2 Friction clutch
Advanced Vehicle Technology
To my long-suffering wife, who has provided support and understanding throughout the preparation
of this book.
Vehicle Technology
Second edition
Heinz Heisler MSc., BSc., F.I.M.I., M.S.O.E., M.I.R.T.E., M.C.I.T., M.I.L.T.
Formerly Principal Lecturer and Head of Transport Studies,
College of North West London, Willesden Centre, London, UK
An imprint of Elsevier Science
Linacre House, Jordan Hill, Oxford OX2 8DP
225 Wildwood Avenue, Woburn, MA 01801-2041
First published by Edward Arnold 1989
Reprinted by Reed Educational and Professional Publishing Ltd 2001
Second edition 2002
Copyright # 1989, 2002 Heinz Heisler. All rights reserved
The right of Heinz Heisler to be identified as the author of this work has been
asserted in accordance with the Copyright, Designs and Patents Act 1988
No part of this publication may be reproduced in any material form (including
photocopying or storing in any medium by electronic means and whether
or not transiently or incidentally to some other use of this publication) without
the written permission of the copyright holder except in accordance with the
provisions of the Copyright, Designs and Patents Act 1988 or under the terms of
a license issued by the Copyright Licensing Agency Ltd, 90 Tottenham Court Road,
London, England W1T 4LP. Applications for the copyright holder's written
permission to reproduce any part of this publication should be addressed
to the publishers
Whilst the advice and information in this book are believed to be true and
accurate at the date of going to press, neither the authors nor the publisher
can accept any legal responsibility or liability for any
errors or omissions that may be made.
Library of Congress Cataloguing in Publication Data
A catalogue record for this book is available from the Library of Congress
ISBN 0 7506 5131 8
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visit our website at www.bh.com
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Printed and bound in Great Britain
1 Vehicle structure
1.1 Integral body construction
Engine, transmission and body structures
1.3 Fifth wheel coupling assembly
1.4 Trailer and caravan drawbar couplings
1.5 Semi-trailer landing gear
1.6 Automatic chassis lubrication system
2 Friction clutch
2.1 Clutch fundamentals
2.2 Angular driven plate cushioning
and torsional damping
2.3 Clutch friction materials
2.4 Clutch
drive and driven member inspection
2.5 Clutch misalignment
2.6 Pull type diaphragm clutch
2.7 Multiplate diaphragm type clutch
2.8 Lipe rollway twin driven plate clutch
2.9 Spicer twin driven plate
angle spring pull type clutch
2.10 Clutch (upshift) brake
2.11 Multiplate hydraulically operated automatic transmission
2.12 Semicentrifugal clutch
2.13 Fully automatic centrifugal clutch
2.14 Clutch pedal actuating mechanisms
2.15 Composite flywheel and integral single plate
diaphragm clutch
3 Manual gearboxes and overdrives
3.1 The necessity for a gearbox
3.2 Five speed and
reverse synchromesh gearboxes
3.3 Gear synchronization and engagement
3.4 Remote controlled gear
selection and engagement m
3.5 Splitter and range change gearboxes
3.6 Transfer box power take-off
3.7 Overdrive considerations
3.8 Setting gear ratios
4 Hydrokinetic fluid couplings
and torque converters
4.1 Hydrokinetic fluid couplings
4.2 Hydrokinetic fluid coupling efficiency
and torque capacity
4.3 Fluid friction coupling
4.4 Hydrokinetic
three element torque converter
4.5 Torque
converter performance terminology
4.6 Overrun clutches
4.7 Three stage hydrokinetic converter
Polyphase hydrokinetic torque converter
4.9 Torque converter with lock-up and gear change
friction clutches
5 Semi....................................................
and fully automatic transmission
5.1 Automatic transmission consideration
5.2 Four speed and reverse longitudinally mounted automatic transmission
mechanical power flow
5.3 The fundamentals
of a hydraulic control system
5.4 Basic principle of a .......................................
hydraulically controlled gearshift
Basic four speed hydraulic control system
5.6 Three speed and reverse transaxle automatic transmission mechanical
power flow
5.7 Hydraulic
gear selection control components
Hydraulic gear selection control operation
5.9 The continuously variable
belt and pulley transmission
5.10 Five speed automatic transmission with electronic-hydraulic
............ control
5.11 Semi-automatic (manual gear change two pedal control) transmission
............................ system
6 Transmission bearings and
constant velocity joints
6.1 Rolling contact bearings
6.2 The need for constant velocity joints
7 Final drive transmission
Crownwheel and pinion axle adjustments
7.2 Differential locks
7.3 Skid reducing differentials
7.4 Double reduction axles
7.5 Two speed axles
7.6 The third (central) differential
7.7 Four wheel drive arrangements
7.8 Electro-hydralic limited slip differential
7.9 Tyre grip when braking and accelerating with good and poor road
7.10 Traction control system
.......................8 Tyres
8.1 Tractive and braking properties of tyres
8.2 Tyre materials
8.3 Tyre tread design
8.4 Cornering properties of tyres
8.5 Vehicle steady state directional stability
8.6 Tyre marking identification
8.7 Wheel balancing
9 Steering
9.1 Steering gearbox fundamental design
9.2 The need for power assisted steering
9.3 Steering linkage ball and socket joints
9.4 Steering geometry and wheel alignment
9.5 Variable-ratio rack and pinion
9.6 Speed sensitive rack and pinion
power assisted steering
9.7 Rack and pinion
electric power assisted steering
10 Suspension
10.1 Suspension geometry
10.2 Suspension roll centres
10.3 Body roll stability analysis
10.4 Anti-roll bars and roll stiffness
10.5 rubber spring bump or limiting stops
10.6 Axle location
10.7 Rear suspension arrangements
10.8 Suspension design consideration
10.9 Hydrogen suspension
10.10 Hydropneumatic automatic height
correction suspension
Commercial vehicle axle beam location
10.12 Variable rate leaf suspension springs.........................................................
10.13 Tandem and tri-axle bogies
10.14 Rubber spring suspension
Air suspensions for commercial vehicles
Lift axle tandem or tri-axle suspension
10.17 Active suspension
10.18 Electronic controlled pneumatic (air) suspension for on and off...road use
11 Brake system
11.1 Braking fun
11.2 Brake shoe and pad fundamentals
11.3 Brake shoe expanders and adjusters
11.4 Disc brake pad support arrangements
11.5 Dual- or split-line braking systems
11.6 Apportional braking
11.7 Antilocking brake system (ABS)
11.8 Brake servos
11.9 Pneumatic operated disk brakes
(for trucks and trailers)
12 Air operated power brake equipment and .................
vehicle retarders
12.1 Introductions to air powered brakes
12.2 Air operated power brake systems
12.3 Air operated power brake equipment
12.4 Vehicle retarders
12.5 Electronic-pneumatic brakes
13 Vehicle refrigeration
13.1 Refrigeration terms
13.2 Principles of a vapour-compression cycle.....................
refrigeration system
13.3 Refrigeration system components
13.4 Vapour-compression cycle refrigeration system with reverse cycle
14 Vehicle body aerodynamics
14.1 Viscous air flow fundamentals
14.2 Aerodynamic drag
14.3 Aerodynamic lift
14.4 Car body drag reduction
14.5 Aerodynamic lift control
14.6 Afterbody drag
14.7 Commercial ...........................................
vehicle aeordynamic fundamentals
14.8 Commercial
vehicle drag reducing devices
.................... Index
Vehicle Structure
1.1 Integral body construction
The integral or unitary body structure of a car can
be considered to be made in the form of three box
compartments; the middle and largest compartment stretching between the front and rear road
wheel axles provides the passenger space, the
extended front box built over and ahead of the front
road wheels enclosing the engine and transmission
units and the rear box behind the back axle
providing boot space for luggage.
Fig. 1.1 (a and b)
These box compartments are constructed in the
form of a framework of ties (tensile) and struts
(compressive), pieces (Fig. 1.1(a & b)) made from
rolled sheet steel pressed into various shapes such
as rectangular, triangular, trapezium, top-hat or a
combination of these to form closed box thin gauge
sections. These sections are designed to resist direct
tensile and compressive or bending and torsional
loads, depending upon the positioning of the members within the structure.
Structural tensile and compressive loading of car body
1.1.1 Description and function of body
components (Fig. 1.2)
The major individual components comprising the
body shell will now be described separately under
the following subheadings:
Cantrails (Fig. 1.2(4)) Cantrails are the horizontal members which interconnect the top ends of the
vertical A and BC or BC and D door pillars (posts).
These rails form the side members which make up
the rectangular roof framework and as such are
subjected to compressive loads. Therefore, they
are formed in various box-sections which offer the
greatest compressive resistance with the minimum
of weight and blend in with the roofing. A drip rail
(Fig. 1.2(4)) is positioned in between the overlapping roof panel and the cantrails, the joins being
secured by spot welds.
Window and door pillars
Windscreen and rear window rails
Roof structure
Upper quarter panel or window
Floor seat and boot pans
Central tunnel
Front longitudinals
Front valance
Rear valance
Toe board
Heel board
Roof structure (Fig. 1.2) The roof is constructed
basically from four channel sections which form
the outer rim of the slightly dished roof panel.
The rectangular outer roof frame acts as the compressive load bearing members. Torsional rigidity
to resist twist is maximized by welding the four
corners of the channel-sections together. The slight
curvature of the roof panel stiffens it, thus preventing winkling and the collapse of the unsupported
centre region of the roof panel. With large cars,
additional cross-rail members may be used to
provide more roof support and to prevent the roof
crushing in should the car roll over.
Window and door pillars (Fig. 1.2(3, 5, 6, and 8))
Windowscreen and door pillars are identified by a
letter coding; the front windscreen to door pillars
are referred to as A post, the centre side door pillars
as BC post and the rear door to quarter panel as
D post. These are illustrated in Fig. 1.2.
These pillars form the part of the body structure
which supports the roof. The short form A pillar and
rear D pillar enclose the windscreen and quarter
windows and provide the glazing side channels,
whilst the centre BC pillar extends the full height of
the passenger compartment from roof to floor and
supports the rear side door hinges. The front and
rear pillars act as struts (compressive members)
which transfer a proportion of the bending effect,
due to underbody sag of the wheelbase, to each end
of the cantrails which thereby become reactive
struts, opposing horizontal bending of the passenger compartment at floor level. The central BC
pillar however acts as ties (tensile members), transferring some degree of support from the mid-span of
the cantrails to the floor structure.
Upper quarter panel or window (Fig. 1.2(6)) This
is the vertical side panel or window which occupies
the space between the rear side door and the rear
window. Originally the quarter panel formed an
important part of the roof support, but improved
pillar design and the desire to maximize visibility
has either replaced them with quarter windows or
reduced their width, and in some car models they
have been completely eliminated.
Floor seat and boot pans (Fig. 1.3) These constitute the pressed rolled steel sheeting shape to
enclose the bottom of both the passenger and luggage compartments. The horizontal spread-out
pressing between the bulkhead and the heel board
is called the floor pan, whilst the raised platform
over the rear suspension and wheel arches is known
as the seat or arch pan. This in turn joins onto a
lower steel pressing which supports luggage and is
referred to as the boot pan.
To increase the local stiffness of these platform
panels or pans and their resistance to transmitted
vibrations such as drumming and droning, many
narrow channels are swaged (pressed) into the steel
sheet, because a sectional end-view would show a
Windscreen and rear window rails (Fig. 1.2(2))
These box-section rails span the front window
pillars and rear pillars or quarter panels depending
upon design, so that they contribute to the resistance opposing transverse sag between the wheel
track by acting as compressive members. The
other function is to support the front and rear
ends of the roof panel. The undersides of the rails
also include the glazing channels.
Fig. 1.2 Load bearing body box-section members
semi-corrugated profile (or ribs). These channels
provide rows of shallow walls which are both bent
and stretched perpendicular to the original flat
sheet. In turn they are spaced and held together
by the semicircular drawn out channel bottoms.
Provided these swages are designed to lay the
correct way and are not too long, and the metal is
not excessively stretched, they will raise the rigidity
Fig. 1.3 (a±c) Platform chassis
of these panels so that they are equivalent to a sheet
which may be several times thicker.
spans between the rear end of the valance, where it
meets the bulkhead, and the door pillar and wing.
The lower edge of the scuttle will merge with the
floor pan so that in some cases it may form part of
the toe board on the passenger compartment side.
Usually these panels form inclined sides to the bulkhead, and with the horizontal ledge which spans the
full width of the bulkhead, brace the bulkhead wall
so that it offers increased rigidity to the structure.
The combined bulkhead dash panel and scuttle will
thereby have both upright and torsional rigidity.
Central tunnel (Fig. 1.3(a and b)) This is the
curved or rectangular hump positioned longitudinally along the middle of the floor pan. Originally it
was a necessary evil to provide transmission space
for the gearbox and propeller shaft for rear wheel
drive, front-mounted engine cars, but since the
chassis has been replaced by the integral boxsection shell, it has been retained with front wheel
drive, front-mounted engines as it contributes
considerably to the bending rigidity of the floor
structure. Its secondary function is now to house
the exhaust pipe system and the hand brake cable
Front longitudinals (Figs 1.2(10) and 1.3(a and b))
These members are usually upswept box-section
members, extending parallel and forward from the
bulkhead at floor level. Their purpose is to withstand the engine mount reaction and to support the
front suspension or subframe. A common feature
of these members is their ability to support vertical
loads in conjunction with the valances. However, in
the event of a head-on collision, they are designed
to collapse and crumble within the engine compartment so that the passenger shell is safeguarded and
is not pushed rearwards by any great extent.
Sills (Figs 1.2(9) and 1.3(a, b and c)) These members
form the lower horizontal sides of the car body
which spans between the front and rear road-wheel
wings or arches. To prevent body sag between the
wheelbase of the car and lateral bending of the
structure, the outer edges of the floor pan are given
support by the side sills. These sills are made in the
form of either single or double box-sections
(Fig. 1.2(9)). To resist the heavier vertical bending
loads they are of relatively deep section.
Open-top cars, such as convertibles, which do not
receive structural support from the roof members,
usually have extra deep sills to compensate for the
increased burden imposed on the underframe.
Front valance (Figs 1.2 and 1.3(a and b)) These
panels project upwards from the front longitudinal
members and at the rear join onto the wall of the
bulkhead. The purpose of these panels is to transfer
the upward reaction of the longitudinal members
which support the front suspension to the bulkhead.
Simultaneously, the longitudinals are prevented
from bending sideways because the valance panels
are shaped to slope up and outwards towards the
top. The panelling is usually bent over near the
edges to form a horizontal flanged upper, thus
presenting considerable lateral resistance. Furthermore, the valances are sometimes stepped and
wrapped around towards the rear where they meet
and are joined to the bulkhead so that additional
lengthwise and transverse stiffness is obtained.
If coil spring suspension is incorporated, the
valance forms part of a semi-circular tower which
houses and provides the load reaction of the spring
so that the merging of these shapes compounds the
rigidity for both horizontal lengthwise and lateral
bending of the forward engine and transmission
compartment body structure. Where necessary,
double layers of sheet are used in parts of the spring
housing and at the rear of the valance where they
are attached to the bulkhead to relieve some of the
concentrated loads.
Bulkhead (Figs 1.2(1) and 1.3(a and b)) This is the
upright partition separating the passenger and
engine compartments. Its upper half may form
part of the dash panel which was originally used to
display the driver's instruments. Some body manufacturers refer to the whole partition between engine
and passenger compartments as the dash panel. If
there is a double partition, the panel next to the
engine is generally known as the bulkhead and that
on the passenger side the dash board or panel. The
scuttle and valance on each side are usually joined
onto the box-section of the bulkhead. This braces
the vertical structure to withstand torsional distortion and to provide platform bending resistance
support. Sometimes a bulkhead is constructed
between the rear wheel arches or towers to reinforce
the seat pan over the rear axle (Fig. 1.3(c)).
Scuttle (Fig. 1.3(a and b)) This can be considered
as the panel formed under the front wings which
Rear valance (Fig. 1.2(7)) This is generally considered as part of the box-section, forming the front
half of the rear wheel arch frame and the panel
immediately behind which merges with the heel
board and seatpan panels. These side inner-side
panels position the edges of the seat pan to its
designed side profile and thus stiffen the underfloor
structure above the rear axle and suspension. When
rear independent coil spring suspension is adopted,
the valance or wheel arch extends upwards to form
a spring tower housing and, because it forms a
semi-vertical structure, greatly contributes to the
stiffness of the underbody shell between the floor
and boot pans.
Torsional rigidity of the platform is usually
derived at the front by the bulkhead, dash pan
and scuttle (Fig. 1.3(a and b)) at the rear by the
heel board, seat pan, wheel arches (Fig. 1.3(a, b and
c)), and if independent rear suspension is adopted,
by the coil spring towers (Fig. 1.3(a and c)).
Between the wheelbase, the floor pan is normally
provided with box-section cross-members to stiffen
and prevent the platform sagging where the
passenger seats are positioned.
1.1.3 Stiffening of platform chassis
(Figs 1.4 and 1.5)
To appreciate the stresses imposed on and the
resisting stiffness offered by sheet steel when it is
subjected to bending, a small segment of a beam
greatly magnified will now be considered (Fig.
1.4(a)). As the beam deforms, the top fibres contract and the bottom fibres elongate. The neutral
plane or axis of the beam is defined as the plane
whose length remains unchanged during deformation and is normally situated in the centre of a
uniform section (Fig. 1.4(a and b)).
The stress distribution from top to bottom within
the beam varies from zero along the neutral axis
(NA), where there is no change in the length of the
fibres, to a maximum compressive stress on the outer
top layer and a maximum tensile stress on the outer
bottom layer, the distortion of the fibres being
greatest at their extremes as shown in Fig. 1.4(b).
It has been found that bending resistance
increases roughly with the cube of its distance
from the neutral axis (Fig. 1.5(a)). Therefore, bending resistance of a given section can be greatly
improved for a given weight of metal by taking
metal away from the neutral axis where the metal
fibres do not contribute very much to resisting
distortion and placing it as far out as possible
where the distortion is greatest. Bending resistance
may be improved by using longitudinal or crossmember deep box-sections (Fig. 1.5(b)) and tunnel
sections (Fig. 1.5(c)) to restrain the platform chassis from buckling and to stiffen the flat horizontal
floor seat and boot pans. So that vibration and
drumming may be reduced, many swaged ribs are
pressed into these sheets (Fig. 1.5(d)).
Toe board The toe board is considered to form
the lower regions of the scuttle and dash panel near
where they merge with the floor pan. It is this
panelling on the passenger compartment side
where occupants can place their feet when the car
is rapidly retarded.
Heel board (Fig. 1.3(b and c)) The heel board is
the upright, but normally shallow, panel spanning
beneath and across the front of the rear seats. Its
purpose is to provide leg height for the passengers
and to form a raised step for the seat pan so that
the rear axle has sufficient relative movement
1.1.2 Platform chassis (Fig. 1.3(a±c))
Most modern car bodies are designed to obtain
their rigidity mainly from the platform chassis and
to rely less on the upper framework of window
and door pillars, quarter panels, windscreen rails
and contrails which are becoming progressively
slender as the desire for better visibility is encouraged.
The majority of the lengthwise (wheelbase) bending stiffness to resist sagging is derived from both
the central tunnel and the side sill box-sections
(Fig. 1.3(a and b)). If further strengthening is
necessary, longitudinal box-section members may
be positioned parallel to, but slightly inwards from,
the sills (Fig. 1.3(c)). These lengthwise members
may span only part of the wheelbase, or the full
length, which is greatly influenced by the design of
road wheel suspension chosen for the car, the depth
of both central tunnel and side sills, which are built
into the platform, and if there are subframes
attached fore and aft of the wheelbase (Fig. 1.6
(a and b)).
1.1.4 Body subframes (Fig. 1.6)
Front or rear subframes may be provided to brace
the longitudinal side members so that independent
suspension on each side of the car receives adequate
support for the lower transverse swing arms (wishbone members). Subframes restrain the two halves
of the suspension from splaying outwards or the
Fig. 1.4 Stress and strain imposed on beam when subjected to bending
longitudinal side members from lozenging as alternative road wheels experience impacts when travelling over the irregularities of a normal road surface.
It is usual to make the top side of the subframe
the cradle for the engine or engine and transmission
mounting points so that the main body structure
itself does not have to be reinforced. This particularly applies where the engine, gearbox and final
drive form an integral unit because any torque
reaction at the mounting points will be transferred
to the subframe and will multiply in proportion to
the overall gear reduction. This may be approximately four times as great as that for the front
mounted engine with rear wheel drive and will
become prominent in the lower gears.
One advantage claimed by using separate subframes attached to the body underframe through
the media of rubber mounts is that transmitted
vibrations and noise originating from the tyres
and road are isolated from the main body shell
and therefore do not damage the body structure
and are not relayed to the occupants sitting
Cars which have longitudinally positioned
engines mounted in the front driven by the rear
wheels commonly adopt beam cross-member
subframes at the front to stiffen and support the
hinged transverse suspension arms (Fig. 1.6(a)).
Saloon cars employing independent rear suspension sometimes prefer to use a similar subframe at
the rear which provides the pivot points for the
semi-trailing arms because this type of suspension
requires greater support than most other arrangements (Fig. 1.6(a)).
Fig. 1.5 Bending resistance for various sheet sections
When the engine, gearbox and final drive are
combined into a single unit, as with the front longitudinally positioned engine driving the front wheels
where there is a large weight concentration, a subframe gives extra support to the body longitudinal
side members by utilising a horseshoe shaped frame
(Fig. 1.6(b)). This layout provides a platform for
the entire mounting points for both the swing arm
and anti-roll bar which between them make up the
lower part of the suspension.
Fig. 1.6 (a±c) Body subframe and underfloor structure
Front wheel drive transversely positioned
engines with their large mounting point reactions
often use a rectangular subframe to spread out
both the power and transmission unit's weight
and their dynamic reaction forces (Fig. 1.6(c)).
This configuration provides substantial torsional
rigidity between both halves of the independent
suspension without relying too much on the main
body structure for support.
modifies the magnitude of frequencies of the
vibrations so that they are less audible to the
The installation of acoustic materials cannot
completely eliminate boom, drumming, droning
and other noises caused by resonance, but merely
reduces the overall noise level.
Insulation Because engines are generally mounted
close to the passenger compartment of cars or the
cabs of trucks, effective insulation is important. In
this case, the function of the material is to reduce
the magnitude of vibrations transmitted through
the panel and floor walls. To reduce the transmission of noise, a thin steel body panel should be
combined with a flexible material of large mass,
based on PVC, bitumen or mineral wool. If the
insulation material is held some distance from the
structural panel, the transmissibility at frequencies
above 400 Hz is further reduced. For this type of
application the loaded PVC material is bonded to a
spacing layer of polyurethane foam or felt, usually
about 7 mm thick. At frequencies below 400 Hz, the
use of thicker spacing layers or heavier materials
can also improve insulation.
Soundproofing the interior of the passenger
compartment (Fig. 1.7)
Interior noise originating outside the passenger
compartment can be greatly reduced by applying
layers of materials having suitable acoustic properties over floor, seat and boot pans, central tunnel,
bulkhead, dash panel, toeboard, side panels, inside
of doors, and the underside of both roof and
bonnet etc. (Fig. 1.7).
Acoustic materials are generally designed for one
of three functions:
a) Insulation from noise Ð This may be created by
forming a non-conducting noise barrier
between the source of the noises (which may
come from the engine, transmission, suspension
tyres etc.) and the passenger compartment.
b) Absorption of vibrations Ð This is the transference of excited vibrations in the body shell to
a media which will dissipate their resultant
energies and so eliminate or at least greatly
reduce the noise.
c) Damping of vibrations Ð When certain vibrations cannot be eliminated, they may be exposed
to some form of material which in some way
Absorption For absorption, urethane foam or
lightweight bonded fibre materials can be used.
In some cases a vinyl sheet is bonded to the foam
to form a roof lining. The required thickness of the
absorbent material is determined by the frequencies
involved. The minimum useful thickness of
polyurethane foam is 13 mm which is effective
with vibration frequencies above 1000 Hz.
Damping To damp resonance, pads are bonded
to certain panels of many cars and truck cabs. They
are particularly suitable for external panels whose
resonance cannot be eliminated by structural
alterations. Bituminous sheets designed for this
purpose are fused to the panels when the paint is
baked on the car. Where extremely high damping
or light weight is necessary, a PVC base material,
which has three times the damping capacity of
bituminous pads, can be used but this material is
rather difficult to attach to the panelling.
1.1.5 Collision safety (Fig. 1.8)
Car safety may broadly be divided into two kinds:
Firstly the active safety, which is concerned with
the car's road-holding stability while being driven,
steered or braked and secondly the passive safety,
Fig. 1.7 Car body sound generation and its dissipation
collision, but overall alignment may also be necessary if the vehicle's steering and ride characteristics
do not respond to the expected standard of a similar vehicle when being driven.
Structural misalignment may be caused by all
sorts of reasons, for example, if the vehicle has
been continuously driven over rough ground at
high speed, hitting an obstacle in the road, mounting steep pavements or kerbs, sliding off the road
into a ditch or receiving a glancing blow from some
other vehicle or obstacle etc. Suspicion that something is wrong with the body or chassis alignment is
focused if there is excessively uneven or high tyre
wear, the vehicle tends to wander or pull over to
one side and yet the track and suspension geometry
appears to be correct.
Alignment checks should be made on a level,
clear floor with the vehicle's tyres correctly inflated
to normal pressure. A plumb bob is required in the
form of a stubby cylindrical bar conical shaped at
one end, the other end being attached to a length of
thin cord. Datum reference points are chosen such
as the centre of a spring eye on the chassis mounting point, transverse wishbone and trailing arm
pivot centres, which are attachment points to the
underframe or chassis, and body cross-member to
side-member attachment centres and subframe
bolt-on points (Fig. 1.9).
Initially the cord with the plumb bob hanging
from its end is lowered from the centre of each
reference point to the floor and the plumb bob contact point with the ground is marked with a chalked
cross. Transverse and diagonal lines between reference points can be made by chalking the full length
of a piece of cord, holding it taut between reference
centres on the floor and getting somebody to pluck
the centre of the line so that it rebounds and leaves
a chalked line on the floor.
A reference longitudinal centre line may be made
with a strip of wood baton of length just greater
than the width between adjacent reference marks
on the floor. A nail is punched through one end
and this is placed over one of the reference marks.
A piece of chalk is then held at the tip of the free
end and the whole wood strip is rotated about
the nailed end. The chalk will then scribe an arc
between adjacent reference points. This is repeated
from the other side. At the points where these two
arcs intersect a straight line is made with a plucked,
chalked cord running down the middle of the vehicle. This procedure should be followed at each end
of the vehicle as shown in Fig. 1.9.
Once all the reference points and transverse and
diagonal joining lines have been drawn on the
Fig. 1.8 Collision body safety
which depends upon body style and design structure to protect the occupants of the car from serious
injury in the event of a collision.
Car bodies can be considered to be made in three
parts (Fig. 1.8); a central cell for the passengers
of the welded bodywork integral with a rigid
platform, acting as a floor pan, and chassis with
various box-section cross- and side-members. This
type of structure provides a reinforced rigid crushproof construction to resist deformation on impact
and to give the interior a high degree of protection.
The extension of the engine and boot compartments at the front and rear of the central passenger
cell are designed to form zones which collapse and
crumble progressively over the short duration of a
collision impact. Therefore, the kinetic energy due
to the car's initial speed will be absorbed fore and
aft primarily by strain and plastic energy within the
crumble zones with very little impact energy actually being dissipated by the central body cell.
1.1.6 Body and chassis alignment checks
(Fig. 1.9)
Body and chassis alignment checks will be necessary if the vehicle has been involved in a major
Table 1.1 Summary of function and application of
soundproofing materials
Acoustic materials
Loaded PVC,
bitumen, with or
without foam or
fibres base,
mineral wool
Floor, bulkhead
dash panel
Bitumen or
Doors, side
underside of roof
Polyurethane foam,
mineral wool, or
bonded fibres
Side panels,
underside of
roof, engine
Fig. 1.9 Body underframe alignment checks
floor, a rule or tape is used to measure the distances
between centres both transversely and diagonally.
These values are then chalked along their respective
lines. Misalignment or error is observed when a
pair of transverse or diagonal dimensions differ
and further investigation will thus be necessary.
Note that transverse and longitudinal dimensions are normally available from the manufacturer's manual and differences between paired
diagonals indicates lozenging of the framework
due to some form of abnormal impact which has
previously occurred.
Both the variations of inertia and gas pressure
forces generate three kinds of vibrations which are
transferred to the cylinder block:
1.2 Engine, transmission and body structure
1 To prevent the fatigue failure of the engine and
gearbox support points which would occur if
they were rigidly attached to the chassis or
body structure.
2 To reduce the amplitude of any engine vibration
which is being transmitted to the body structure.
3 To reduce noise amplification which would occur
if engine vibration were allowed to be transferred
directly to the body structure.
1 Vertical and/or horizontal shake and rock
2 Fluctuating torque reaction
3 Torsional oscillation of the crankshaft
1.2.2 Reasons for flexible mountings
It is the objective of flexible mounting design to
cope with the many requirements, some having
conflicting constraints on each other. A list of the
duties of these mounts is as follows:
1.2.1 Inherent engine vibrations
The vibrations originating within the engine are
caused by both the cyclic acceleration of the reciprocating components and the rapidly changing
cylinder gas pressure which occurs throughout
each cycle of operation.
4 To reduce human discomfort and fatigue by
partially isolating the engine vibrations from
the body by means of an elastic media.
5 To accommodate engine block misalignment
and to reduce residual stresses imposed on the
engine block and mounting brackets due to
chassis or body frame distortion.
6 To prevent road wheel shocks when driving
over rough ground imparting excessive rebound
movement to the engine.
7 To prevent large engine to body relative movement due to torque reaction forces, particularly
in low gear, which would cause excessive misalignment and strain on such components as
the exhaust pipe and silencer system.
8 To restrict engine movement in the fore and aft
direction of the vehicle due to the inertia of the
engine acting in opposition to the accelerating
and braking forces.
1.2.3 Rubber flexible mountings (Figs 1.10, 1.11
and 1.12)
A rectangular block bonded between two metal
plates may be loaded in compression by squeezing
the plates together or by applying parallel but
opposing forces to each metal plate. On compression, the rubber tends to bulge out centrally from
the sides and in shear to form a parallelogram
(Fig. 1.10(a)).
To increase the compressive stiffness of the
rubber without greatly altering the shear stiffness,
an interleaf spacer plate may be bonded in between
the top and bottom plate (Fig. 1.10(b)). This interleaf plate prevents the internal outward collapse of
the rubber, shown by the large bulge around the
sides of the block, when no support is provided,
whereas with the interleaf a pair of much smaller
bulges are observed.
Fig. 1.11
Fig. 1.10 (a and b)
Modes of loading rubber blocks
When two rubber blocks are inclined to each other
to form a `V' mounting, see Fig. 1.11, the rubber will
be loaded in both compression and shear shown by
the triangle of forces. The magnitude of compressive
force will be given by Wc and the much smaller shear
force by WS. This produces a resultant reaction force
WR. The larger the wedge angle , the greater the
proportion of compressive load relative to the shear
load the rubber block absorbs.
The distorted rubber provides support under
light vertical static loads approximately equal in
both compression and shear modes, but with
heavier loads the proportion of compressive stiffness
`V' rubber block mounting
These modes of movement may be summarized
as follows:
Linear motions
1 Horizontal
2 Horizontal lateral
3 Vertical
Rotational motions
4 Roll
5 Pitch
6 Yaw
1.2.6 Positioning of engine and gearbox
mountings (Fig. 1.15)
If the mountings are placed underneath the combined engine and gearbox unit, the centre of gravity
is well above the supports so that a lateral (side)
force acting through its centre of gravity, such as
experienced when driving round a corner, will cause
the mass to roll (Fig. 1.15(a)). This condition is
undesirable and can be avoided by placing the
mounts on brackets so that they are in the
same plane as the centre of gravity (Fig. 1.15(b)).
Thus the mounts provide flexible opposition to
any side force which might exist without creating a
roll couple. This is known as a decoupled condition.
An alternative method of making the natural
modes of oscillation independent or uncoupled is
achieved by arranging the supports in an inclined
`V' position (Fig. 1.15(c)). Ideally the aim is to
make the compressive axes of the mountings meet
at the centre of gravity, but due to the weight of the
power unit distorting the rubber springing the
inter-section lines would meet slightly below this
point. Therefore, the mountings are tilted so that
the compressive axes converge at some focal point
above the centre of gravity so that the actual lines
of action of the mountings, that is, the direction
of the resultant forces they exert, converge on the
centre of gravity (Fig. 1.15(d)).
The compressive stiffness of the inclined mounts
can be increased by inserting interleafs between
the rubber blocks and, as can be seen in
Fig. 1.15(e), the line of action of the mounts converges at a lower point than mounts which do not
have interleaf support.
Engine and gearbox mounting supports are
normally of the three or four point configuration.
Petrol engines generally adopt the three point
support layout which has two forward mounts
(Fig. 1.13(a and c)), one inclined on either side of
the engine so that their line of action converges on
the principal axis, while the rear mount is supported
centrally at the rear of the gearbox in approximately
the same plane as the principal axis. Large diesel
engines tend to prefer the four point support
Fig. 1.12 Load±deflection curves for rubber block
to that of shear stiffness increases at a much faster
rate (Fig. 1.12). It should also be observed that the
combined compressive and shear loading of the
rubber increases in direct proportion to the static
deflection and hence produces a straight line graph.
1.2.4 Axis of oscillation (Fig. 1.13)
The engine and gearbox must be suspended so that
it permits the greatest degree of freedom when
oscillating around an imaginary centre of rotation
known as the principal axis. This principal axis
produces the least resistance to engine and gearbox
sway due to their masses being uniformly distributed about this axis. The engine can be considered
to oscillate around an axis which passes through
the centre of gravity of both the engine and gearbox
(Figs 1.13(a, b and c)). This normally produces an
axis of oscillation inclined at about 10±20 to the
crankshaft axis. To obtain the greatest degree of
freedom, the mounts must be arranged so that they
offer the least resistance to shear within the rubber
1.2.5 Six modes of freedom of a suspended body
(Fig. 1.14)
If the movement of a flexible mounted engine is
completely unrestricted it may have six modes of
vibration. Any motion may be resolved into three
linear movements parallel to the axes which pass
through the centre of gravity of the engine but at
right angles to each other and three rotations about
these axes (Fig. 1.14).
Fig. 1.13
Axis of oscillation and the positioning of the power unit flexible mounts
arrangement where there are two mounts either side
of the engine (Fig. 1.13(b)). The two front mounts
are inclined so that their lines of action pass through
the principal axis, but the rear mounts which are
located either side of the clutch bell housing are not
inclined since they are already at principal axis level.
down at a uniform rate. The amplitude of this cyclic
movement will progressively decrease and the number of oscillations per minute of the rubber mounting
is known as its natural frequency of vibration.
There is a relationship between the static deflection imposed on the rubber mount springing by the
suspended mass and the rubber's natural frequency
of vibration, which may be given by
1.2.7 Engine and transmission vibrations
n0 ˆ p
Natural frequency of vibration (Fig. 1.16) A sprung
body when deflected and released will bounce up and
Fig. 1.14 Six modes of freedom for a suspended block
the engine out of balance forces and the fluctuating
cylinder gas pressure and the natural frequency of
oscillation of the elastic rubber support mounting,
i.e. resonance occurs when
n0 = natural frequency of vibration
x = static deflection of the rubber (m)
This relationship between static deflection and
natural frequency may be seen in Fig. 1.16.
Resonance Resonance is the unwanted synchronization of the disturbing force frequency imposed by
n = disturbing frequency
n0 = natural frequency
Transmissibility (Fig. 1.17) When the designer
selects the type of flexible mounting the Theory of
Transmissibility can be used to estimate critical
resonance conditions so that they can be either
prevented or at least avoided.
Transmissibility (T) may be defined as the ratio
of the transmitted force or amplitude which passes
through the rubber mount to the chassis to that of
the externally imposed force or amplitude generated
by the engine:
transmitted force or amplitude
imposed disturbing force or
This relationship between transmissibility and
the ratio of disturbing frequency and natural
frequency may be seen in Fig. 1.17.
Fig. 1.16 Relationship of static deflection and natural
Fig. 1.15 (a±e)
Coupled and uncoupled mounting points
The transmissibility to frequency ratio graph
(Fig. 1.17) can be considered in three parts as follows:
rubber mountings is greater than 11¤2 and the transmissibility is less than one. Under these conditions
off-peak partial resonance vibrations passing to the
body structure will be minimized.
Range (I) This is the resonance range and should be
avoided. It occurs when the disturbing frequency
is very near to the natural frequency. If steel mounts
are used, a critical vibration at resonance would go
to infinity, but natural rubber limits the transmissibility to around 10. If Butyl synthetic rubber is
adopted, its damping properties reduce the peak
transmissibility to about 21¤2. Unfortunately, high
damping rubber compounds such as Butyl rubber
are temperature sensitive to both damping and
dynamic stiffness so that during cold weather a
noticeably harsher suspension of the engine results.
Damping of the engine suspension mounting is
necessary to reduce the excessive movement of a
flexible mounting when passing through resonance,
but at speeds above resonance more vibration is
transmitted to the chassis or body structure than
would occur if no damping was provided.
Range (III) This is known as the shock reduction
range and only occurs when the disturbing
frequency is lower than the natural frequency.
Generally it is only experienced with very soft
rubber mounts and when the engine is initially
cranked for starting purposes and so quickly passes
through this frequency ratio region.
Example An engine oscillates vertically on its
flexible rubber mountings with a frequency of 800
vibrations per minute (vpm). With the information
provided answer the following questions:
a) From the static deflection±frequency graph,
Fig. 1.16, or by formula, determine the natural frequency of vibration when the static deflection of
the engine is 2 mm and then find the disturbing to
natural frequency ratio. Comment on these results.
b) If the disturbing to natural frequency ratio is
increased to 2.5 determine the natural frequency
Range (II) This is the recommended working
range where the ratio of the disturbing frequency
to that of the natural frequency of vibration of the
Fig. 1.17 Relationship of transmissibility and
the ratio of disturbing and natural frequencies
for natural rubber, Butyl rubber and steel
of vibration and the new static deflection of the
engine. Comment of these conditions.
1.2.9 Subframe to body mountings
(Figs 1.6 and 1.19)
One of many problems with integral body design is
the prevention of vibrations induced by the engine,
transmission and road wheels from being transmitted
through the structure. Some manufacturers adopt a
subframe (Fig. 1.6(a, b and c)) attached by resilient
mountings (Fig. 1.19(a and b)) to the body to which
the suspension assemblies, and in some instances the
engine and transmission, are attached. The mass
of the subframes alone helps to damp vibrations.
It also simplifies production on the assembly line,
and facilitates subsequent overhaul or repairs.
In general, the mountings are positioned so that
they allow strictly limited movement of the
subframe in some directions but provide greater
freedom in others. For instance, too much lateral
freedom of a subframe for a front suspension
assembly would introduce a degree of instability
into the steering, whereas some freedom in vertical
and longitudinal directions would improve the
quality of a ride.
a) n0 ˆ p ˆ p
ˆ 670:84 vib/min
ˆ 1:193
n0 670:84
The ratio 1.193 is very near to the resonance
condition and should be avoided by using softer
ˆ 2:5
ˆ 320 vib/min
Now n0 ˆ p
thus x ˆ
2 30
30 2
; n0 ˆ
1.2.10 Types of rubber flexible mountings
A survey of typical rubber mountings used for
power units, transmissions, cabs and subframes
are described and illustrated as follows:
ˆ 0:008789 m or 8:789 mm
A low natural frequency of 320 vib/min is well
within the insulation range, therefore from either
the deflection±frequency graph or by formula
the corresponding rubber deflection necessary is
8.789 mm when the engine's static weight bears
down on the mounts.
Double shear paired sandwich mounting (Fig.
1.18(a)) Rubber blocks are bonded between the
jaws of a `U' shaped steel plate and a flat interleaf
plate so that a double shear elastic reaction takes
place when the mount is subjected to vertical loading. This type of shear mounting provides a large
degree of flexibility in the upright direction and
thus rotational freedom for the engine unit about
its principal axis. It has been adopted for both
engine and transmission suspension mounting
points for medium-sized diesel engines.
1.2.8 Engine to body/chassis mountings
Engine mountings are normally arranged to
provide a degree of flexibility in the horizontal
longitudinal, horizontal lateral and vertical axis of
rotation. At the same time they must have sufficient stiffness to provide stability under shock
loads which may come from the vehicle travelling
over rough roads. Rubber sprung mountings
suitably positioned fulfil the following functions:
Double inclined wedge mounting (Fig. 1.18(b)) The
inclined wedge angle pushes the bonded rubber
blocks downwards and outwards against the
bent-up sides of the lower steel plate when loaded
in the vertical plane. The rubber blocks are subjected
to both shear and compressive loads and the proportion of compressive to shear load becomes greater
with vertical deflection. This form of mounting is
suitable for single point gearbox supports.
1 Rotational flexibility around the horizontal
longitudinal axis which is necessary to allow the
impulsive inertia and gas pressure components
of the engine torque to be absorbed by rolling of
the engine about the centre of gravity.
2 Rotational flexibility around both the horizontal
lateral and the vertical axis to accommodate any
horizontal and vertical shake and rock caused by
unbalanced reciprocating forces and couples.
Inclined interleaf rectangular sandwich mounting
(Fig. 1.18(c)) These rectangular blocks are
Fig. 1.18 (a±h)
Types of rubber flexible mountings
Fig. 1.18
Fig. 1.18 contd
designed to be used with convergent `V' formation
engine suspension system where the blocks are
inclined on either side of the engine. This configuration enables the rubber to be loaded in both shear
and compression with the majority of engine rotational flexibility being carried out in shear. Vertical
deflection due to body pitch when accelerating or
braking is absorbed mostly in compression. Vertical
elastic stiffness may be increased without greatly
effecting engine roll flexibility by having metal
spacer interleafs bonded into the rubber.
on either side of the power unit's bell housing
at principal axis level may be used. Longitudinal
movement is restricted by the double `V' formed
between the inner and two outer members seen in
a plan view. This `V' and wedge configuration provides a combined shear and compressive strain to
the rubber when there is a relative fore and aft movement between the engine and chassis, in addition to
that created by the vertical loading of the mount.
This mounting's major application is for the rear
mountings forming part of a four point suspension
for heavy diesel engines.
Double inclined wedge with longitudinal control
mounting (Fig. 1.18(d)) Where heavy vertical
loads and large rotational reactions are to be
absorbed, double inclined wedge mounts positioned
Metaxentric bush mounting (Fig. 1.18(e)) When
the bush is in the unloaded state, the steel inner
sleeve is eccentric relative to the outer one so that
there is more rubber on one side of it than on the
other. Precompression is applied to the rubber
expanding the inner sleeve. The bush is set so that
the greatest thickness of rubber is in compression
in the laden condition. A slot is incorporated in
the rubber on either side where the rubber is at its
minimum in such a position as to avoid stressing
any part of it in tension.
When installed, its stiffness in the fore and aft
direction is greater than in the vertical direction, the
ratio being about 2.5 : 1. This type of bush provides
a large amount of vertical deflection with very little
fore and aft movement which makes it suitable for
rear gearbox mounts using three point power unit
suspension and leaf spring eye shackle pin bushes.
distortion within the rubber. Under small deflection conditions the shear and compression is
almost equal, but as the load and thus deflection
increases, the proportion of compression over the
shear loading predominates.
These mounts provide very good lateral stability
without impairing vertical deflection flexibility and
progressive stiffness control. When used for road
wheel axle suspension mountings, they offer good
insulation against road and other noises.
Flanged sleeve bobbin mounting with rebound
control (Fig. 1.19(a and b)) These mountings
have the rubber moulded partially around the outer
flange sleeve and in between this sleeve and an inner
tube. A central bolt attaches the inner tube to the
body structure while the outer member is bolted on
two sides to the subframe.
When loaded in the vertical downward direction,
the rubber between the sleeve and tube walls will be
in shear and the rubber on the outside of the
flanged sleeve will be in compression.
There is very little relative sideway movement
between the flanged sleeve and inner tube due to
rubber distortion. An overload plate limits the downward deflection and rebound is controlled by the
lower plate and the amount and shape of rubber
trapped between it and the underside of the flanged
sleeve. A reduction of rubber between the flanged
sleeve and lower plate (Fig. 1.19(a)) reduces the
rebound, but an increase in depth of rubber increases
rebound (Fig. 1.19(b)). The load deflection characteristics are given for both mounts in Fig. 1.19c.
These mountings are used extensively for body to
subframe and cab to chassis mounting points.
Metacone sleeve mountings (Fig. 1.18(f and g))
These mounts are formed from male and female
conical sleeves, the inner male member being
centrally positioned by rubber occupying the
space between both surfaces (Fig. 1.18(f)). During
vertical vibrational deflection, the rubber between
the sleeves is subjected to a combined shear and
compression which progressively increases the stiffness of the rubber as it moves towards full distortion. The exposed rubber at either end overlaps the
flanged outer sleeve and there is an upper and
lower plate bolted rigidly to the ends of the inner
sleeve. These plates act as both overload (bump)
and rebound stops, so that when the inner member
deflects up or down towards the end of its movement it rapidly stiffens due to the surplus rubber
being squeezed in between. Mounts of this kind are
used where stiffness is needed in the horizontal
direction with comparative freedom of movement
for vertical deflection.
An alternative version of the Metacone mount
uses a solid aluminium central cone with a flanged
pedestal conical outer steel sleeve which can be
bolted directly onto the chassis side member, see
Fig. 1.18(g). An overload plate is clamped between
the inner cone and mount support arm, but no
rebound plate is considered necessary.
These mountings are used for suspension applications such as engine to chassis, cab to chassis,
bus body and tanker tanks to chassis.
Hydroelastic engine mountings (Figs 1.20(a±c) and
1.21) A flanged steel pressing houses and supports an upper and lower rubber spring diaphragm.
The space between both diaphragms is filled and
sealed with fluid and is divided in two by a separator
plate and small transfer holes interlink the fluid
occupying these chambers (Fig. 1.20(a and b)).
Under vertical vibratory conditions the fluid will
be displaced from one chamber to the other
through transfer holes. During downward deflection (Fig. 1.20(b)), both rubber diaphragms are
subjected to a combined shear and compressive
action and some of the fluid in the upper chamber
will be pushed into the lower and back again by
way of the transfer holes when the rubber rebounds
(Fig. 1.20(a)). For low vertical vibratory frequencies,
Double inclined rectangular sandwich mounting
(Fig. 1.18(h)) A pair of rectangular sandwich
rubber blocks are supported on the slopes of a
triangular pedestal. A bridging plate merges the
resilience of the inclined rubber blocks so that
they provide a combined shear and compressive
the movement of fluid between the chambers is
unrestricted, but as the vibratory frequencies
increase, the transfer holes offer increasing resistance to the flow of fluid and so slow down the up
and down motion of the engine support arm. This
damps and reduces the amplitude of mountings
vertical vibratory movement over a number of
cycles. A comparison of conventional rubber and
hydroelastic damping resistance over the normal
operating frequency range for engine mountings is
shown in Fig. 1.20(c).
Instead of adopting a combined rubber mount
with integral hydraulic damping, separate diagonally mounted telescopic dampers may be used in
conjunction with inclined rubber mounts to reduce
both vertical and horizontal vibration (Fig. 1.21).
1.3 Fifth wheel coupling assembly
(Fig. 1.22(a and b))
The fifth wheel coupling attaches the semi-trailer to
the tractor unit. This coupling consists of a semicircular table plate with a central hole and a vee
section cut-out towards the rear (Fig. 1.22(b)).
Attached underneath this plate are a pair of pivoting coupling jaws (Fig. 1.22(a)). The semi-trailer
has an upper fifth wheel plate welded or bolted to
the underside of its chassis at the front and in the
centre of this plate is bolted a kingpin which faces
downwards (Fig. 1.22(a)).
When the trailer is coupled to the tractor unit,
this upper plate rests and is supported on top of the
tractor fifth wheel table plate with the two halves of
the coupling jaws engaging the kingpin. To permit
Fig. 1.19 (a±c) Flanged sleeve bobbin mounting with
rebound control
relative swivelling between the kingpin and jaws,
the two interfaces of the tractor fifth wheel
tables and trailer upper plate should be heavily
greased. Thus, although the trailer articulates
about the kingpin, its load is carried by the tractor
Flexible articulation between the tractor and
semi-trailer in the horizontal plane is achieved by
permitting the fifth wheel table to pivot on horizontal trunnion bearings that lie in the same vertical
plane as the kingpin, but with their axes at right
angles to that of the tractor's wheel base (Fig.
1.22(b)). Rubber trunnion rubber bushes normally
provide longitudinal oscillations of about 10 .
The fifth wheel table assembly is made from
either a machined cast or forged steel sections, or
from heavy section rolled steel fabrications, and the
upper fifth wheel plate is generally hot rolled steel
welded to the trailer chassis. The coupling locking
system consisting of the jaws, pawl, pivot pins and
kingpin is produced from forged high carbon manganese steels and the pressure areas of these components are induction hardened to withstand shock
loading and wear.
Fig. 1.20 (a±c)
1.3.1 Operation of twin jaw coupling
(Fig. 1.23(a±d))
With the trailer kingpin uncoupled, the jaws will be
in their closed position with the plunger withdrawn
from the lock gap between the rear of the jaws,
which are maintained in this position by the pawl
contacting the hold-off stop (Fig. 1.23(a)). When
coupling the tractor to the trailer, the jaws of the
Hydroelastic engine mount
Fig. 1.21 Diagonally mounted hydraulic dampers suppress both vertical and horizontal vibrations
fifth wheel strike the kingpin of the trailer. The
jaws are then forced open and the kingpin enters
the space between the jaws (Fig. 1.23(b)). The kingpin contacts the rear of the jaws which then
automatically pushes them together. At the same
time, one of the coupler jaws causes the trip pin to
strike the pawl. The pawl turns on its pivot against
the force of the spring, releasing the plunger, allowing it to be forced into the jaws' lock gap by its
spring (Fig. 1.23(c)). When the tractor is moving,
the drag of the kingpin increases the lateral force of
the jaws on the plunger.
To disconnect the coupling, the release hand
lever is pulled fully back (Fig. 1.23(d)). This
draws the plunger clear of the rear of the jaws
and, at the same time, allows the pawl to swing
round so that it engages a projection hold-off stop
situated at the upper end of the plunger, thus jamming the plunger in the fully out position in readiness for uncoupling.
spring load notched pawl will then snap over the
jaw projection to lock the kingpin in the coupling
position (Fig. 1.24(c)). The securing pin should
then be inserted through the pull lever and table
eye holes. When the tractor is driving forward, the
reaction on the kingpin increases the locking
force between the jaw projection and the notched
To disconnect the coupling, lift out the securing
pin and pull the release hand lever fully out
(Fig. 1.24(d)). With both the tractor and trailer
stationary, the majority of the locking force
applied to notched pawl will be removed so that
with very little effort, the pawl is able to swing clear
of the jaw in readiness for uncoupling, that is, by
just driving the tractor away from the trailer. Thus
the jaw will simply swivel allowing the kingpin to
pull out and away from the jaw.
1.4 Trailer and caravan drawbar couplings
1.4.1 Eye and bolt drawbar coupling for heavy
goods trailers (Figs 1.25 and 1.26)
Drawbar trailers are normally hitched to the truck
by means of an `A' frame drawbar which is coupled
by means of a towing eye formed on the end of the
drawbar (Fig. 1.25). When coupled, the towing eye
hole is aligned with the vertical holes in the upper
and lower jaws of the truck coupling and an eye
bolt passes through both coupling jaws and drawbar eye to complete the attachment (Fig. 1.26).
Lateral drawbar swing is permitted owing to the
eye bolt pivoting action and the slots between the
1.3.2 Operation of single jaw and pawl coupling
(Fig. 1.24(a±d))
With the trailer kingpin uncoupled, the jaw will be
held open by the pawl in readiness for coupling
(Fig. 1.24(a)). When coupling the tractor to the
trailer, the jaw of the fifth wheel strikes the kingpin
of the trailer and swivels the jaw about its pivot pin
against the return spring, slightly pushing out the
pawl (Fig. 1.24(b)). Further rearward movement of
the tractor towards the trailer will swing the jaw
round until it traps and encloses the kingpin. The
Fig. 1.22 (a and b)
Fifth wheel coupling assembly
jaws on either side. Aligning the towing eye to the
jaws is made easier by the converging upper and
lower lips of the jaws which guide the towing eye as
the truck is reversed and the jaws approach the
drawbar. Isolating the coupling jaws from the
truck draw beam are two rubber blocks which act
as a damping media between the towing vehicle and
trailer. These rubber blocks also permit additional
deflection of the coupling jaw shaft relative to the
draw beam under rough abnormal operating conditions, thus preventing over-straining the drawbar
and chassis system.
Fig. 1.23 (a±d)
Fifth wheel coupling with twin jaws plunger and pawl
Fig. 1.24 (a±d)
Fifth wheel coupling with single jaw and pawl
1.4.2 Ball and socket towing bar coupling for
light caravan/trailers (Fig. 1.27)
Light trailers or caravans are usually attached to
the rear of the towing car by means of a ball and
socket type coupling. The ball part of the attachment is bolted onto a bracing bracket fitted directly
to the boot pan or the towing load may be shared
out between two side brackets attached to the rear
longitudinal box-section members of the body.
A single channel section or pair of triangularly
arranged angle-section arms may be used to form
the towbar which both supports and draws the
Attached to the end of the towbar is the socket
housing with an internally formed spherical cavity.
This fits over the ball member of the coupling so
that it forms a pivot joint which can operate in both
the horizontal and vertical plane (Fig. 1.27).
To secure the socket over the ball, a lock device
must be incorporated which enables the coupling to
be readily connected or disconnected. This lock
may take the form of a spring-loaded horizontally
positioned wedge with a groove formed across its
top face which slips underneath and against the
ball. The wedge is held in the closed engaged position by a spring-loaded vertical plunger which has
a horizontal groove cut on one side. An uncoupling
lever engages the plunger's groove so that when the
coupling is disconnected the lever is squeezed to lift
and release the plunger from the wedge. At the
same time the whole towbar is raised by the handle
to clear the socket and from the ball member.
Coupling the tow bar to the car simply reverses
the process, the uncoupling lever is again squeezed
against the handle to withdraw the plunger and the
socket housing is pushed down over the ball member. The wedge moves outwards and allows the ball
to enter the socket and immediately the wedge
springs back into the engaged position. Releasing
the lever and handle completes the coupling by
permitting the plunger to enter the wedge lock
Sometimes a strong compression spring is interposed between the socket housing member and the
towing (draw) bar to cushion the shock load when
the car/trailer combination is initially driven away
from a standstill.
Fig. 1.25 Drawbar trailer
The coupling jaws, eye bolt and towing eye are
generally made from forged manganese steel with
induction hardened pressure areas to increase the
wear resistance.
Operation of the automatic drawbar coupling
(Fig. 1.26) In the uncoupled position the eyebolt
is held in the open position ready for coupling
(Fig. 1.26(a)). When the truck is reversed, the jaws
of the coupling slip over the towing eye and in the
process strike the conical lower end of the eye bolt
(Fig. 1.26(b)). Subsequently, the eye bolt will lift. This
trips the spring-loaded wedge lever which now rotates
clockwise so that it bears down on the eye bolt.
Further inward movement of the eye bolt between
the coupling jaws aligns the towing eye with the eye
bolt. The spring pressure now acts through the wedge
lever to push the eye bolt through the towing eye and
the lower coupling jaw (Fig. 1.26(c)). When the eye
bolt stop-plate has been fully lowered by the spring
tension, the wedge lever will slot into its groove
formed in the centre of the eye bolt so that it locks
the eye bolt in the coupled position.
To uncouple the drawbar, the handle is pulled
upwards against the tension of the coil spring
mounted on the wedge level operating shaft
(Fig. 1.26(d)). This unlocks the wedge, freeing the
eyebolt and then raises the eye bolt to the
uncoupled position where the wedge lever jams it
in the open position (Fig. 1.26(a)).
1.5 Semi-trailer landing gear (Fig. 1.28)
Landing legs are used to support the front of the
semi-trailer when the tractor unit is uncoupled.
Extendable landing legs are bolted vertically to
each chassis side-member behind the rear wheels of
Fig. 1.26 (a±e)
Automatic drawbar coupling
1.6 Automatic chassis lubrication system
1.6.1 The need for automatic lubrication system
(Fig. 1.29)
Owing to the heavy loads they carry commercial
vehicles still prefer to use metal to metal joints which
are externally lubricated. Such joints are kingpins
and bushes, shackle pins and bushes, steering ball
joints, fifth wheel coupling, parking brake linkage
etc. (Fig. 1.29). These joints require lubricating in
proportion to the amount of relative movement and
the loads exerted. If lubrication is to be effective in
reducing wear between the moving parts, fresh oil
must be pumped between the joints frequently. This
can best be achieved by incorporating an automatic
lubrication system which pumps oil to the bearing's
surfaces in accordance to the distance travelled by
the vehicle.
Fig. 1.27 Ball and socket caravan/trailer towing
the tractor unit, just sufficiently back to clear the
rear tractor road wheels when the trailer is coupled
and the combination is being manoeuvred
(Fig. 1.28(a)). To provide additional support for
the legs, bracing stays are attached between the legs
and from the legs diagonally to the chassis crossmember (Fig. 1.28(b)).
The legs consist of inner and outer high tensile
steel tubes of square section. A jackscrew with a
bevel wheel attached at its top end supported by the
outer leg horizontal plate in a bronze bush bearing.
The jawscrew fits into a nut which is mounted at
the top of the inner leg and a taper roller bearing
race is placed underneath the outer leg horizontal
support plate and the upper part of the jackscrew
to minimize friction when the screw is rotated (Fig.
1.28(b)). The bottom ends of the inner legs may
support either twin wheels, which enable the trailer
to be manoeuvred, or simply flat feet. The latter are
able to spread the load and so permit greater load
To extend or retract the inner legs, a winding
handle is attached to either the low or high speed
shaft protruding from the side of the gearbox. The
upper high speed shaft supports a bevel pinion
which meshes with a vertically mounted bevel
wheel forming part of the jackscrew.
Rotating the upper shaft imparts motion directly
to the jackscrew through the bevel gears. If greater
leverage is required to raise or lower the front of the
trailer, the lower shaft is engaged and rotated.
This provides a gear reduction through a compound gear train to the upper shaft which then
drives the bevel pinion and wheel and hence the
1.6.2 Description of airdromic automatic chassis
lubrication system (Fig. 1.30)
This lubrication system comprises four major components; a combined pump assembly, a power unit,
an oil unloader valve and an air control unit.
Pump assembly (Fig. 1.30) The pump assembly
consists of a circular housing containing a ratchet
operated drive (cam) shaft upon which are
mounted one, two or three single lobe cams (only
one cam shown). Each cam operates a row of 20
pumping units disposed radially around the pump
casing, the units being connected to the chassis
bearings by nylon tubing.
Power unit (Fig. 1.30) This unit comprises a
cylinder and spring-loaded air operated piston
which is mounted on the front face of the pump
assembly housing, the piston rod being connected
indirectly to the drive shaft ratchet wheel by way of
a ratchet housing and pawl.
Oil unloader valve (Fig. 1.30) This consists of a
shuttle valve mounted on the front of the pump
assembly housing. The oil unloader valve allows air
pressure to flow to the power unit for the power
stroke. During the exhaust stroke, however, when
air flow is reversed and the shuttle valve is lifted
from its seat, any oil in the line between the power
unit and the oil unloader valve is then discharged to
Fig. 1.28 (a and b)
Semi-trailer landing gear
Fig. 1.29 Tractor unit automatic lubrication system
Air control unit (Fig. 1.30) This unit is mounted
on the gearbox and is driven via the speedometer
take-off point. It consists of a worm and wheel drive
which operates an air proportioning control
unit. This air proportioning unit is operated by a
single lift face cam which actuates two poppet
valves, one controlling air supply to the power
unit, the other controlling the exhaust air from the
power unit.
housing. Because the pawl meshes with one of the
ratchet teeth and the ratchet wheel forms part of
the camshaft, air pressure in the power cylinder will
partially rotate both the ratchet and pawl housing
and the camshaft clockwise. The cam (or cams) are
in contact with one or more pump unit, and so each
partial rotation contributes to a proportion of the
jerk plunger and barrel pumping cycle of each unit
(Fig. 1.30).
As the control unit face cam continues to rotate,
the inlet poppet inlet valve is closed and the exhaust
poppet valve opens. Compressed air in the air control unit and above the oil control shuttle valve will
now escape through the air control unit exhaust
port to the atmosphere. Consequently the compressed air underneath the oil unloader shuttle
valve will be able to lift it and any trapped air and
oil in the power cylinder will now be released via
the hole under the exhaust port. The power unit
piston will be returned to its innermost position by
the spring and in doing so will rotate the ratchet
and pawl housing anti-clockwise. The pawl is thus
1.6.3 Operation of airdromic automatic chassis
lubrication system (Fig. 1.30)
Air from the air brake auxiliary reservoir passes by
way of the safety valve to the air control (proportioning) unit inlet valve. Whilst the inlet valve is
held open by the continuously rotating face cam
lobe, air pressure is supplied via the oil unloader
valve to the power unit attached to the multipump
assembly housing. The power unit cylinder is supported by a pivot to the pump assembly casing,
whilst the piston is linked to the ratchet and pawl
Fig. 1.30
Airdromic automatic chassis lubrication system
able to slip over one or more of the ratchet teeth to
take up a new position. The net result of the power
cylinder being charged and discharged with compressed air is a slow but progressive rotation of the
camshaft (Fig. 1.30).
A typical worm drive shaft to distance travelled
relationship is 500 revolutions per 1 km. For 900
worm drive shaft revolutions the pumping cam
revolves once. Therefore, every chassis lubrication
point will receive one shot of lubricant in this
When the individual lubrication pump unit's
primary plunger is in its outermost position, oil
surrounding the barrel will enter the inlet port,
filling the space between the two plungers. As the
cam rotates and the lobe lifts the primary plunger,
it cuts off the inlet port. Further plunger rise will
partially push out the secondary plunger and so
open the check valve. Pressurised oil will then
pass between the loose fitting secondary plunger
and barrel to lubricate the chassis moving part it
services (Fig. 1.30).
Friction clutch
2.1 Clutch fundamentals
Clutches are designed to engage and disengage the
transmission system from the engine when a vehicle
is being driven away from a standstill and when the
gearbox gear changes are necessary. The gradual
increase in the transfer of engine torque to the
transmission must be smooth. Once the vehicle is
in motion, separation and take-up of the drive for
gear selection must be carried out rapidly without
any fierceness, snatch or shock.
2.1.3 Multi-pairs of rubbing surfaces (Fig. 2.1)
An alternative approach to raising the transmitted
torque capacity of the clutch is to increase the
number of pairs of rubbing surfaces. Theoretically
the torque capacity of a clutch is directly proportional to the number of pairs of surfaces for a given
clamping load. Thus the conventional single driven
plate has two pairs of friction faces so that a twin
or triple driven plate clutch for the same spring
thrust would ideally have twice or three times the
torque transmitting capacity respectively of that of
the single driven plate unit (Fig. 2.1). However,
because it is very difficult to dissipate the extra
heat generated in a clutch unit, a larger safety factor
is necessary per driven plate so that the torque
capacity is generally only of the order 80% per pair
of surfaces relative to the single driven plate clutch.
2.1.1 Driven plate inertia
To enable the clutch to be operated effectively, the
driven plate must be as light as possible so that
when the clutch is disengaged, it will have the minimum of spin, i.e. very little flywheel effect. Spin
prevention is of the utmost importance if the various pairs of dog teeth of the gearbox gears, be they
constant mesh or synchromesh, are to align in the
shortest time without causing excessive pressure,
wear and noise between the initial chamfer of the
dog teeth during the engagement phase.
Smoothness of clutch engagement may be
achieved by building into the driven plate some
sort of cushioning device, which will be discussed
later in the chapter, whilst rapid slowing down of
the driven plate is obtained by keeping the diameter,
centre of gravity and weight of the driven plate to
the minimum for a given torque carrying capacity.
2.1.4 Driven plate wear (Fig. 2.1)
Lining life is also improved by increasing the
number of pairs of rubbing surfaces because wear
is directly related to the energy dissipation per unit
area of contact surface. Ideally, by doubling the
surface area as in a twin plate clutch, the energy
input per unit lining area will be halved for a given
slip time which would result in a 50% decrease in
facing wear. In practice, however, this rarely occurs
(Fig. 2.1) as the wear rate is also greatly influenced
by the peak surface rubbing temperature and the
intermediate plate of a twin plate clutch operates at
a higher working temperature than either the flywheel or pressure plate which can be more effectively cooled. Thus in a twin plate clutch, half the
energy generated whilst slipping must be absorbed
by the intermediate plate and only a quarter each
by the flywheel and pressure plate. This is usually
borne out by the appearance of the intermediate
plate and its corresponding lining faces showing
evidence of high temperatures and increased wear
compared to the linings facing the flywheel and
pressure plate. Nevertheless, multiplate clutches
do have a life expectancy which is more or less
related to the number of pairs of friction faces for
a given diameter of clutch.
For heavy duty applications such as those
required for large trucks, twin driven plates are
used, while for high performance cars where very
2.1.2 Driven plate transmitted torque capacity
The torque capacity of a friction clutch can be
raised by increasing the coefficient of friction of
the rubbing materials, the diameter and/or the
spring thrust sandwiching the driven plate. The
friction lining materials now available limit the
coefficient of friction to something of the order of
0.35. There are materials which have higher coefficient of friction values, but these tend to be
unstable and to snatch during take-up. Increasing
the diameter of the driven plate unfortunately
raises its inertia, its tendency to continue spinning
when the driven plate is freed while the clutch is in
the disengaged position, and there is also a limit to
the clamping pressure to which the friction lining
material may be subjected if it is to maintain its
friction properties over a long period of time.
Fig. 2.1 Relationship of torque capacity wear rate and pairs of rubbing faces for multiplate clutch
rapid gear changes are necessary and large
amounts of power are to be developed, small
diameter multiplate clutches are preferred.
2.2 Angular driven plate cushioning and torsional
damping (Figs 2.2±2.8)
2.2.1 Axial driven plate friction lining cushioning
(Figs 2.2, 2.3 and 2.4)
In its simplest form the driven plate consists of
a central splined hub. Mounted on this hub is a
thin steel disc which in turn supports, by means of
a ring of rivets, both halves of the annular friction
linings (Figs 2.2 and 2.3).
Axial cushioning between the friction lining
faces may be achieved by forming a series of evenly
spaced `T' slots around the outer rim of the disc.
This then divides the rim into a number of segments (Arcuate) (Fig. 2.4(a)). A horseshoe shape
is further punched out of each segment. The central
portion or blade of each horseshoe is given a permanent set to one side and consecutive segments
have opposite sets so that every second segment is
riveted to the same friction lining. The alternative
set of these central blades formed by the horseshoe
punch-out spreads the two half friction linings apart.
An improved version uses separately attached, very
thin spring steel segments (borglite) (Fig. 2.4(b)), positioned end-on around a slightly thicker disc plate.
These segments are provided with a wavy `set' so as
to distance the two half annular friction linings.
Both forms of crimped spring steel segments
situated between the friction linings provide
Fig. 2.2
Clutch driven centre plate (pictorial view)
Fig. 2.3 Clutch driven centre plate (sectional view)
Fig. 2.4 (a and b)
progressive take-up over a greater pedal travel and
prevent snatch. The separately attached spring
segments are thinner than the segments formed out
of the single piece driven plate, so that the squeeze
take-up is generally softer and the spin inertia of the
thinner segments is noticeably reduced.
A further benefit created by the spring segments
ensures satisfactory bedding of the facing material
and a more even distribution of the work load. In
addition, cooling between the friction linings occurs
when the clutch is disengaged which helps to stabilise the frictional properties of the face material.
The advantages of axial cushioning of the face
linings provide the following:
Driven plate cushion take-up
The spring take-up characteristics of the driven
plate are such that when the clutch is initially
engaged, the segments are progressively flattened so
that the rate of increase in clamping load is provided
by the rate of reaction offered by the spring
segments (Fig. 2.5). This first low rate take-up
period is followed by a second high rate engagement, caused by the effects of the pressure plate
springs exerting their clamping thrust as they are
allowed to expand against the pressure plate and
so sandwich the friction lining between the flywheel
and pressure plate faces.
2.2.2 Torsional damping of driven plate
a) Better clutch engagement control, allowing
lower engine speeds to be used at take-up thus
prolonging the life of the friction faces.
b) Improved distribution of the friction work over
the lining faces reduces peak operating temperatures and prevents lining fade, with the resulting
reduction in coefficient of friction and subsequent clutch slip.
Crankshaft torsional vibration (Fig. 2.6) Engine
crankshafts are subjected to torsional wind-up
and vibration at certain speeds due to the power
impulses. Superimposed onto some steady mean
rotational speed of the crankshaft will be additional
fluctuating torques which will accelerate and decelerate the crankshaft, particularly at the front pulley
the gear teeth, wear, and noise in the form of
gear clatter. To overcome the effects of crankshaft
torsional vibrations a torsion damping device is
normally incorporated within the driven plate hub
assembly which will now be described and explained.
Construction and operation of torsional damper
springs (Figs 2.2, 2.3 and 2.7) To transmit torque
more smoothly and progressively during take-up of
normal driving and to reduce torsional oscillations
being transmitted from the crankshaft to the transmission, compressed springs are generally arranged
circumferentially around the hub of the driven
plate (Figs 2.2 and 2.3). These springs are inserted
in elongated slots formed in both the flange of the
splined hub and the side plates which enclose the
hub's flange (Fig. 2.3). These side plates are riveted
together by either three or six rivet posts which pass
through the flanged hub limit slots. This thus
provides a degree of relative angular movement
between hub and side plates. The ends of the helical
coil springs bear against both central hub flange
and the side plates. Engine torque is therefore
transmitted from the friction face linings and side
plates through the springs to the hub flange, so that
any fluctuation of torque will cause the springs to
compress and rebound accordingly.
Multistage driven plate torsional spring dampers
may be incorporated by using a range of different
springs having various stiffnesses and spring location slots of different lengths to produce a variety
of parabolic torsional load±deflection characteristics (Fig. 2.7) to suit specific vehicle applications.
The amount of torsional deflection necessary
varies for each particular application. For example,
with a front mounted engine and rear wheel drive
vehicle, a moderate driven plate angular movement
is necessary, say six degrees, since the normal transmission elastic wind-up is almost adequate, but with
an integral engine, gearbox and final drive arrangement, the short transmission drive length necessitates considerably more relative angular deflection,
say twelve degrees, within the driven plate hub
assembly to produce the same quality of take-up.
Fig. 2.5 Characteristics of driven plate axial clamping
load to deflection take-up
Fig. 2.6 Characteristics of crankshaft torsional
vibrations undamped and damped
end and to a lesser extent the rear flywheel end
(Fig. 2.6). If the flywheel end of the crankshaft
were allowed to twist in one direction and then the
other while rotating at certain critical speeds, the
oscillating angular movements would take up the
backlash between meshing gear teeth in the transmission system. Consequently, the teeth of the driving
gears would be moving between the drive (pressure
side) and non-drive tooth profiles of the driven gears.
This would result in repeated shockloads imposed on
Construction and operation of torsional damper
washers (Figs 2.2, 2.3 and 2.8) The torsional
energy created by the oscillating crankshaft is
partially absorbed and damped by the friction
washer clutch situated on either side of the hub
flange (Figs 2.2 and 2.3). Axial damping load is
achieved by a Belleville dished washer spring
mounted between one of the side plates and a four
lug thrust washer.
Fig. 2.8 Characteristics of driven plate torsional
damper thrust spring
Fig. 2.7 Characteristics of driven plate torsional spring
torques to deflection take-up
The outer diameter of this dished spring presses
against the side plate and the inner diameter pushes
onto the lugged thrust washer. In its free state
the Belleville spring is conical in shape but when
assembled it is compressed almost flat. As the friction washers wear, the dished spring cone angle
increases. This exerts a greater axial thrust, but
since the distance between the side plate and lugged
thrust washer has increased, the resultant clamping
thrust remains almost constant (Fig. 2.8).
i) a high degree of interface contamination tolerance without affecting its friction take-up and
grip characteristics.
2.3.1 Asbestos-based linings (Figs 2.2 and 2.3)
Generally, clutch driven plate asbestos-based linings are of the woven variety. These woven linings
are made from asbestos fibre spun around lengths
of brass or zinc wire to make lengths of threads
which are both heat resistant and strong. The
woven cloth can be processed in one of two ways:
2.3 Clutch friction materials
Clutch friction linings or buttons are subjected to
severe rubbing and generation of heat for relatively
short periods. Therefore it is desirable that they
have a combination of these properties:
a) The fibre wire thread is woven into a cloth and
pressed out into discs of the required diameter,
followed by stitching several of these discs
together to obtain the desired thickness. The
resultant disc is then dipped into resin to bond
the woven asbestos threads together.
b) The asbestos fibre wire is woven in three dimensions in the form of a disc to obtain in a single
stage the desired thickness. It is then pressed
into shape and bonded together by again dipping it into a resin solution. Finally, the rigid
lining is machined and drilled ready for riveting
to the driven plate.
a) Relatively high coefficient of friction under
operating conditions,
b) capability of maintaining friction properties
over its working life,
c) relatively high energy absorption capacity for
short periods,
d) capability of withstanding high pressure plate
compressive loads,
e) capability of withstanding bursts of centrifugal
force when gear changing,
f) adequate shear strength to transmit engine
g) high level of cyclic working endurance without
the deterioration in friction properties,
h) good compatibility with cast iron facings over
the normal operating temperature range,
Development in weaving techniques has, in
certain cases, eliminated the use of wire coring so
that asbestos woven lining may be offered as either
non- or semi-metallic to match a variety of working
Asbestos is a condensate produced by the solidification of rock masses which cool at differential
rates. When the moisture content of one layer
is transferred to another, fibres are produced on
solidification from which, as a result of high compression, these brittle, practically straight and
exceptionally fine needle-like threads are made.
During processing, these break down further with
a diameter of less than 0.003 mm. They exhibit a
length/thickness ratio of at least three to one. It is
these fine fibres which can readily be inhaled into
the lungs which are so dangerous to health.
The normal highest working temperature below
which these asbestos linings will operate satisfactorily giving uniform coefficient of friction between
0.32 and 0.38 and a reasonable life span is about
260 C. Most manufacturers of asbestos-based
linings quote a maximum temperature (something
like 360 C) beyond which the lining, if operated
continuously or very frequently, will suffer damage,
with consequent alteration to its friction characteristics and deterioration in wear resistance.
these fibres are difficult to inhale because of their
shape and size.
2.3.3 Metallic friction materials
Metallic and semi-metallic facings have been only
moderately successful. The metallic linings are
normally made from either sintered iron or copperbased sintered bronze and the semi-metallic facings
from a mixture of organic and metallic materials.
Metallic lining materials are made from a powder
produced by crushing metal or alloy pieces into
many small particles. They are then compressed
and heated in moulds until sufficient adhesion and
densification takes place. This process is referred to
as sintering. The metallic rings are then ground flat
and are then riveted back to back onto the driven
Generally the metallic and semi-metallic linings
have a higher coefficient of friction, can operate at
higher working temperatures, have greater torque
capacity and have extended life compared to that
of the organic asbestos based linings. The major
disadvantages of metallic materials are their
relatively high inertia, making it difficult to obtain
rapid gear changes; high quality flywheel and pressure plate. Cast iron must be used to match their
friction characteristics and these facings are more
expensive than organic materials.
2.3.2 Asbestos substitute friction material
(Figs 2.2 and 2.3)
The DuPont Company has developed a friction
material derived from aromatic polyamide fibres
belonging to the nylon family of polymers and it
has been given the trade name Kevlar aramid.
The operating properties relative to asbestos
based linings are as follows:
1 High endurance performance over its normal
working pressure and temperature range.
2 It is lighter in weight than asbestos material
therefore a reduction in driven plate spin shortens the time required for gear changing.
3 Good take-up characteristics, particularly with
vehicles which were in the past prone to grab.
4 Weight for weight Kevlar has five times the
tensile strength of steel.
5 Good centrifugal strength to withstand lining
disintegration as a result of sudden acceleration
which may occur during the changing of gears.
6 Stable rubbing properties at high operating
temperatures. It is not until a temperature of
425 C is reached that it begins to break down
and then it does not simply become soft and
melt, but steadily changes to carbon, the disintegration process being completed at about
500 C.
2.3.4 Cerametallic friction materials (Fig. 2.9)
Cerametallic button friction facings are becoming
increasingly popular for heavy duty clutches.
Instead of a full annular shaped lining, as with
organic (asbestos or substitute) friction materials,
four or six cerametallic trapezoidal-shaped buttons
are evenly spaced on both sides around the driven
plate. The cerametallic material is made from a
powder consisting mainly of ceramic and copper.
It is compressed into buttons and heated so that
the copper melts and flows around each particle of
solid ceramic. After solidification, the copper
forms a strong metal-ceramic interface bond.
These buttons are then riveted to the clutch driven
plate and then finally ground flat.
The inherent advantages of these cerametalliclined driven plates are:
1 A very low inertia (about 10% lower than the
organic disc and 45% lower than a comparable
sintered iron disc). Consequently it will result in
quicker gear changes and, in the case of synchronized transmission, will increase synchronizer life.
2 A relatively high and stable coefficient of friction,
providing an average value in the region of
Kevlar exists in two states; firstly as a 0.12 mm
thick endless longitudinal fibre, which has a cut
length varying between 6 and 100 mm, and secondly
in the form of an amorphous structure of crushed
and ground fibre known as pulp. In either form
flywheel and pressure plate facings. A prolonged
development programme has virtually eliminated this problem and has considerably extended
the driven plate life span compared to driven
plates using organic (asbestos-based) annular
disc linings.
2.4 Clutch drive and driven member inspection
This inspection entails the examination of both the
driven plate linings and the flywheel and pressure
plate facings and will now be considered.
2.4.1 Driven plate lining face inspection
Driven plate friction facings should, after a short
period of service, give a polished appearance due to
the frequent interface rubbing effect. This smooth
and polished condition will provide the greatest
friction grip, but it must not be confused with a
glazed surface created by the formation of films of
grease or oil worked into the rubbing surfaces,
heated and oxidized.
A correctly bedded-in friction facing will appear
highly polished through which the grain of the
friction material can be clearly seen. When in
perfect condition, these polished facings are of a
grey or mid-brown colour. A very small amount
of lubricant on the facings will burn off due to the
generated heat. This will only slightly darken the
facings, but providing polished facings remain so
that the grain of the material can be clearly distinguished, it does not reduce its effectiveness.
Large amounts of lubricant gaining access to the
friction surfaces may result in the following:
Fig. 2.9 Clutch driven plate with ceramic facings
0.4, which increases the torque capacity of
clutches using these driven plates.
The capability of operating at high working
temperatures of up to 440 C for relatively long
periods without showing signs of fade.
Button type driven plates expose more than 50%
of the flywheel and pressure plate surfaces to the
atmosphere during clutch engagement, so that
heat transfer to the surrounding by convection
may be improved by as much as 100%.
Button type friction pads do not suffer from
warpage as do full ring metallic or organic linings
and therefore are less prone to distort and cause
clutch drag.
Button type friction pads permit the dust worn
from the friction surfaces to be thrown clear of
the clutch areas, thus preventing the possibility
of any trapped work-hardened particles from
scoring the friction faces.
Cerametallic materials are not as sensitive to
grease and oil contamination as organic asbestos
based linings.
The early ceramic-metallic friction buttons had a
poor reputation as they tended to wear tracks in
a) The burning of the grease or oil may leave a carbon
deposit and a high glaze, this hides the grain of the
material and is likely to cause clutch slip.
b) If the grease or oil is only partially burnt and
leaves a resinous deposit on the facings it may
result in a fierce clutch and may in addition
produce clutch spin caused by the rubbing interfaces sticking.
c) If both carbon and resinous deposits are formed
on the linings, clutch judder may develop during
clutch take-up.
2.4.2 Flywheel and pressure plate facing inspection
Cast iron flywheel or pressure plate faces should
have a smooth polished metallic appearance, but
abnormal operating conditions or badly worn
driven plate linings may be responsible for the
following defects:
a) Overheated clutch friction faces can be identified by colouring of the swept polished tracks.
The actual surface temperatures reached can be
distinguished broadly by the colours; straw,
brown, purple and blue which relate to 240 C,
260 C, 280 C and 320 C respectively.
b) Severe overheating will create thermal stresses
within the cast iron mass of the flywheel and
pressure plate, with the subsequent appearance
of radial hairline cracks.
c) Excessively worn driven plate linings with
exposed rivets and trapped work-hardened
dust particles will cause scoring of the rubbing
faces in the form of circular grooves.
observe the reading. Acceptable end float values
are normally between 0.08 and 0.30 mm.
2.5.2 Crankshaft flywheel flange runout
(Fig. 2.10(a))
The crankshaft flange flywheel joint face must be
perpendicular to its axis of rotation with no permissible runout. To check for any misalignment, keep
the dial gauge assembly mounted as for the end
float check. Zero gauge the dial and rotate the
crankshaft by hand for one complete revolution
whilst observing any dial movement. Investigate
further if runout exists.
2.5 Clutch misalignment (Fig. 2.10(a±d))
Clutch faults can sometimes be traced to misalignment of the crankshaft to flywheel flange
joint, flywheel housing and bell housing. Therefore,
if misalignment exists, the driven plate plane of
rotation will always be slightly skewed to that of
the restrained hub which is made to rotate about
the spigot shaft's axis. Misalignment is generally
responsible for the following faults:
2.5.3 Flywheel friction face and rim face runout
(Fig. 2.10(a and b))
When the flywheel is centred by the crankshaft axis,
it is essential that the flywheel friction face and rim
rotate perpendicularly to the crankshaft axis.
Mount the dial gauge magnetic base to the
engine flywheel housing. First set the indicator
pointer against the friction face of the flywheel
near the outer edge (Fig. 2.10(a and b)) and set
gauge to zero. Turn the flywheel one revolution
and observe the amount of variation. Secondly
reset indicator pointer against the flywheel rim
and repeat the test procedure (Fig. 2.10(b)). Maximum permissible runout in both tests is 0.02 mm
per 20 mm of flywheel radius. Thus with a 300 mm
diameter clutch fitted, maximum run-out would be
0.15 mm. Repeat both tests 2 or 3 times and compare readings to eliminate test error.
1 Rapid wear on the splines of the clutch driven
plate hub, this being caused mainly by the tilted
hub applying uneven pressure over the interface
length of the splines.
2 The driven plate breaking away from the splined
hub due to the continuous cyclic flexing of the
plate relative to its hub.
3 Excessively worn pressure plate release mechanism, causing rough and uneven clutch
4 Fierce chattering or dragging clutch resulting in
difficult gear changing.
2.5.4 Flywheel housing runout (Fig. 2.10(c))
When the gearbox bell housing is centred by the
inside diameter and rear face of the engine flywheel
housing, it is essential that the inside diameter and
rear face of the housing should be concentric and
parallel respectively with the flywheel.
Mount the dial gauge magnetic base to the flywheel friction face and position. Set the indicator
pointer against the face of the housing. Make sure
that the pointer is not in the path of the fixing holes
in the housing face or else incorrect readings may
result. Zero the indicator and observe the reading
whilst the crankshaft is rotated one complete revolution. Reset the indicator pointer against the internally machined recess of the clutch housing and repeat
the test procedure. Maximum permissible runout is
0.20 mm. Repeat both tests two or three times and
compare readings to eliminate errors.
If excessive clutch drag, backlash and poor
changes are evident and the faults cannot be
corrected, then the only remedy is to remove both
gearbox and clutch assembly so that the flywheel
housing alignment can be assessed (Fig. 2.10).
2.5.1 Crankshaft end float (Fig. 2.10(a))
Before carrying out engine crankshaft, flywheel or
flywheel housing misalignment tests, ensure that
the crankshaft end float is within limits. (Otherwise
inaccurate run-out readings may be observed.)
To measure the crankshaft end float, mount the
magnetic dial gauge base to the back of the flywheel
housing and position the indicator pointer against
the crankshaft flanged end face. Zero the dial gauge
and with the assistance of a suitable lever, force the
crankshaft back and forth and, at the same time,
Fig. 2.10 (a±d)
Crankshaft flywheel and clutch housing alignment
2.5.5 Detachable bell housing runout
(Fig. 2.10(c and d))
When the gearbox bell housing is located by dowel
pins instead of the inside diameter of the engine
flywheel housing (Fig. 2.10(c)) (shouldered bell
housing), it is advisable to separate the clutch bell
housing from the gearbox and mount it to the
flywheel housing for a concentric check.
Mount the dial gauge magnetic base onto the
flywheel friction face and position the indicator
pointer against the internal recess of the bell
housing gearbox joint bore (Fig. 2.10(d)). Set the
gauge to zero and turn the crankshaft by hand one
complete revolution. At the same time, observe
the dial gauge reading.
Maximum permissible runout should not exceed
0.25 mm.
2.6 Pull type diaphragm clutch (Fig. 2.11)
With this type of diaphragm clutch, the major components of the pressure plate assembly are a cast iron
pressure plate, a spring steel diaphragm disc and a
low carbon steel cover pressing (Fig. 2.11). To actuate
the clutch release, the diaphragm is made to pivot
between a pivot ring positioned inside the rear of the
cover and a raised circumferential ridge formed on
the back of the pressure plate. The diaphragm disc is
divided into fingers caused by radial slits originating
from the central hole. These fingers act both as leaf
springs to provide the pressure plate thrust and as
release levers to disengage the driven plate from the
drive members.
When the driven and pressure plates are bolted
to the flywheel, the diaphragm is distorted into a
dished disc which therefore applies an axial thrust
between the pressure plate and the cover pressing.
This clutch design reverses the normal method of
operation by pulling the diaphragm spring outwards
to release the driven plate instead of pushing it.
Owing to its configuration, the pull type clutch
allows a larger pressure plate and diaphragm
spring to be used for a given diameter of clutch.
Advantages of this design over a similar push type
clutch include lower pedal loads, higher torque
capacity, improved take-up and increased durability. This clutch layout allows the ratio of the
diaphragm finger release travel to pressure plate
movement to be reduced. It is therefore possible
to maintain the same pressure plate movement as
that offered by a conventional push type clutch,
and yet increase the ratio between clamp load and
pedal load from 4:1 to 5:1.
Fig. 2.11
Diaphragm single plate pull type clutch
2.7 Multiplate diaphragm type clutch (Fig. 2.12)
These clutches basically consist of drive and driven
plate members. The drive plates are restrained from
rotating independently by interlocking lugs and
slots which permit axial movement, but not relative
rotational spin, whilst the driven plates are
attached and supported by internally splined hubs
to corresponding splines formed on the gearbox
spigot shaft, see Fig. 2.12.
The diaphragm spring is in the form of a dished
annular disc. The inner portion of the disc is
radially slotted, the outer ends being enlarged
with a circular hole to prevent stress concentration
when the spring is distorted during disengagement.
These radial slots divide the disc into a number of
release levers (fingers).
The diaphragm spring is located in position with
a shouldered pivot post which is riveted to the
cover pressing. These rivets also hold a pair of
fulcrum rings in position which are situated either
side of the diaphragm.
Whilst in service, the diaphragm cone angle will
change continuously as wear occurs and as the
clutch is engaged and disengaged during operation.
To enable this to happen, the diaphragm pivots
and rolls about the fulcrum rings. When the clutch
is engaged the diaphragm bears against the outer
Fig. 2.12
Multiplate diaphragm type clutch
ring, but when disengagement takes place the reaction load is then taken by the inner ring.
As the friction linings wear, the spring diaphragm will become more dished and subsequently
will initially exert a larger axial clamping load. It is
only when the linings are very worn, so that the
distance between the cover pressing and pressure
plate become excessive, that the axial thrust will
begin to decline.
bearings. There are two types of pressure plate
cover housings; one with a deep extended cover
rim which bolts onto a flat flywheel facing and the
shallow cover type in which the pressure plate
casting fits into a recessed flywheel.
The release mechanism is comprised of three
lever fingers. The outer end of each lever pivots
on a pin and needle race mounted inside each
of the adjustable eye bolt supports, which are
attached to the cover housing through an internally
and externally threaded sleeve which is secured to
the cover housing with a lock nut. Inwards from
the eye bolt, one-sixth of the release lever length, is
a second pin which pivots on a pair of needlebearing races situated inside the pressure plate
lugs formed on either side of each layer.
2.8 Lipe rollway twin driven plate clutch (Fig. 2.13)
These clutches have two circular rows of helical coil
springs which act directly between the pressure
plate and the cover housing, see Fig. 2.13. The
release mechanism is of the pull type in which a
central release bearing assembly is made to withdraw (pull out) three release levers to disengage the
clutch. The clutch pressure plate assembly is bolted
to the flywheel and the driven plate friction linings
are sandwiched between the flywheel, intermediate
plate and pressure plate facings. The central hub of
the driven plates is mounted on a splined gearbox
spigot shaft (input shaft). The splined end of the
input shaft is supported by a ball race bearing
mounted inside the flywheel-crankshaft attachment
flange. The other end of this shaft is supported
inside the gearbox by either ball or taper roller
Release lever adjustment
Initially, setting up of the release levers is achieved
by slackening the locknuts and then rotating each
sleeve in turn with a two pronged fork adaptor tool
which fits into corresponding slots machined out of
the adjustment sleeve end. Rotating the sleeve one
way or the other will screw the eye bolts in or out
until the correct dimension is obtained between the
back of the release lever fingers and the outer cover
rim edge. This setting dimension is provided by the
Fig. 2.13 (a±b)
Twin driven plate pull type clutch
manufacturers for each clutch model and engine
application. Finally, tighten the locknuts of each
eye bolt and re-check each lever dimension again.
this instance one-sixth of the value if the springs
were direct acting.
The operating characteristics of the clutch
mechanism are described as follows:
Release bearing adjustment
Slacken sleeve locknut with a `C' shaped spanner.
Rotate the inner sleeve either way by means of the
slotted adjusting nut until the recommended clearance is obtained between the bearing housing cover
face and clutch brake.
9.5 mm for 355 mm
13 mm for 355 mm
13 mm for 294 mm
New engaged position (Fig. 2.14(a))
The spring thrust horizontal component of 2.2 kN,
multiplied by the lever ratio, provides a pressure
plate clamping load of 13.2 kN (Fig. 2.14(a)). The
axial thrust horizontal component pushing on the
pressure plate does not vary in direct proportion
with the spring load exerted between its ends, but is
a function of the angle through which the mounted
springs operate relative to the splined input shaft.
Finally tighten sleeve locknut and re-check clearance.
Worn engagement position (Fig. 2.14(b))
When the driven plate facings wear, the release
bearing moves forward to the pressure plate so
that the springs elongate. The spring load exerted
between the spring ends is thus reduced. Fortunately, the inclined angle of spring axis to that of
the thrust bearing axis is reduced so that as the
spring load along its axis declines, the horizontal
thrust component remains essentially the same.
Therefore, the pressure plate clamping load
remains practically constant throughout the life of
the clutch (Fig. 2.14(b)).
2.9 Spicer twin driven plate angle spring pull type
clutch (Fig. 2.14)
An interesting clutch engagement and release pressure plate mechanism employs three pairs of coil
springs which are inclined to the axial direction of
the driven plates. These springs are mounted
between the pressure plate cover housing, which
takes the spring reaction, and the release lever central hub (Fig. 2.14). The axial clamping thrust is
conveyed by the springs to the six to one leverage
ratio release levers (six of them) spaced evenly
around the release hub. These release levers span
between the release hub and a large annular shaped
adjustable pivot ring which is screwed inside the
pressure plate cover housing. Towards the pivot
pin end of the release levers a kink is formed so
that it can bear against the pressure plate at one
point. The pressure plate and intermediate plate are
both prevented from spinning with the driven
plates by cast-in drive lugs which fit into slots
formed into the cover housing.
In the engaged position, the six springs expand
and push the release hub and, subsequently, the
release levers towards the pressure plate so that
the driven plates are squeezed together to transmit
the drive torque.
To release the clutch driven plates, the release
bearing assembly is pulled out from the cover housing. This compels the release lever hub to compress
and distort the thrust springs to a much greater
inclined angle relative to the input shaft axis and
so permits the pressure plate to be withdrawn by
means of the retraction springs.
Because the spring thrust does not operate
directly against the pressure plate, but is relayed
through the release levers, the actual spring's stiffness is reduced by a factor of the leverage ratio; in
Release position (Fig. 2.14(c))
When the clutch is released, that is when the bearing is pulled rearwards, the springs compress and
increase in load, but the spring angle relative to the
thrust bearing axis increases so that a greater proportion of the spring load will be acting radially
instead of axially. Consequently, the horizontal
component of axial release bearing load, caused
by the spring thrust, gradually reduces to a value
of about 1.7 kN as the bearing moves forwards,
which results in the reduced pedal effort. This is
shown in Fig. 2.14(c).
Internal manual adjustment
Release bearing adjustment is made by unscrewing
the ring lock plate bolt and removing the plate. The
clutch pedal is then held down to relieve the release
levers and adjusting ring load. The adjusting ring is
then rotated to screw it in or out so that it alters the
release lever hub axial position.
Turning the adjusting ring clockwise moves the
release bearing towards the gearbox (increasing
free pedal movement). Conversely, turning the
adjusting ring anticlockwise moves the release
bearing towards the flywheel (decreasing free
pedal movement).
Fig. 2.14 (a±c) Twin driven plate angle spring pull type clutch
The adjusting ring outer face is notched so that
it can be levered round with a screwdriver when
adjustment is necessary. The distance between each
notch represents approximately 0.5 mm. Thus three
notches moved means approximately 1.5 mm
release bearing movement.
With the pedal released, there should be approximately 13 mm clearance between the release bearing face and clutch brake.
2.10 Clutch (upshift) brake (Fig. 2.15)
The clutch brake is designed primarily for use with
unsynchronized (crash or constant mesh) gearboxes to permit shifting into first and reverse gear
without severe dog teeth clash. In addition, the
brake will facilitate making unshafts by slowing
down the input shaft so that the next higher gear
may be engaged without crunching of teeth.
The brake disc assembly consists of a pair of
Belleville spring washers which are driven by a
hub having internal lugs that engage machined
slots in the input shaft. These washers react against
the clutch brake cover with facing material positioned between each spring washer and outer
cover (Fig. 2.15).
When the clutch pedal is fully depressed, the disc
will be squeezed between the clutch release bearing
housing and the gearbox bearing housing, causing
the input spigot shaft to slow down or stop. The
hub and spring washer combination will slip with
respect to the cover if the applied torque load
exceeds 34 Nm, thus preventing the disc brake
being overloaded.
In general, the clutch brake comes into engagement only during the last 25 mm of clutch pedal
Internal self-adjustment
A clutch self-adjustment version has teeth cut on
the inside of the adjusting ring and a small worm
and spring self-adjusting device replaces the lock
plate. The worm meshes with the adjusting ring.
One end of the spring is located in a hole formed
in the release lever hub whilst the other end is in
contact with the worm. Each time the clutch is
engaged and disengaged, the release lever movement will actuate the spring. Consequently, once
the driven plates have worn sufficiently, the
increased release lever movement will rotate the
worm which in turn will partially screw round the
adjusting ring to compensate and so reset the position of the release levers.
Fig. 2.15
Clutch upshift brake (torque limiting)
Fig. 2.16 Multiplate hydraulically actuated clutches
travel. Therefore, the pedal must be fully depressed
to squeeze the clutch brake. The clutch pedal should
never be fully depressed before the gearbox is put
into neutral. If the clutch brake is applied with the
gearbox still in gear, a reverse load will be put on the
gears making it difficult to get the gearbox out of
gear. At the same time it will have the effect of trying
to stop or decelerate the vehicle with the clutch brake
and rapid wear of the friction disc will take place.
Never apply the clutch brake when making down
shifts, that is do not fully depress the clutch pedal
when changing from a higher to a lower gear.
up with the input and output splined drive members
respectively (Fig. 2.16). When these plates are
squeezed together, torque will be transmitted from
the input to the output members by way of these
splines and grooves and the friction torque generated between pairs of rubbing surfaces. These steel
plates are faced with either resinated paper linings
or with sintered bronze linings, depending whether
moderate or large torques are to be transmitted.
Because the whole gear cluster assembly will be
submerged in fluid, these linings are designed to
operate wet (in fluid). These clutches are hydraulically operated by servo pistons either directly or
indirectly through a lever disc spring to multiplate,
the clamping load which also acts as a piston return
spring. In this example of multiplate clutch utilization hydraulic fluid is supplied under pressure
through radial and axial passages drilled in the output shaft. To transmit pressurized fluid from one
member to another where there is relative angular
movement between components, the output shaft
has machined grooves on either side of all the radial
supply passages. Square sectioned nylon sealing
rings are then pressed into these grooves so that
2.11 Multiplate hydraulically operated automatic
transmission clutches (Fig. 2.16)
Automatic transmissions use multiplate clutches in
addition to band brakes extensively with epicyclic
compound gear trains to lock different stages of the
gearing or gear carriers together, thereby providing
a combination of gear ratios.
These clutches are comprised of a pack of annular
discs or plates, alternative plates being internally
and externally circumferentially grooved to match
when the shaft is in position, these rings expand and
seal lengthwise portions of the shaft with their corresponding bore formed in the outer members.
lower speed range, the necessary extra clamping
thrust being supplemented by the centrifugal force
at higher speeds.
The release levers are made with offset bob
weights at their outer ends, so that they are centrifugally out of balance (Fig. 2.17). The movement
due to the centrifugal force about the fixed pivot
tends to force the pressure plate against the driven
plate, thereby adding to the clamping load. While
the thrust due to the clamping springs is constant,
the movement due to the centrifugal force varies as
the square of the speed (Fig. 2.18). The reserve
factor for the thrust spring can be reduced to 1.1
compared to 1.4±1.5 for a conventional helical coil
spring clutch unit. Conversely, this clutch design
may be used for heavy duty applications where
greater torque loads are transmitted.
Front clutch (FC)
When pressurized, fluid is supplied to the front
clutch piston chamber. The piston will move over
to the right and, through the leverage of the disc
spring, will clamp the plates together with considerable thrust. The primary sun gear will now be
locked to the input turbine shaft and permit torque
to be transmitted from the input turbine shaft to
the central output shaft and primary sun gear.
Rear clutch (RC)
When pressurized, fluid is released from the front
clutch piston chamber, and is transferred to the
rear clutch piston chamber. The servo piston will
be forced directly against the end plate of the rear
clutch multiplate pack. This compresses the release
spring and sandwiches the drive and driven plates
together so that the secondary sun gear will now be
locked to the input turbine shaft. Torque can now
be transmitted from the input turbine shaft to the
secondary sun gear.
2.13 Fully automatic centrifugal clutch
(Figs 2.19 and 2.20)
Fully automatic centrifugal clutches separate
the engine from the transmission system when the
engine is stopped or idling and smoothly take up
the drive with a progressive reduction in slip within
a narrow rising speed range until sufficient engine
power is developed to propel the vehicle directly.
Above this speed full clutch engagement is
To facilitate gear changes whilst the vehicle
is in motion, a conventional clutch release
2.12 Semicentrifugal clutch (Figs 2.17 and 2.18)
With this design of clutch lighter pressure plate
springs are used for a given torque carrying capacity, making it easier to engage the clutch in the
Fig. 2.17
Semicentrifugal clutch
Fig. 2.18 Semicentrifugal clutch characteristics
lever arrangement is additionally provided. This
mechanism enables the driver to disengage and
engage the clutch independently of the flyweight
action so that the drive and driven gearbox member
speeds can be rapidly and smoothly unified during
the gear selection process.
The automatic centrifugal mechanism consists of
a reaction plate situated in between the pressure
plate and cover pressing. Mounted on this reaction
plate by pivot pins are four equally spaced bobweights (Fig. 2.19). When the engine's speed
increases, the bobweight will tend to fly outward.
Since the centre of gravity of their masses is to one
side of these pins, they will rotate about their pins.
This will be partially prevented by short struts
offset to the pivot pins which relay this movement
and effort to the pressure plate. Simultaneously,
the reaction to this axial clamping thrust causes
the reaction plate to compress both the reaction
and pressure springs so that it moves backwards
towards the cover pressing.
The greater the centrifugal force which tends to
rotate the bobweights, the more compressed the
springs will become and their reaction thrust will
be larger, which will increase the pressure plate
clamping load.
To obtain the best pressure plate thrust to engine
speed characteristics (Fig. 2.20), adjustable reactor
springs are incorporated to counteract the main
compression spring reaction. The initial compression length and therefore loading of these springs is
set up by the adjusting nut after the whole unit has
been assembled. Thus the resultant thrust of both
lots of springs determine the actual take-up engine
speed of the clutch.
Gear changes are made when the clutch is disengaged, which is achieved by moving the release
bearing forwards. This movement pulls the reactor
plate rearwards by means of the knife-edge link and
also withdraws the pressure plate through the
retractor springs so as to release the pressure plate
clamping load.
2.14 Clutch pedal actuating mechanisms
Some unusual ways of operating a clutch unit will
now be described and explained.
2.14.1 Clutch pedal with over-centre spring
(Fig. 2.21)
With this clutch pedal arrangement, the overcentre spring supplements the foot pressure applied
when disengaging the clutch, right up until the
diaphragm spring clutch is fully disengaged (Fig.
2.21). It also holds the clutch pedal in the `off'
position. When the clutch pedal is pressed, the
master cylinder piston forces the brake fluid into
the slave cylinder. The slave piston moves the push
rod, which in turn disengages the clutch. After the
pedal has been depressed approximately 25 mm of
its travel, the over-centre spring change over point
has been passed. The over-centre spring tension is
then applied as an assistance to the foot pressure.
Adjustment of the clutch is carried out by adjusting the pedal position on the master cylinder push
2.14.2 Clutch cable linkage with automatic
adjuster (Fig. 2.22)
The release bearing is of the ball race type and is
kept in constant contact with the fingers of the
diaphragm spring by the action of the pedal selfadjustment mechanism. In consequence, there is
no pedal free movement adjustment required
(Fig. 2.22).
Fig. 2.19
Fully automatic centrifugal clutch
Fig. 2.20 Fully automatic centrifugal clutch characteristics
Fig. 2.21 Hydraulic clutch operating system with over-centre spring
When the pedal is released, the adjustment pawl
is no longer engaged with the teeth on the pedal
quadrant. The cable, however, is tensioned by the
spring which is located between the pedal and
quadrant. As the pedal is depressed, the pawl
engages in the nearest vee between the teeth. The
particular tooth engagement position will gradually change as the components move to compensate
for wear in the clutch driven plate and stretch in the
whenever the driver depresses the clutch pedal
or maintains the pedal in a partially depressed
position, as may be necessary under pull-away
conditions. Movement of the clutch pedal is immediately relayed by way of the servo to the clutch in
proportion to the input pedal travel.
As the clutch's driven plate wears, clutch actuating linkage movement is automatically taken up
by the air piston moving further into the cylinder.
Thus the actual servo movement when the clutch is
being engaged and disengaged remains approximately constant. In the event of any interruption
of the air supply to the servo the clutch will still
operate, but without any servo assistance.
Immediately the clutch pedal is pushed down,
the fluid from the master cylinder is displaced into
2.14.3 Clutch air/hydraulic servo (Fig. 2.23)
In certain applications, to reduce the driver's foot
effort in operating the clutch pedal, a clutch air/
hydraulic servo may be incorporated into the actuating linkage. This unit provides power assistance
Fig. 2.22
Clutch cable linkage with automatic adjuster
the servo hydraulic cylinder. The pressure created
will act on both the hydraulic piston and the reaction plunger. Subsequently, both the hydraulic
piston and the reaction plunger move to the right
and allow the exhaust valve to close and the inlet
valve to open. Compressed air will now pass
through the inlet valve port and the passage connecting the reaction plunger chamber to the compressed air piston cylinder. It thereby applies
pressure against the air piston. The combination
of both hydraulic and air pressure on the hydraulic
and air piston assembly causes it to move over, this
movement being transferred to the clutch release
bearing which moves the clutch operating mechanism to the disengaged position (Fig. 2.23(d)).
When the clutch pedal is held partially
depressed, the air acting on the right hand side of
the reaction plunger moves it slightly to the left
which now closes the inlet valve. In this situation,
further air is prevented from entering the air
cylinder. Therefore, since no air can move in or
out of the servo air cylinder and both valves are
in the lapped position (both seated), the push rod
will not move unless the clutch pedal is again
moved (Fig. 2.23(c)).
When the clutch pedal is released fluid returns
from the servo to the master cylinder. This permits
the reaction plunger to move completely to the left
and so opens the exhaust valve. Compressed air
in the air cylinder will now transfer to the reaction
plunger chamber. It then passes through the
exhaust valve and port where it escapes to
the atmosphere. The released compressed air from
the cylinder allows the clutch linkage return spring
to move the air and hydraulic piston assembly back
to its original position in its cylinder and at the
same time this movement will be relayed to the
clutch release bearing, whereby the clutch operating mechanism moves to the engaged drive position
(Fig. 2.23(a)).
2.15 Composite flywheel and integral single plate
diaphragm clutch (Fig. 2.24)
This is a compact diaphragm clutch unit built as
an integral part of the two piece flywheel. It is
designed for transaxle transmission application
where space is at a premium and maximum torque
transmitting capacity is essential.
The flywheel and clutch drive pressing acts as a
support for the annular flywheel mass and functions as the clutch pressure plate drive member.
The advantage of having the flywheel as a two
piece assembly is that its mass can be concentrated
more effectively at its outer periphery so that its
overall mass can be reduced for the same cyclic
torque and speed fluction which it regulates.
Fig. 2.23 (a±c) Clutch air/hydraulic servo
Fig. 2.24
Integral single plate clutch and composite flywheel
The diaphragm spring takes the shape of a
dished annular disc. The inner portion of the disc
is radially slotted, the outer ends being enlarged
with a circular hole to prevent stress concentration
when the spring is deflected during disengagement.
These radial slots divide the disc into many
inwardly pointing fingers which have two functions, firstly to provide the pressure plate with
an evenly distributed multileaf spring type thrust,
and secondly to act as release levers to separate
the driven plate from the sandwiching flywheel
and pressure plate friction faces.
To actuate the clutch release, the diaphragm
spring is made to pivot between a pivot spring
positioned inside the flywheel/clutch drive pressing
near its outer periphery and a raised circumferential rim formed on the back of the pressure plate.
The engagement and release action of the clutch is
similar to the pull type diaphragm clutch where the
diaphragm is distorted into a dished disc when
assembled and therefore applies on axial thrust
between the pressure plate and its adjacent
flywheel/clutch drive pressing. With this spring
leverage arrangement, a larger pressure plate and
diaphragm spring can be utilised for a given overall
diameter of clutch assembly. This design therefore
has the benefits of lower pedal effort, higher transmitting torque capacity, a highly progressive
engagement take-up and increased clutch life compared to the conventional push type diaphragm
The engagement and release mechanism consists
of a push rod which passes through the hollow
gearbox input shaft and is made to enter and contact the blind end of a recess formed in the release
plunger. The plunger is a sliding fit in the normal
spigot bearing hole made in the crankshaft end
flange. It therefore guides the push rod and transfers its thrust to the diaphragm spring fingers via
the release plate.
Manual gearboxes and overdrives
3.1 The necessity for a gearbox
Power from a petrol or diesel reciprocating engine
transfers its power in the form of torque and angular
speed to the propelling wheels of the vehicle to
produce motion. The object of the gearbox is to
enable the engine's turning effect and its rotational
speed output to be adjusted by choosing a range of
under- and overdrive gear ratios so that the vehicle
responds to the driver's requirements within the
limits of the various road conditions. An insight
of the forces opposing vehicle motion and engine
performance characteristics which provide the
background to the need for a wide range of gearbox
designs used for different vehicle applications will
now be considered.
3.1.1 Resistance to vehicle motion
To keep a vehicle moving, the engine has to develop
sufficient power to overcome the opposing road
resistance power, and to pull away from a standstill
or to accelerate a reserve of power in addition to that
absorbed by the road resistance must be available
when required.
Road resistance is expressed as tractive resistance
(kN). The propelling thrust at the tyre to road
interface needed to overcome this resistance is
known as tractive effect (kN) (Fig. 3.1). For matching engine power output capacity to the opposing
road resistance it is sometimes more convenient to
express the opposing resistance to motion in terms
of road resistance power.
The road resistance opposing the motion of the
vehicle is made up of three components as follows:
Fig. 3.1 Vehicle tractive resistance and effort
performance chart
more energy as the wheel speed increases and therefore the rolling resistance will also rise slightly as
shown in Fig. 3.1. Factors which influence the
magnitude of the rolling resistance are the laden
weight of the vehicle, type of road surface, and
the design, construction and materials used in the
manufacture of the tyre.
1 Rolling resistance
2 Air resistance
3 Gradient resistance
Air resistance (Fig. 3.1) Power is needed to
counteract the tractive resistance created by the
vehicle moving through the air. This is caused by
air being pushed aside and the formation of turbulence over the contour of the vehicle's body. It has
been found that the air resistance opposing force
and air resistance power increase with the square
and cube of the vehicle's speed respectively. Thus at
very low vehicle speeds air resistance is insignificant, but it becomes predominant in the upper
Rolling resistance (Fig. 3.1) Power has to be
expended to overcome the restraining forces caused
by the deformation of tyres and road surfaces and
the interaction of frictional scrub when tractive
effect is applied. Secondary causes of rolling resistance are wheel bearing, oil seal friction and the
churning of the oil in the transmission system. It
has been found that the flattening distortion of the
tyre casing at the road surface interface consumes
speed range. Influencing factors which determine
the amount of air resistance are frontal area of
vehicle, vehicle speed, shape and streamlining of
body and the wind speed and direction.
gear with a small surplus of about 0.2% gradeability.
The two extreme operating conditions just
described set the highest and lowest gear ratios.
To fix these conditions, the ratio of road speed in
highest gear to road speed in lowest gear at a given
engine speed should be known. This quantity is
referred to as the ratio span.
Gradient resistance (Fig. 3.1) Power is required
to propel a vehicle and its load not only along a
level road but also up any gradient likely to be
encountered. Therefore, a reserve of power must be
available when climbing to overcome the potential
energy produced by the weight of the vehicle as it
is progressively lifted. The gradient resistance
opposing motion, and therefore the tractive effect
or power needed to drive the vehicle forward, is
directly proportional to the laden weight of the
vehicle and the magnitude of gradient. Thus driving
up a slope of 1 in 5 would require twice the reserve of
power to that needed to propel the same vehicle up a
gradient of 1 in 10 at the same speed (Fig. 3.1).
Road speed in highest gear
Road speed in lowest gear
(both road speeds being achieved at similar engine
Car and light van gearboxes have ratio spans of
about 3.5:1 if top gear is direct, but with overdrive
this may be increased to about 4.5:1. Large commercial vehicles which have a low power to weight
ratio, and therefore have very little surplus power
when fully laden, require ratio spans of between 7.5
and 10:1, or even larger for special applications.
An example of the significance of ratio span is
shown as follows:
3.1.2 Power to weight ratio
When choosing the lowest and highest gearbox
gear ratios, the most important factor to consider
is not just the available engine power but also the
weight of the vehicle and any load it is expected to
propel. Consequently, the power developed per
unit weight of laden vehicle has to be known. This
is usually expressed as the power to weight ratio.
Ratio span ˆ
Calculate the ratio span for both a car and heavy
commercial vehicle from the data provided.
Brake power developed
Power to weight
Laden weight of vehicle
There is a vast difference between the power to
weight ratio for cars and commercial vehicles
which is shown in the following examples.
Determine the power to weight ratio for the
following modes of transport:
Type of vehicle
vehicle (CV)
ˆ 4:0:1
Commercial vehicle ratio span ˆ
ˆ 8:0:1
Car ratio span ˆ
a) A car fully laden with passengers and luggage
weighs 1.2 tonne and the maximum power produced by the engine amounts to 120 kW.
b) A fully laden articulated truck weighs 38 tonne
and a 290 kW engine is used to propel this load.
ˆ 100 kW/tonne
a) Power to weight ratio ˆ
b) Power to weight ratio ˆ
ˆ 7:6 kW/tonne.
3.1.4 Engine torque rise and speed operating
range (Fig. 3.2)
Commercial vehicle engines used to pull large loads
are normally designed to have a positive torque
rise curve, that is from maximum speed to peak
torque with reducing engine speed the available
torque increases (Fig. 3.2). The amount of engine
torque rise is normally expressed as a percentage of
the peak torque from maximum speed (rated
power) back to peak torque.
3.1.3 Ratio span
Another major consideration when selecting gear
ratios is deciding upon the steepest gradient the
vehicle is expected to climb (this may normally be
taken as 20%, that is 1 in 5) and the maximum level
road speed the vehicle is expected to reach in top
% torque rise ˆ
Maximum speed torque
Peak torque
Fig. 3.2 Engine performance and gear split chart for an eight speed gearbox
The torque rise can be shaped depending upon
engine design and taking into account such features
as naturally aspirated, resonant induction tuned,
turbocharged, turbocharged with intercooling and
so forth. Torque rises can vary from as little as 5 to
as high as 50%, but the most common values for
torque rise range from 15 to 30%.
A large torque rise characteristic raises the
engine's operating ability to overcome increased
loads if the engine's speed is pulled down caused
by changes in the road conditions, such as climbing
steeper gradients, and so tends to restore the original running conditions. If the torque rise is small
it cannot help as a buffer to supplement the high
torque demands and the engine speed will rapidly
fade. Frequent gear changes therefore become
necessary compared to engines operating with
high torque rise characteristics. Once the engine
speed falls below peak torque, the torque rise
becomes negative and the pulling ability of the
engine drops off very quickly.
Vehicle driving technique should be such that
engines are continuously driven between the speed
range of peak torque and governed speed. The
driver can either choose to operate the engine's
speed in a range varying just below the maximum
rated power to achieve maximum performance and
journey speed or, to improve fuel economy, wear
and noise, within a speed range of between 200 to
400 rev/min on the positive torque rise side of the
engine torque curve that is in a narrow speed band
just beyond peak torque. Fig. 3.2 shows that the
economy speed range operates with the specific fuel
consumption at its minimum and that the engine
speed band is in the most effective pulling zone.
3.2 Five speed and reverse synchromesh gearboxes
With even wider engine speed ranges (1000 to 6000
rev/min) higher car speeds (160 km/h and more)
and high speed motorways, it has become desirable,
and in some cases essential, to increase the number
of traditional four speed ratios to five, where the
fifth gear, and sometimes also the fourth gear, is an
overdrive ratio. The advantages of increasing the
number of ratio steps are several; not only does
the extra gear provide better acceleration response,
but it enables the maximum engine rotational speed
to be reduced whilst in top gear cruising, fuel
Lubrication to the mainshaft gears is obtained by
radial branch holes which feed the rubbing surfaces
of both mainshaft and gears.
Table 3.1 Typical four and five speed gearbox gear
Five speed box
Four speed box
3.2.2 Five speed and reverse single stage
synchromesh gearbox (Fig. 3.4)
This two shaft gearbox has only one gear reduction
stage formed between pairs of different sized constant mesh gear wheels to provide a range of gear
ratios. Since only one pair of gears mesh, compared
to the two pairs necessary for the double stage
gearbox, frictional losses are halved.
Power delivered to the input primary shaft can
follow five different flow paths to the secondary
shaft via first, second, third, fourth and fifth gear
wheel pairs, but only one pair is permitted to transfer the drive from one shaft to another at any one
time (Fig. 3.4).
The conventional double stage gearbox is
designed with an input and output drive at either
end of the box but a more convenient and compact
arrangement with transaxle units where the final
drive is integral to the gearbox is to have the input
and output power flow provided at one end only of
the gearbox.
In the neutral position, first and second output
gear wheels will be driven by the corresponding
gear wheels attached to the input primary shaft,
but they will only be able to revolve about their
own axis relative to the output secondary shaft.
Third, fourth and fifth gear wheel pairs are driven
by the output second shaft and are free to revolve
only relative to the input primary shaft because
they are not attached to this shaft but use it only
as a supporting axis.
When selecting individual gear ratios, the appropriate synchronizing sliding sleeve is pushed
towards and over the dog teeth forming part of
the particular gear wheel required. Thus with first
and second gear ratios, the power flow passes from
the input primary shaft and constant mesh pairs of
gears to the output secondary shaft via the first and
second drive hub attached to this shaft. Gear
engagement is completed by the synchronizing
sleeve locking the selected output gear wheel to
the output secondary shaft. Third, fourth and
fifth gear ratios are selected when the third and
fourth or fifth gear drive hub, fixed to the input
primary shaft, is locked to the respective gear wheel
dog clutch by sliding the synchronizing sleeve in to
mesh with it. The power flow path is now transferred from the input primary shaft drive hub and
selected pair of constant mesh gears to the output
secondary shaft.
consumption is improved and engine noise and wear
are reduced. Typical gear ratios for both four and five
speed gearboxes are as shown in Table 3.1.
The construction and operation of four speed
gearboxes was dealt with in Vehicle and Engine
Technology. The next section deals with five speed
synchromesh gearboxes utilized for longitudinal
and transverse mounted engines.
3.2.1 Five speed and reverse double stage
synchromesh gearbox (Fig. 3.3)
With this arrangement of a five speed double stage
gearbox, the power input to the first motion shaft
passes to the layshaft and gear cluster via the first
stage pair of meshing gears. Rotary motion is
therefore conveyed to all the second stage layshaft
and mainshaft gears (Fig. 3.3). Because each pair of
second stage gears has a different size combination,
a whole range of gear ratios are provided. Each
mainshaft gear (whilst in neutral) revolves on the
mainshaft but at some relative speed to it. Therefore, to obtain output powerflow, the selected
mainshaft gear has to be locked to the mainshaft.
This then completes the flow path from the first
motion shaft, first stage gears, second stage gears
and finally to the mainshaft.
In this example the fifth gear is an overdrive gear
so that to speed up the mainshaft output relative to
the input to the first motion shaft, a large layshaft
fifth gear wheel is chosen to mesh with a much
smaller mainshaft gear.
For heavy duty operations, a forced feed lubrication system is provided by an internal gear crescent
type oil pump driven from the rear end of the
layshaft (Fig. 3.3). This pump draws oil from the
base of the gearbox casing, pressurizes it and then
forces it through a passage to the mainshaft. The
oil is then transferred to the axial hole along the
centre of the mainshaft by way of an annular
passage formed between two nylon oil seals.
Fig. 3.3 Five speed and reverse double stage synchromesh gearbox
Transference of power from the gearbox output
secondary shaft to the differential left and right
hand drive shafts is achieved via the final drive
pinion and gear wheel which also provide a permanent gear reduction (Fig. 3.4). Power then flows
from the differential cage which supports the final
drive gear wheel to the cross-pin and planet gears
where it then divides between the two side sun gears
and accordingly power passes to both stub drive
3.3 Gear synchronization and engagement
The gearbox basically consists of an input shaft
driven by the engine crankshaft by way of the
clutch and an output shaft coupled indirectly either
Fig. 3.4 Five speed and reverse single stage synchromesh gearbox with integral final drive (transaxle unit)
through the propellor shaft or intermediate gears to
the final drive. Between these two shafts are pairs of
gear wheels of different size meshed together.
If the gearbox is in neutral, only one of these
pairs of gears is actually attached rigidly to one of
these shafts while the other is free to revolve on the
second shaft at some speed determined by the existing speeds of the input and output drive shafts.
To engage any gear ratio the input shaft has to
be disengaged from the engine crankshaft via the
clutch to release the input shaft drive. It is then only
the angular momentum of the input shaft, clutch
drive plate and gear wheels which keeps them revolving. The technique of good gear changing is to be
able to judge the speeds at which the dog teeth of
both the gear wheel selected and output shaft are
rotating at a uniform speed, at which point in time
the dog clutch sleeve is pushed over so that both sets
of teeth engage and mesh gently without grating.
Because it is difficult to know exactly when to
make the gear change a device known as the synchromesh is utilized. Its function is to apply a friction clutch braking action between the engaging
gear and drive hub of the output shaft so that
their speeds will be unified before permitting the
dog teeth of both members to contact.
Synchromesh devices use a multiplate clutch or a
conical clutch to equalise the input and output
rotating members of the gearbox when the process
of gear changing is taking place. Except for special
applications, such as in some splitter and range
change auxiliary gearboxes, the conical clutch
method of synchronization is generally employed.
With the conical clutch method of producing silent
gear change, the male and female cone members
are brought together to produce a synchronizing
frictional torque of sufficient magnitude so that one
or both of the input and output members' rotational
speed or speeds adjust automatically until they
revolve as one. Once this speed uniformity has been
achieved, the end thrust applied to the dog clutch
sleeve is permitted to nudge the chamfered dog teeth
of both members into alignment, thereby enabling the
two sets of teeth to slide quietly into engagement.
ately the balls are pushed out of their groove, the
chamfered edges of the outer hub's internal teeth will
then be able to align with the corresponding teeth
spacing on the first motion gear. Both sets of teeth
will now be able to mesh so that the outer hub can be
moved into the fully engaged position (Fig. 3.5(c)).
Note the bronze female cone insert frictional face
is not smooth, but consists of a series of tramline
grooves which assist in cutting away the oil film so
that a much larger synchronizing torque will be
generated to speed up the process.
3.3.2 Positive baulk ring synchromesh unit
(Fig. 3.6(a, b and c))
The gearbox mainshaft rotates at propellor shaft
speed and, with the clutch disengaged, the first
motion shaft gear, layshaft cluster gears, and
mainshaft gears rotate freely.
Drive torque will be transmitted when a gear
wheel is positively locked to the mainshaft. This is
achieved by means of the outer synchromesh hub
internal teeth which slide over the inner synchromesh hub splines (Fig. 3.6(a)) until they engage
with dog teeth formed on the constant mesh gear
wheel being selected.
When selecting and engaging a particular gear
ratio, the gear stick slides the synchromesh outer
hub in the direction of the chosen gear (towards
the left). Because the shift plate is held radially
outwards by the two energizing ring type springs
and the raised middle hump of the plate rests in the
groove formed on the inside of the hub, the end of
the shift plate contacts the baulking ring and pushes
it towards and over the conical surface, forming
part of the constant mesh gear wheel (Fig. 3.6(b)).
The frictional grip between the male and female
conical members of the gear wheel and baulking
ring and the difference in speed will cause the baulking ring to be dragged around relative to the inner
hub and shift plate within the limits of the clearance
between the shift plate width and that of the
recessed slot in the baulking ring. Owing to the
designed width of the shift plate slot in the baulking
ring, the teeth on the baulking ring are now out of
alignment with those on the outer hub by approximately half a tooth width, so that the chamfered
faces of the teeth of the baulking ring and outer hub
bear upon each other.
As the baulking ring is in contact with the gear
cone and the outer hub, the force exerted by the
driver on the gear stick presses the baulking ring
female cone hard against the male cone of the gear.
Frictional torque between the two surfaces will
eventually cause these two members to equalize
3.3.1 Non-positive constant load synchromesh
unit (Fig. 3.5(a, b and c))
When the gear stick is in the neutral position the
spring loaded balls trapped between the inner and
outer hub are seated in the circumferential groove
formed across the middle of the internal dog teeth
(Fig. 3.5(a)). As the driver begins to shift the gear
stick into say top gear (towards the left), the outer
and inner synchromesh hubs move as one due to the
radial spring loading of the balls along the splines
formed on the main shaft until the female cone of the
outer hub contacts the male cone of the first motion
gear (Fig. 3.5(b)). When the pair of conical faces
contact, frictional torque will be generated due to
the combination of the axial thrust and the difference in relative speed of both input and output shaft
members. If sufficient axial thrust is applied to the
outer hub, the balls will be depressed inwards
against the radial loading of the springs. Immedi66
Fig. 3.5 Non-positive constant load synchromesh unit
Fig. 3.6 (a±c)
Positive baulk ring synchromesh unit
their speeds. Until this takes place, full engagement
of the gear and outer hub dog teeth is prevented by
the out of alignment position of the baulking ring
teeth. When the gear wheel and main shaft have
unified their speeds, the synchronizing torque will
have fallen to zero so that the baulking ring is no
longer dragged out of alignment. Therefore the
outer hub can now overcome the baulk and follow
through to make a positive engagement between
hub and gear (Fig. 3.6(c)). It should be understood
that the function of the shift plate and springs is to
transmit just sufficient axial load to ensure a rapid
bringing together of the mating cones so that the
baulking ring dog teeth immediately misalign with
their corresponding outer hub teeth. Once the cone
faces contact, they generate their own friction
torque which is sufficient to flick the baulking
ring over, relative to the outer hub. Thus the chamfers of both sets of teeth contact and oppose further
outer hub axial movement towards the gear dog
centre, hold the bronze synchronizing cone rings
apart. Alternating with the shouldered pins on the
same pitch circle are diametrically split pins, the
ends of which fit into blind bores machined in
the synchronizing cone rings. The pin halves are
sprung apart, so that a chamfered groove around the
middle of each half pin registers with a chamfered
hole in the drive hub.
If the gearbox is in the neutral position, both sets
of shouldered and split pins are situated with their
grooves aligned with the central drive hub (Fig.
3.8(a and b)).
When an axial load is applied to the drive hub by
the gear stick, it moves over (in this case to the left)
until the synchronizing ring is forced against the
adjacent first motion gear cone. The friction (synchronizing) torque generated between the rubbing
tapered surfaces drags the bronze synchronizing
ring relative to the mainshaft and drive hub until
the grooves in the shouldered pins are wedged against
the chamfered edges of the drive hub (Fig. 3.8(c)) so
that further axial movement is baulked.
Immediately the input and output shaft speeds
are similar, that is, synchronization has been
achieved, the springs in the split pins are able to
expand and centralize the shouldered pins relative
to the chamfered holes in the drive hub. The drive
hub can now ride out of the grooves formed around
the split pins, thus permitting the drive hub to shift
further over until the internal and external dog
teeth of both gear wheel hub mesh and fully engage
(Fig. 3.8(d)).
3.3.3 Positive baulk pin synchromesh unit
(Fig. 3.7(a, b, c and d))
Movement of the selector fork synchronizing sleeve
to the left (Fig. 3.7(a and b)) forces the female
(internal) cone to move into contact with the male
(external) cone on the drive gear. Frictional torque
will then synchronize (unify) the input and output
speeds. Until speed equalization is achieved, the collars on the three thrust pins (only one shown) will be
pressed hard into the enlarged position of the slots
(Fig. 3.5(c)) in the synchronizing sleeve owing to the
frictional drag when the speeds are dissimilar. Under
these conditions, unless extreme pressure is exerted,
the dog teeth cannot be crushed by forcing the collars
into the narrow portion of the slots. However, when
the speeds of the synchromesh hub and drive gear are
equal (synchronized) the collars tend to `float' in
the enlarged portion of the slots, there is only the
pressure of the spring loaded balls to be overcome.
The collars will then slide easily into the narrow
portion of the slots (Fig. 3.5(d)) allowing the synchronizer hub dog teeth to shift in to mesh with the
dog teeth on the driving gear.
3.3.5 Split ring synchromesh unit
(Fig. 3.9(a, b, c and d))
In the neutral position the sliding sleeve sits centrally over the drive hub (Fig. 3.9(a)). This permits
the synchronizing ring expander band and thrust
block to float within the constraints of the recess
machine in the side of the gear facing the drive hub
(Fig. 3.9(b)).
For gear engagement to take place, the sliding
sleeve is moved towards the gear wheel selected (to
the left) until the inside chamfer of the sliding sleeve
contacts the bevelled portion of the synchronizing
ring. As a result, the synchronizing ring will be
slightly compressed and the friction generated
between the two members then drags the synchronizing ring round in the direction of whichever
member is rotating fastest, be it the gear or driven
hub. At the same time, the thrust block is pulled
round so that it applies a load to one end of the
expander band, whilst the other end is restrained
from moving by the anchor block (Fig. 3.9(c)).
3.3.4 Split baulk pin synchromesh unit
(Fig. 3.8(a, b, c and d))
The synchronizing assembly is composed of two
thick bronze synchronizing rings with tapered
female conical bores, and situated between them
is a hardened steel drive hub internally splined with
external dog teeth at each end (Fig. 3.8(a)). Three
shouldered pins, each with a groove around its
Fig. 3.7 (a±d) Positive baulk pin synchromesh unit
Whilst this is happening the expander is also
pushed radially outwards. Consequently, there
will be a tendency to expand the synchronizing
slit ring, but this will be opposed by the chamfered
mouth of the sliding sleeve. This energizing action
attempting to expand the synchronizing ring prevents the sliding sleeve from completely moving
over and engaging the dog teeth of the selected
Fig. 3.8 (a±c) Split baulk pin synchromesh unit
gear wheel until both the drive hub and constant
mesh gear wheel are revolving at the same speed.
When both input and output members are unified, that is, rotating as one, there cannot be any
more friction torque because there is no relative
speed to create the frictional drag. Therefore
the expander band immediately stops exerting
radial force on the inside of the synchronizing ring.
Fig. 3.9 (a±d) Split baulk ring synchromesh unit
The axial thrust applied by the gear stick to the
sliding sleeve will now be sufficient to compress the
split synchronizing ring and subsequently permits
the sleeve to slide over the gear wheel dog teeth for
full engagement (Fig. 3.9(d)).
3.4 Remote controlled gear selection and
engagement mechanisms
Gear selection and engagement is achieved by two
distinct movements:
1 The selection of the required gear shift gate and
the positioning of the engagement gate lever.
2 The shifting of the chosen selector gate rod into
the engagement gear position.
These two operations are generally performed
through the media of the gear shift lever and the
remote control tube/rod. Any transverse movement of the gear shift lever by the driver selects
the gear shift gate and the engagement of the gate
is obtained by longitudinal movement of the gear
shift lever.
Movement of the gear shift lever is conveyed to
the selection mechanism via the remote control
tube. Initially the tube is twisted to select the
gate shift gate, followed by either a push or pull
movement of the tube to engage the appropriate
For the gear shift control to be effective it must
have some sort of flexible linkage between the gear
shift lever supported on the floor of the driver's
compartment and the engine and transmission integral unit which is suspended on rubber mountings.
This is essential to prevent engine and transmission
vibrations being transmitted back to the body and
floor pan and subsequently causing discomfort to
the driver and passengers.
Fig. 3.10 Remote controlled double rod and bell crank
lever gearshift mechanism suitable for both four and
five speed transversely mounted gearbox
ment shaft. Consequently, this shifts the transverse
selector/engagement shaft so that it pushes the
synchronizing sliding sleeve into engagement with
the selected gear dog teeth.
3.4.2 Remote controlled bell cranked lever gear
shift mechanism for a four speed transverse
mounted gearbox (Ford) (Fig. 3.11)
Gear selection and engagement movement is
conveyed from the gear shift lever pivot action to
the remote control rod universal joint and to the
control shift and relay lever guide (Fig. 3.11).
Rocking the gear shift lever transversely rotates
the control shaft and relay guide. This tilts the
selector relay lever and subsequently the selection relay lever guide and shaft until the striker
finger aligns with the chosen selector gate. A further push or pull movement to the gear shift lever
by the driver then transfers a forward or
backward motion via the remote control rod, control shaft and relay lever guide to the engagement
relay lever. Movement is then redirected at right
angles to the selector relay guide and shaft.
Engagement of the gear required is finally obtained
by the selector/engagement shaft forcing the striking finger to shift the gate and selector fork along
the single selector rod so that the synchronizing sleeve meshes with the appropriate gear
wheel dog clutch.
3.4.1 Remote controlled double rod and bell
cranked lever gear shift mechanism, suitable for
both four and five speed transverse mounted
gearbox (Talbot) (Fig. 3.10)
Twisting the remote control tube transfers movement to the first selector link rod. This motion is
then redirected at right angles to the transverse
gate selector/engagement shaft via the selector
relay lever (bell crank) to position the required
gear gate (Fig. 3.10). A forward or backward
movement of the remote control tube now conveys
motion via the first engagement relay lever (bell
crank), engagement link rod and second relay
lever to rotate the transverse gate selector/engage73
Fig. 3.11 Remote controlled bell crank level gear shift mechanism for a four speed transversely mounted gearbox
3.4.3 Remote controlled sliding ball joint gear
shift mechanism suitable for both four and five
speed longitudinal or transverse mounted gearbox
(VW) (Fig. 3.12)
Selection and engagement of the different gear
ratios is achieved with a swivel ball end pivot gear
shift lever actuating through a sliding ball relay
lever a single remote control rod (Fig. 3.12). The
remote control rod transfers both rotary and pushpull movement to the gate selector and engagement
shaft. This rod is also restrained in bushes between
the gear shift lever mounting and the bulkhead.
It thus permits the remote control rod to transfer
both rotary (gate selection) and push-pull (select rod
engagement shift) movement to the gate selector and
engagement shaft. Relative movement between the
suspended engine and transmission unit and the car
body is compensated by the second sliding ball
relay lever. As a result the gate engagement striking
finger is able to select and shift into engagement the
appropriate selector rod fork.
This single rod sliding ball remote control
linkage can be used with either longitudinally or
transversely mounted gearboxes, but with the latter
Fig. 3.12 Remote controlled sliding ball joint gear shift
mechanism suitable for both four and five speed
longitudinally or transversely mounted gearbox
an additional relay lever mechanism (not shown) is
needed to convey the two distinct movements of
selection and engagement through a right angle.
3.4.4 Remote controlled double rod and hinged
relay joint gear shift mechanism suitable for both
four and five speed longitudinal mounted gearbox
(VW) (Fig. 3.13)
With this layout the remote control is provided by
a pair of remote control rods, one twists and selects
the gear gate when the gear shift lever is given a
transverse movement, while the other pushes or
pulls when the gear shift lever is moved longitudinally (Fig. 3.13). Twisting movement is thus conveyed to the engagement relay lever which makes
the engagement striking finger push the aligned
selector gate and rod. Subsequently, the synchronizing sleeve splines mesh with the corresponding
dog clutch teeth of the selected gear wheel. Relative
movement between the gear shift lever swivel support and rubber mounted gearbox is absorbed by
the hinged relay joint and the ball joints at either
end of the remote control engagement rod.
3.4.5 Remote controlled single rod with self
aligning bearing gear shift mechanism suitable for
both five and six speed longitudinal mounted
gearbox (Ford) (Fig. 3.14)
A simple and effective method of selecting and
engaging the various gear ratios suitable for
commercial vehicles where the driver cab is forward of the gearbox is shown in Fig. 3.14.
Fig. 3.14 Remote controlled single rod with self-aligning
bearing gear shift mechanism suitable for both five
and six speed longitudinally mounted gearbox
Movement of the gear shift lever in the usual
transverse and longitudinal directions provides
both rotation and a push-pull action to the remote
control tube. Twisting the remote control tube
transversely tilts the relay gear shift lever about its
ball joint so that the striking finger at its lower
end matches up with the selected gear gate. Gear
engagement is then completed by the driver tilting
the gear shift lever away or towards himself. This
permits the remote control tube to move axially
through the mounted self-aligning bearing. As a
result, a similar motion will be experienced by the
relay gear shift lever, which then pushes the striking
finger, selector gate and selector fork into the gear
engaged position. It should be observed that the
self-aligning bearing allows the remote control tube
to slide to and fro. At the same time it permits the
inner race member to tilt if any relative movement
between the gearbox and chassis takes place.
3.4.6 Remote controlled single rod with swing
arm support gear shift mechanism suitable for five
and six speed longitudinally mounted gearbox
(ZF) (Fig. 3.15)
This arrangement which is used extensively on
commercial vehicles employs a universal crosspin joint to transfer both the gear selection and
Fig. 3.13 Remote controlled double rod and hinged
relay joint gear shift mechanism suitable for both four and
five speed longitudinally mounted gearbox
pivots the suspended selector gate relay lever so
that the transverse gate selector/engagement shift
moves across the selector gates until it aligns with
the selected gate. The gear shift lever is then given a
to or fro movement. This causes the transverse
selector/engagement shaft to rotate, thereby forcing the striking finger to move the selector rod
and fork. The synchronizing sleeve will now be
able to engage the dog clutch of the appropriate
gear wheel. Any misalignment between the gear
shift lever support mounting and the gear shift
mechanism forming part of the gearbox (caused
by engine shake or rock) is thus compensated by
the swing rod which provides a degree of float for
the selector gate relay lever pivot point.
3.5 Splitter and range change gearboxes
Ideally the tractive effect produced by an engine and
transmission system developing a constant power
output from rest to its maximum road speed would
vary inversely with its speed. This characteristic can
be shown as a smooth declining tractive effect curve
with rising road speed (Fig. 3.16).
In practice, the transmission has a fixed number
of gear ratios so that the ideal smooth tractive
effect curve would be interrupted to allow for loss
Fig. 3.15 Remote controlled single rod with swing arm
support gear shift mechanism suitable for five and six
speed longitudinally mounted gearbox
engagement motion to the remote control tube
(Fig. 3.15). Twisting this remote control tube by
giving the gear shift lever a transverse movement
Fig. 3.16 Ideal and actual tractive effort-speed characteristics of a vehicle
of engine speed and power between each gear
change (see the thick lines of Fig. 3.16).
For a vehicle such as a saloon car or light van
which only weighs about one tonne and has a large
power to weight ratio, a four or five speed gearbox
is adequate to maintain tractive effect without too
much loss in engine speed and vehicle performance
between gear changes.
Unfortunately, this is not the situation for heavy
goods vehicles where large loads are being hauled
so that the power to weight ratio is usually very
low. Under such operating conditions if the gear
ratio steps are too large the engine speed will drop
to such an extent during gear changes that the
engine torque recovery will be very sluggish
(Fig. 3.17). Therefore, to minimize engine speed
fall-off whilst changing gears, smaller gear ratio
steps are required, that is, more gear ratios are
necessary to respond to the slightest change in
vehicle load, road conditions and the driver's
requirements. Figs 3.2 and 3.18 show that by doubling the number of gear ratios, the fall in engine
speed between gear shifts is much smaller. To cope
with moderate payloads, conventional double
stage compound gearboxes with up to six forward
speeds are manufactured, but these boxes tend to be
large and heavy. Therefore, if more gear ratios are
essential for very heavy payloads, a far better way of
extending the number of gear ratios is to utilize a two
speed auxiliary gearbox in series with a three, four,
five or six speed conventional compound gearbox.
The function of this auxiliary box is to double the
number of gear ratios of the conventional gearbox.
With a three, four, five or six speed gearbox, the
gear ratios are increased to six, eight, ten or twelve
respectively (Figs 3.2 and 3.18). For very special
Fig. 3.18 Engine speed ratio chart for a vehicle
employing either a ten speed range change or a splitter
change gearbox
applications, a three speed auxiliary gearbox can be
incorporated so that the gear ratios are trebled.
Usually one of these auxiliary gear ratios provides a
range of very low gear ratios known as crawlers or
deep gears. The auxiliary gearbox may be situated
either in front or to the rear of the conventional
compound gearbox.
The combination of the auxiliary gearbox and
the main gearbox can be designed to be used either
as a splitter gear change or as a range gear change
in the following way.
3.5.1 Splitter gear change (Figs 3.19 and 3.20)
With the splitter arrangement, the main gearbox
gear ratios are spread out wide between adjacent
gears whilst the two speed auxiliary gearbox has one
direct gear ratio and a second gear which is either a
step up or down ratio (Fig. 3.19). The auxiliary
second gear ratio is chosen so that it splits the main
gearbox ratio steps in half, hence the name splitter
gear change. The splitter indirect gear ratio normally is set between 1.2 and 1.4:1. A typical ratio
would be 1.25:1.
A normal upchange sequence for an eight speed
gearbox (Fig. 3.20), consisting of a main gearbox
with four forward gear ratios and one reverse and a
two speed auxiliary splitter stage, would be as
Auxiliary splitter low ratio and main gearbox first
gear selected results in `first gear low' (1L); auxiliary
splitter switched to high ratio but with the main gearbox still in first gear results in `first gear high' (1H);
Fig. 3.17 Engine speed ratio chart for a vehicle
employing a five speed gearbox
Fig. 3.19 Eight speed constant mesh gearbox with two speed front mounted splitter change
splitter switched again to low ratio and main gearbox second gear selected results in 2L; splitter
switched to high ratio, main gearbox gear remaining
in second gives 2H; splitter switched to low ratio
main gearbox third gear selected gives 3L; splitter
switched to high ratio main gearbox still in third
gives 3 H. This procedure continues throughout
the upshift from bottom to top gear ratio. Thus the
overall upshift gear ratio change pattern would be:
Gear ratio 1 2
3 4
5 6
7 8
1L 1H 2L 2H 3L 3H 4L 4H RL
It can therefore be predicted that alternate
changes involve a simultaneous upchange in the
Fig. 3.20
Splitter change gear ratio±speed chart
main gearbox and downchange in the splitter stage,
or vice versa.
Referring to the thick lines in Figs 3.2, 3.17 and
3.18, these represent the recommended operating
speed range for the engine for best fuel economy,
but the broken lines in Fig. 3.17 represent the gear
shift technique if maximum road speed is the
criteria and fuel consumption, engine wear and
noise become secondary considerations.
Through the gear ratios from bottom to top
`low gear range' is initially selected, the main gearbox order of upchanges are first, second, third and
fourth gears. At this point the range change is
moved to `high gear range' and the sequence of
gear upchanges again becomes first, second, third
and fourth. Therefore the total number of gear
ratios is the sum of both low and high ranges,
that is, eight. A tabulated summary of the upshift
gear change pattern will be:
3.5.2 Range gear change (Figs 3.21 and 3.22)
In contrast to the splitter gear change, the range
gear change arrangement (Fig. 3.21) has the gear
ratios between adjacent gear ratio steps set close
together. The auxiliary two speed gearbox will have
one ratio direct drive and the other one normally
equal to just over half the gear ratio spread from
bottom to top. This is slightly larger than the main
gearbox gear ratio spread.
To change from one gear ratio to the next with,
for example, an eight speed gearbox comprising
four normal forward gears and one reverse and a
two speed auxiliary range change, the pattern of
gear change would be as shown in Fig. 3.22.
Fig. 3.21
Gear ratio 1 2 3 4 5
1L 2L 3L 4L 1H 2H 3H 4H RL
3.5.3 Sixteen speed synchromesh gearbox with
range change and integral splitter gears
(Fig. 3.23)
This heavy duty commercial gearbox utilizes both a
two speed range change and a two speed splitter
gear change to enable the four speed gearbox to
Eight speed constant mesh gearbox with two speed rear mounted range change
When low or high splitter gears are engaged,
the first motion shaft drive hub conveys power to
the first or second pair of splitter gear wheels and
hence to the layshaft gear cluster.
Mid-four speed gearbox power flow (Fig. 3.23)
Power from the first motion shaft at a reduced
speed is transferred to the layshaft cluster of gears
and subsequently provides the motion to all the
other mainshaft gear wheels which are free to
revolve on the mainshaft, but at relatively different
speeds when in the neutral gear position.
Engagement of one mid-gearbox gear ratio dog
clutch locks the corresponding mainshaft drive hub
to the chosen gear so that power is now able to pass
from the layshaft to the mainshaft through the
selected pair of gear wheels.
Reverse gear is provided via an idler gear which,
when meshed between the layshaft and mainshaft,
alters the direction of rotation of the mainshaft in
the usual manner.
Fig. 3.22 Range change gear ratio±speed chart
extend the gear ratio into eight steps and, when
required, to sixteen split (narrow) gear ratio
The complete gearbox unit can be considered to
be divided into three sections; the middle section
(which is basically a conventional double stage four
speed gearbox), and the first two pairs of gears at
the front end which make up the two speed splitter
gearbox. Mounted at the rear is an epicyclic gear
train providing a two speed low and high range
change (Fig. 3.23).
The epicyclic gear train at the rear doubles the
ratios of the four speed gearbox permitting the
driver to initially select the low (L) gear range
driving through this range 1, 2, 3 and 4 then selecting the high (H) gear range. The gear change
sequence is again repeated but the gear ratios now
become 5, 6, 7 and 8.
If heavy loads are being carried, or if maximum
torque is needed when overtaking on hills, much
closer gear ratio intervals are desirable. This is
provided by splitting the gear steps in half with
the two speed splitter gears; the gear shift pattern
of 1st low, 1st high, 2nd low, 2nd high, 3rd low and
so on is adopted.
Rear end range two speed gearbox power flow
(Fig. 3.23) When the range change is in the neutral position, power passes from the mainshaft and
sun gear to the planet gears which then revolve on
the output shaft's carrier pin axes and in turn spin
round the annular gear and synchronizing drive
Engaging the low range gear locks the synchronizing drive hub to the gearbox casing. This forces
the planet gears to revolve and walk round the
inside of the annular gear. Consequently, the carrier
and output shafts which support the planet gear
axes will also be made to rotate but at a speed
lower than that of the input shaft.
Changing to high range locks the annular gear
and drive hub to the output shaft so that power
flow from the planet gears is then divided between
the carrier and annular, but since they need to
rotate at differing speeds, the power flow forms a
closed loop and jams the gearing. As a result, there
is no gear reduction but just a straight through
drive to the output shaft.
Front end splitter two speed gearbox power flow
(Fig. 3.23) Input power to the gearbox is supplied
to the first motion shaft. When the splitter synchronizing sliding sleeve is in neutral, both the splitter
low and high input gear wheels revolve on their
needle bearings independently of their supporting
first motion shaft and mainshaft respectively.
3.5.4 Twin counter shaft ten speed constant mesh
gearbox with synchromesh two speed rear
mounted range change (Fig. 3.24)
With the quest for larger torque carrying capacity,
closer steps between gear ratio changes, reduced
gearbox length and weight, a unique approach
to fulfil these requirements has been developed
Fig. 3.23
Sixteen speed synchromesh with range change and integral splitter gears
Fig. 3.24 Twin countershaft ten speed constant mesh gearbox with synchromesh two speed range change
adopting the two countershaft constant mesh gearbox which incorporates a synchromesh two speed
rear mounted range change (Fig. 3.24).
The main gearbox is in fact a double stage compound conventional gearbox using two countershafts (layshaft) instead of the normal single
Power flow path Power flows into the main gearbox through the input first motion shaft and gear
wheel. Here it is divided between the two first stage
countershaft gears and is then conveyed via each
countershaft gear wheel to the corresponding second stage mainshaft gears. Each of these rotate at
relative speeds about the mainshaft. Torque is only
transmitted to the mainshaft when the selected dog
clutch drive hub is slid in to mesh with the desired
gear dog teeth.
The power flow can then pass directly to the output shaft by engaging the synchromesh high range
dog teeth. Conversely, a further gear reduction can
be made by engaging the low range synchromesh
dog teeth so that the power flow from the mainshaft
auxiliary gear is split between the two auxiliary
countershafts. The additional speed reduction is
then obtained when the split power path comes
together through the second stage auxiliary output
gear. It should be observed that, unlike the mainshaft, the auxiliary gear reduction output shaft has
no provision for radial float.
Reverse gear is obtained by incorporating an idle
gear between the second stage countershaft reverse
gears and the mainshaft reverse gear so that the
mainshaft reverse gear is made to rotate in the
opposite direction to all the other forward drive
mainshaft gears.
Design and construction Referring to Fig. 3.24,
there is a countershaft either side of the mainshaft
and they are all in the same plane. What cannot be
seen is that this single plane is inclined laterally at
19 to the horizontal to reduce the overall height of
the gearbox.
The mainshaft is hollow and is allowed to float in
the following manner: each end is counterbored,
and into each counterbore is pressed a stabilizing
rod. The front end of this rod projects into the rear
of the input shaft which is also counterbored to
house a supporting roller bearing for the stabilizer
rod. The rear projecting stabilizer rod has a spherically shaped end which rests in a hole in the centre
of a steel disc mounted inside the auxiliary drive
gear immediately behind the mainshaft. This gear
itself is carried by a ball bearing mounted in the
gearbox housing. When torque is transmitted
through the gearbox, the centrally waisted 11 mm
diameter section of both stabilizers deflects until
radial loads applied by the two countershaft gears
to the mainshaft gear are equalized. By these
means, the input torque is divided equally between
the two countershafts and two diametrically opposite teeth on the mainshaft gear at any one time.
Therefore, the face width of the gear teeth can be
reduced by about 40% compared to gearboxes using
single countershafts. Another feature of having a
mainshaft which is relatively free to float in all radial
directions is that it greatly reduces the dynamic
loads on the gear teeth caused by small errors of
tooth profile during manufacture. A maximum
radial mainshaft float of about 0.75 mm has proved
to be sufficient to permit the shaft to centralize and
distribute the input torque equally between the two
countershafts. To minimize end thrust, all the gears
have straight spur teeth which run acceptably
quietly due to the balanced loading of the gears.
Each of the five forward speeds and reverse are
engaged by dog teeth clutches machined on both
ends of the drive hubs. The ends of the external
teeth on the drive hubs and the internal teeth in the
mainshaft gears are chamfered at about 35 to
provide some self-synchronizing action before
3.6 Transfer box power take-off (PTO) (Fig. 3.25)
A power take-off (PTO) provides some shaft drive
and coupling to power specialized auxiliary equipment at a specified speed and power output. Power
take-offs (PTOs) can be driven directly from the
engine's timing gears, but it is more usual and
practical to take the drive from some point off
the gearbox. Typical power take-off applications
are drives for hydraulic pumps, compressors, generators, hoists, derricks, capstain or cable winch
platform elevators, extended ladders, hose reels,
drain cleaning vehicles, tippers, road sweepers,
snow plough blade and throwing operations and
any other mechanical mechanism that needs a
separate source of power drive output.
The power take-off can be driven either by one of
the layshaft cluster gears, so that it is known as a
side mounted transfer box, or it may be driven from
the back end of the layshaft, in which case it is
known as a rear mounted transfer box (Fig. 3.25).
Transfer boxes can either be single or two speed
arrangements depending upon the intended application. The gear ratios of the transfer box are so
chosen that output rotational speeds may be anything from 50 to 150% of the layshaft input speed.
Fig. 3.25 Six speed constant mesh gearbox illustrating different power take-off point arrangement
3.6.1 Side mounted single speed transfer box
(Fig. 3.25)
With the single speed side mounted transfer box,
the drive is conveyed to the output gear and shaft
by means of an intermediate gear mounted on a
splined idler shaft which is itself supported by two
spaced out ball bearings (Fig. 3.25). Engagement of
the transfer output shaft is obtained by sliding the
intermediate straight toothed gear into mesh with
both layshaft gear and output shaft gear by a selector fork mounted on a gear shift not shown.
ated. With this gear train layout, the drive is conveyed to the intermediate shaft by a gear wheel
which is in constant mesh with both the layshaft
gear wheel and the high speed output gear
(Fig. 3.25). The output shaft supports the high
speed output gear which is free to revolve relative
to it when the transfer drive is in neutral or low gear
is engaged. Also attached to this shaft on splines is
the low speed output gear.
High transfer gear ratio engagement is obtained
by sliding the low speed output gear towards the
high speed output gear until its internal splines
mesh with the dog teeth on the side of the gear.
This then transfers the drive from the layshaft to
the output shaft and coupling through a simple
single stage gear reduction.
3.6.2 Side mounted two speed transfer box
(Fig. 3.25)
If a more versatile transfer power take-off is
required, a two speed transfer box can be incorpor84
Low transfer gear ratio engagement occurs when
the low speed output gear is slid into mesh with
the smaller intermediate shaft gear. The power
flow then takes place through a double stage
(compound) gear reduction.
By selecting a 20% overdrive top gear, say, the
transmission gear ratios can be so chosen that the
engine and road resistance power curves coincide at
peak engine power (Fig. 3.26). The undergearing
has thus permitted the whole of the engine power
curve to be shifted nearer the opposing road resistance power curve so that slightly more engine
power is being utilized when the two curves intersect. As a result, a marginally higher maximum
vehicle speed is achieved. In other words, the
engine will be worked at a lower speed but at a
higher load factor whilst in this overdrive top gear.
If the amount of overdrive for top gear is
increased to 40%, the engine power curve will be
shifted so far over that it intersects the road resistance power curve before peak engine power has
been obtained (Fig. 3.26) and therefore the maximum possible vehicle speed cannot be reached.
Contrasting the direct drive 20% and 40%
overdrive with direct drive top gear power curves
with respect to the road resistance power curve at
70 km/h, as an example, it can be seen (Fig. 3.26)
that the reserve of power is 59%, 47% and 38%
respectively. This surplus of engine power over the
power absorbed by road resistance is a measure of
the relative acceleration ability for a particular
transmission overall gear ratio setting.
A comparison of the three engine power curves
shows that with direct drive top gear the area in the
loop made between the developed and opposing
power curves is the largest and therefore the engine
would respond to the changing driving conditions
with the greatest flexibility.
If top gear is overdriven by 20%, as shown in
Fig. 3.26, the maximum engine power would be
developed at maximum vehicle speed. This then
provides the highest possible theoretical speed,
but the amount of reserve power over the road
resistance power is less, so that acceleration
response will not be as rapid as if a direct drive
top gear is used. Operating under these conditions,
the engine speed would never exceed the peak
power speed and so the engine could not `over-rev',
and as a result engine wear and noise would be
reduced. Benefits are also gained in fuel consumption as shown in Fig. 3.26. The lowest specific fuel
consumption is shifted to a higher cruising speed
which is desirable on motorway journeys.
Indulging in an excessive 40% overdrive top gear
prevents the engine ever reaching peak power so that
not only would maximum vehicle speed be reduced
compared to the 20% overdrive gearing, but the
much smaller difference in power developed to
power dissipated shown on the power curves would
Rear mounted two speed transfer box (Fig. 3.25)
In some gearbox designs, or where the auxiliary
equipment requires it, a rear mounted transfer
box may be more convenient. This transfer drive
arrangement uses either an extended monolayshaft
or a short extension shaft attached by splines to the
layshaft so that it protrudes out from the rear of the
gearbox (Fig. 3.25). The extended layshaft supports a pair of high and low speed gears which are
in permanent mesh with corresponding gears
mounted on the output shaft.
When the transfer box is in neutral, the gears on
the extended layshaft are free to revolve independently on this shaft. Engagement of either high or
low gear ratios is achieved by sliding the output
drive hub sleeve in to mesh with one or other sets of
adjacent dog teeth forming part of the transfer box
layshaft constant mesh gears. Thus high gear ratio
power flow passes from the layshaft to the constant
mesh high range gears to the output shaft and
coupling. Conversely, low gear ratio power transmission goes from the layshaft through the low
range gears to the output drive.
3.7 Overdrive considerations
Power is essential to propel a vehicle because it is a
measure of the rate of doing work, that is, the
amount of work being developed by the engine in
unit time. With increased vehicle speed, more work
has to be done by the engine in a shorter time.
The characteristic power curve over a speed
range for a petrol engine initially increases linearly
and fairly rapidly. Towards mid-speed the steepness of the power rise decreases until the curve
reaches a peak. It then bends over and declines
with further speed increase due to the difficulties
experienced in breathing at very high engine speeds
(Fig. 3.26).
A petrol engined car is usually geared so that in its
normal direct top gear on a level road the engine
speed exceeds the peak power speed by about 10 to
20% of this speed. Consequently, the falling power
curve will intersect the road resistance power curve.
The point where both the engine and road resistance
power curves coincide fixes the road speed at which
all the surplus power has been absorbed. Therefore
it sets the maximum possible vehicle speed.
Fig. 3.26 Effect of over and undergearing on vehicle performance
severely reduce the flexibility of driving in this gear.
It therefore becomes essential for more frequent
down changes with the slightest fall-off in road
speed. A further disadvantage with excessive overdrive is that the minimum specific fuel consumption
would be shifted theoretically to the engine's upper
speed range which in practice could not be reached.
An analysis of matching an engine's performance
to suit the driving requirements of a vehicle shows
that with a good choice of undergearing in top
gear for motorway cruising conditions, benefits of
prolonged engine life, reduced noise, better fuel
economy and less driver fatigue will be achieved.
Another major consideration is the unladen and
laden operation of the vehicle, particularly if it is
to haul heavy loads. Therefore most top gear overdrive ratios are arrived at as a compromise.
3.7.1 Epicyclic overdrive gearing
Epicyclic gear train overdrives are so arranged that
the input shaft drives the pinion carrier while the
output shaft is driven by the annular gear ring
(Figs 3.27 and 3.28). The gear train may be either of
simple (single stage) or compound (double stage)
design and the derived formula for each arrangement is as follows:
The amount of overdrive (undergearing) used
for cars, vans, coaches and commercial vehicles
varies from as little as 15% to as much as 45%.
This corresponds to undergearing ratios of between
0.87:1 and 0.69:1 respectively. Typical overdrive
ratios which have been frequently used are 0.82:1
(22%), 0.78:1 (28%) and 0.75:1 (37%).
Simple gear train (Fig. 3.27)
A ˆ S ‡ 2P
A ˆ number of annulus ring
gear teeth
S ˆ number of sun gear teeth
P ˆ number of planet
gear teeth
Overdrive gear ratio ˆ
Example 1 An overdrive simple epicyclic gear
train has sun and annulus gears with 21 and 75
teeth respectively. If the input speed from the engine
drives the planet carrier at 3000 rev/min, determine
the overdrive gear ratio,
the number of planet gear teeth,
the annulus ring and output shaft speed,
the percentage of overdrive.
a) Overdrive gear ratio ˆ
A ‡ S 75 ‡ 21
Compound gear train (Fig. 3.28)
PL (PL ‡ PS ‡ S)
PL (PL ‡ PS ‡ S) ‡ PS S
also A ˆ PL ‡ PS ‡ S
where A ˆ number of annulus ring
gear teeth
S ˆ number of sun
gear teeth
PS ˆ number of small planet
gear teeth
PL ˆ number of large planet
gear teeth
Overdrive gear ratio ˆ
Fig. 3.27
ˆ 0:78125
b) A ˆ S ‡ 2P
Simple epicycle overdrive gear train
ˆ 27 teeth
ˆ 3840 rev/min
3840 3000
d) Percentage of overdrive ˆ
c) Output speed ˆ
b) A ˆ PL ‡ PS ‡ S
ˆ 21 ‡ 24 ‡ 15
ˆ 60 teeth
c) Output speed ˆ
840 100
ˆ 28%
Example 2 A compound epicyclic gear train overdrive has sun, small planet and large planet gears
with 21, 15 and 24 teeth respectively. Determine
the following if the engine drives the input planet
carrier at 4000 rev/min.
d) Percentage of overdrive ˆ
24 60
(24 60) ‡ 315 1755
878 100
3.7.2 Simple epicyclic overdrive gear train
(Fig. 3.27)
If the sun gear is prevented from rotating and
the input shaft and planet carrier are rotated, the
pinion gears will be forced to revolve around the
fixed sun gear and these pinions will revolve simultaneously on their own axes provided by the carrier
As a result, motion will be transferred from the
carrier and pinion gears to the annulus ring gear
due to the separate rotary movement of both the
planet carrier and the revolving planet gears, thus
PL (PL ‡ PS ‡ S)
a) Overdrive
gear ratio
PL (PL ‡ PS ‡ S) ‡ PS S
24 (24 ‡ 15 ‡ 21)
24 (24 ‡ 15 ‡ 21) ‡ (15 21)
4878 4000
ˆ 21:95%
The overdrive gear ratio,
the number of annulus ring gear teeth,
the annulus ring and output shaft speed,
the percentage of overdrive.
ˆ 4878 rev/min
ˆ 0:82
Fig. 3.28 Compound epicycle gear train
the annulus and therefore the output shaft will be
compelled to revolve at a slightly faster speed.
the engine, the output shaft will try to run faster
than the input shaft and so tend to release the
unidirectional clutch rollers, but this is prevented
by the inner cone clutch locking the sun gear to the
annulus, thereby jamming the sun, planet and
annular epicyclic gear train so that they cannot
revolve relative to each other.
Engagement of the inner cone clutch to the external cone surface of the annulus gear is provided by
four stationary thrust springs (only one shown)
which are free to exert their axial load against a
thrust plate. This in turn transfers thrust by way of
a ball bearing to the rotating cone clutch support
member splined to the sun gear sleeve. This overrun and reverse torque will be transmitted between
the engine and transmission in direct drive.
Owing to the helical cut teeth of the gear wheels,
an end thrust exists between the planet gears and
the sun gear during overrun and reverse which
tends to push the latter rearwards. Therefore, additional clamping load between the cone clutch faces
is necessary.
3.7.3 Compound epicyclic overdrive gear train
(Fig. 3.28)
For only small degrees of overdrive (undergearing),
for example 0.82:1 (22%), the simple epicyclic
gearing would need a relatively large diameter
annulus ring gear; about 175 mm if the dimension
of the gear teeth are to provide adequate strength.
A way of reducing the diameter of the annulus ring
gear for a similar degree of overdrive is to utilize a
compound epicyclic gear train which uses double
pinion gears on each carrier pin instead of one size
of pinion. By this method, the annulus diameter is
reduced to about 100 mm and there are only 60
teeth compared to the 96 teeth annulus used with
the simple epicyclic gear train.
To transmit power, the sun gear is held still
whilst the input shaft and planet carrier are rotated.
This compels the large planet gear to roll around
the stationary sun gear and at the same time forces
each pair of combined pinion gears to revolve
about their carrier pin axes.
Consequently the small pinion gear will impart
both the pinion carrier orbiting motion and the
spinning pinion gear motion to the annulus ring
gear so that the output shaft will be driven at a
higher speed to that of the input shaft.
Overdrive (Fig. 3.29) When overdrive is engaged,
the cone clutch, which is supported on the splined
sleeve of the sun gear, is moved over so that its
outer friction facing is in contact with the internal
cone brake attached to the casing. Consequently
the sun gear is held stationary. With the sun gear
held still and the input shaft and planet carrier
rotating, the planet gears are forced to rotate
about their own axes and at the same time roll
around the fixed sun gear, with the result that the
annulus gear is driven at a faster speed than the
input shaft. This causes the unidirectional clutch
outer member (annular carrier) to overrun the
inner member (planet carrier) so that the wedged
rollers on their ramps are released. Pulling the cone
clutch away from the annulus cone and into frictional contact with the brake casing cone against
the axial load of the six thrust springs is achieved
by means of hydraulic oil pressure. This pressure
acts upon two slave pistons (only one shown) when
a valve is opened by operating the driver controlled
selector switch.
The outward movement of the slave pistons,
due to the hydraulic pressure, draws the stationary thrust plate, ball bearing and rotating clutch
member away from the annular cone and into
engagement with the outer brake cone, thereby
locking the sun gear to the casing. Of the helix
angle of the gear teeth, the torque reaction tends
to push the sun gear forward so that extra end
3.7.4 Laycock simple gear train overdrive
Description (Fig. 3.29) The overdrive unit is
attached to the rear of the gearbox and it consists
of a constant mesh helical toothed epicyclic gear
train which has a central sunwheel meshing with
three planet gears which also mesh with an internally toothed annulus gear. The planet gears are
supported on a carrier driven by the input shaft
whilst the annulus is attached to the output shaft
via a carrier forming an integral part of both
members. A double cone clutch selects the different
ratios; when engaged one side of the clutch provides direct drive and when the other side is used,
Direct drive (Fig. 3.29) Direct drive is obtained
when the inner cone clutch engages with the outer
cone of the annular gear. Power will then be conveyed via the unidirectional clutch to the output
shaft by means of the rollers which are driven up
inclined ramps and wedged between the inner and
outer clutch members. When the vehicle overruns
Fig. 3.29 Laycock single epicycle overdrive
thrust is necessary to maintain sufficient clamping
thrust between the frictional faces of the cones in
the brake position.
cone clutch frees the sun gear and removes the load
from the engine. The engine speed increases immediately until it catches up with the output shaft, at
which point the unidirectional clutch rollers climb
up their respective ramps and jam. The input shaft's
power coming from the engine is now permitted to
drive the output shaft, which in turn transmits drive
to the propellor shaft. At the same time the doublesided cone clutch completes its movement and
engages the annular ring cone.
Direct and overdrive controlled gear change action
(Fig. 3.29) When direct drive is selected, hydraulic pressure is steadily increased and this gradually
releases the double-sided cone clutch member from
the cone brake fixed to the casing. The release of the
Overdrive is engaged when the double-sided
cone clutch moves away from the annulus gear
cone and makes contact with the stationary cone
brake, thus bringing the sliding cone clutch member and sun gear to rest. As a result of the sun gear
being held stationary, the gears now operate as an
epicyclic step up gear ratio transmission. During
the time the double-sided cone clutch member
moves from the annulus cone to the brake cone
the clutch will slip. This now permits the unidirectional roller clutch to transmit the drive. Whilst the
input ramp member rotates as fast as the output
ramp member the roller clutch drives. However, as
the annular ring gear speed rises above that of the
input shaft, the rollers will disengage themselves
from their respective ramps thereby diverting the
drive to the epicyclic gear train.
cylinders (only one shown) and also to the solenoid
controlled valve and dashpot regulator relief valve.
The dashpot pressure regulator ensures a smooth
overdrive engagement and disengagement under differing operating conditions. When in direct drive the
pump to slave cylinder's line pressure is determined
by the regulator relief valve spring tension which
controls the blow-off pressure of the oil escaping to
the lubrication system. This line residual pressure in
direct drive is normally maintained at about 2.8 bar,
but when engaging overdrive it is considerably raised
by the action of the dashpot to about 20±40 bar.
Overdrive engagement Energizing the solenoid
draws down the armature, thereby opening the
inlet valve and closing the outlet valve. Oil at residual line pressure will now pass through the control
orifice to the base of the dashpot regulator relief
valve causing the dashpot to rise and compress
both the dashpot spring and relief valve spring.
Consequently, the pump to slave cylinder pressure
circuit will gradually build up as the dashpot spring
shortens and increases in stiffness until the dashpot
piston has reached its stop, at which point the
operating pressure will be at a maximum. It is this
gradual increase in line pressure which provides the
progressive compression of the clutch thrust springs
and the engagement of the cone clutch with the fixed
cone brake.
Electrical system (Fig. 3.29) Overdrive or direct
drive gear ratio selection is controlled by an electrical circuit which includes an overdrive on/off
switch, inhibitor switch and a relay switch. An
inhibitor switch is incorporated in the circuit to
prevent the engagement of overdrive in reverse
and some or all of the indirect gears. A relay switch
is also included in the circuit so that the overdrive
on/off switch current rating may be small compared to the current draw requirements of the
control solenoid. The overdrive may be designed
to operate only in top gear, but sometimes the
overdrive is permitted to be used in third or even
second gear. Selection and engagement of overdrive by the driver is obtained by a steering column
or fascia panel switch. When the driver selects
overdrive in top gear or one of the permitted indirect gears, say third, the on/off switch is closed and
the selected gearbox gear ratio selector rod will
have pushed the inhibitor switch button into the
closed switch position. Current will now flow from
the battery to the relay switch, magnetizing the
relay winding so that as the relay contacts close, a
larger current will immediately energize the solenoid and open the control valve so that overdrive
will be engaged.
Direct drive engagement De-energizing the solenoid closes the inlet valve and opens the outlet
valve. This prevents fresh oil entering the dashpot
cylinder and allows the existing oil under the dashpot to exhaust by way of the control orifice and the
outlet valve back to the sump. The control orifice
restricts the flow of escaping oil so that the pressure
drop is progressive. This enables the clutch thrust
springs to shift the cone clutch very gradually into
contact with the annulus cone.
3.7.5 Laycock compound gear train overdrive
(Fig. 3.30)
Overdrive When overdrive is selected, the doublesided cone clutch contacts the brake cone which
forms part of the casing. This brings the sun gear
which is attached to the sliding clutch member to a
The input drive passes from the pinion carrier to
the annulus ring and hence to the output shaft
through the small planet gear. At the same time, the
Hydraulic system (Fig. 3.29) A plunger type
pump driven by an eccentric formed on the input
shaft supplies the hydraulic pressure to actuate the
slave pistons and thereby operates the clutch. The
pump draws oil from the sump through a filter (not
shown). It is then pressurized by the plunger and
delivered through a non-return valve to both slave
Fig. 3.30 Laycock double epicycle overdrive
large planet gear absorbs the driving torque reaction
and in the process is made to revolve around the
braked sun gear. The overdrive condition is created
by the large planet gears being forced to roll `walk'
about the sun gear, while at the same time revolving
on their own axes. As a result, the small planet gears,
also revolving on the same carrier pins as the large
planet gears, drive forward the annular ring gear at a
faster speed relative to that of the input.
The overall gear ratio step up is achieved by
having two stages of meshing gear teeth; one
between the large pinion and sun gear and the
other between the small pinion and annulus ring
gear. By using this compound epicyclic gear train, a
relatively large step up gear ratio can be obtained
for a given diameter of annulus ring gear compared
to a single stage epicyclic gear train.
the low pressure ball valve will open and relieve the
excess pressure. Under these conditions the axial
load exerted by the clutch thrust springs will clamp
the double-sided floating conical clutch member to
the external conical shaped annulus ring gear.
Direct drive (Fig. 3.30) Direct drive is attained by
releasing the double-sided cone clutch member from
the stationary conical brake and shifting it over so
that it contacts and engages the conical frictional
surface of the annulus ring gear. The power flow
from the input shaft and planet carrier now divides
into two paths Ð the small planet gear to annulus
ring gear route and the large planet gear, sun gear
and double-sided clutch member route, again finishing up at the annulus ring gear. With such a closed
loop power flow arrangement, where the gears cannot revolve independently to each other, the gears
jam so that the whole gear train combination rotates
as one about the input to output shaft axes. It
thereby provides a straight through direct drive. It
should be observed that the action of the unidirectional roller clutch is similar to that described for the
single stage epicyclic overdrive.
Overdrive engagement To select overdrive the
solenoid is energized. This closes the solenoid ball
valve, preventing oil escaping via the lubrication
system back to the sump. Oil pressure will now
build up to about 26±30 bar, depending on vehicle
application, until sufficient thrust acts on both
slave pistons to compress the clutch thrust springs,
thereby permitting the double-sided clutch member
to shift over and engage the conical surface of the
stationary brake. To enable the engagement action
to overdrive to progress smoothly and to limit the
maximum hydraulic pressure, a high pressure valve
jumper is made to be pushed back and progressively open. This controls and relieves the pressure
rise which would otherwise cause a rough, and
possibly sudden, clutch engagement.
Clutch operating (Fig. 3.30) Engagement of
direct drive and overdrive is achieved in a similar
manner to that explained under single stage epicyclic overdrive unit.
Direct drive is provided by four powerful springs
holding the double-sided conical clutch member in
frictional contact with the annulus ring gear. Conversely, overdrive is obtained by a pair of hydraulic
slave pistons which overcome and compress the
clutch thrust springs, pulling the floating conical
clutch member away from the annulus and into
engagement with the stationary conical brake.
3.8 Setting gear ratios
Matching the engine's performance characteristics
to suit a vehicle's operating requirements is provided by choosing a final drive gear reduction and
then selecting a range of gear ratios for maximum
performance in terms of the ability to climb gradients, achievement of good acceleration through the
gears and ability to reach some predetermined
maximum speed on a level road.
3.8.1 Setting top gear
To determine the maximum vehicle speed, the engine
brake power curve is superimposed onto the power
requirement curve which can be plotted from the
sum of both the rolling (Rr ) and air (Ra ) resistance
covering the entire vehicle's speed range (Fig. 3.31).
The total resistance R opposing motion at any
speed is given by:
Hydraulic system (Fig. 3.30) Pressure supplied by
the hydraulic plunger type pump draws oil from the
sump and forces it past the non-return ball valve to
both the slave cylinders and to the solenoid valve
and the relief valve.
R ˆ Rr ‡ Ra
Direct drive engagement When direct drive is
engaged, the solenoid valve opens due to the solenoid being de-energized. Oil therefore flows not
only to the slave cylinders but also through the
solenoid ball valve to the overdrive lubrication system where it then spills and returns to the sump. A
relatively low residual pressure will now be maintained within the hydraulic system. Should the oil
pressure rise due to high engine speed or blockage,
ˆ 10Cr W ‡ CD AV 2
where Cr ˆ coefficient of rolling resistance
W ˆ gross vehicle weight (kg)
CD ˆ coefficient of aerodynamic resistance (drag)
A ˆ projected frontal area of vehicle (m2 )
V ˆ speed of vehicle (km/h)
Fig. 3.31
Forces opposing vehicle motion over its speed
Fig. 3.32 Relationship of power developed and road
power required over the vehicle's speed range
3.8.2 Setting bottom gear
The maximum payload and gradient the vehicle is
expected to haul and climb determines the necessary
tractive effort, and hence the required overall gear
ratio. The greatest gradient that is likely to be
encountered is decided by the terrain the vehicle is
to operate over. This normally means a maximum
gradient of 5 to 1 and in the extreme 4 to 1. The
minimum tractive effort necessary to propel a vehicle
up the steepest slope may be assumed to be approximately equivalent to the sum of both the rolling and
gradient resistances opposing motion (Fig. 3.31).
The rolling resistance opposing motion may be
determined by the formula:
The top gear ratio is chosen so that the maximum road speed corresponds to the engine speed at
which maximum brake power is obtained (or just
beyond) (Fig. 3.32).
Gearing is necessary to ensure that the vehicle
speed is at a maximum when the engine is developing approximately peak power.
Linear wheel speed ˆ Linear road speed
dN 1000
V (m/min)
60 dN
; Final drive gear ratio GF ˆ
100 V
ˆ 0:06
Rr ˆ 10Cr W
ˆ final drive gear ratio
ˆ engine speed (rev/min)
ˆ effective wheel diameter (m)
ˆ road speed at which peak power is
developed (km/h)
Rr ˆ rolling resistance (N)
Cr ˆ coefficient of rolling resistance
W ˆ gross vehicle weight (kg)
Average values for the coefficient of rolling
resistance for different types of vehicles travelling
at very slow speed over various surfaces have been
determined and are shown in Table 3.2.
Likewise, the gradient resistance (Fig. 3.33)
opposing motion may be determined by the formula:
Example A vehicle is to have a maximum road
speed of 150 km/h. If the engine develops its peak
power at 6000 rev/min and the effective road wheel
diameter is 0.54 m, determine the final drive gear ratio.
0:06 dN
0:06 3:142 0:54 6000
GF ˆ
or 10W sin G
where Rg ˆ gradient resistance (N)
Rg ˆ
W ˆ gross vehicle weight (10W kg ˆ WN)
ˆ 4:07 : 1
G ˆ gradient (1 in x) ˆ sin 94
Table 3.2
Average values of coefficient of rolling
Vehicle type
Coefficient of rolling resistance (Cr)
Passenger Car 0.015
Medium hard soil
ˆ final drive gear ratio
ˆ bottom gear ratio
ˆ mechanical efficiency
ˆ tractive effort (N)
ˆ maximum engine torque (Nm)
ˆ effective road wheel radius (m)
Example A vehicle weighing 1500 kg has a
coefficient of rolling resistance of 0.015. The transmission has a final drive ratio 4.07:1 and an overall
mechanical efficiency of 85%.
If the engine develops a maximum torque of
100 Nm (Fig. 3.34) and the effective road wheel
radius is 0.27 m, determine the gearbox bottom
gear ratio.
Assume the steepest gradient to be encountered
is a one in four.
Note The coefficient of rolling resistance is the ratio of the
rolling resistance to the normal load on the tyre.
Cr ˆ W
Rr ˆ 10Cr W
ˆ 10 0:015 150 ˆ 225N
Rg ˆ
10 1500
ˆ 3750N
E ˆ Rr ‡ Rg
ˆ 3750 ‡ 225 ˆ 3975N
GB ˆ
Fig. 3.33
Gradient resistance to motion
3975 0:27
ˆ 3:1:1
100 4:07 0:85
Tractive effort ˆ Resisting forces opposing motion
ˆ Rr ‡ Rg (N)
E ˆ tractive effort (N)
R ˆ resisting forces (N)
Once the minimum tractive effort has been calculated, the bottom gear ratio can be derived in the
following way:
Driving torque ˆ Available torque
; Bottom gear ratio GB ˆ
Fig. 3.34
Engine torque to speed characteristics
3.8.3 Setting intermediate gear ratios
Ratios between top and bottom gears should be
spaced in such a way that they will provide the
tractive effort±speed characteristics as close to the
ideal as possible. Intermediate ratios can be best
selected as a first approximation by using a geometric progression. This method of obtaining the
gear ratios requires the engine to operate within the
same speed range in each gear, which is normally
selected to provide the best fuel economy.
Consider the engine to vehicle speed characteristics for each gear ratio as shown (Fig. 3.35). When
changing gear the engine speed will drop from the
highest NH to the lowest NL without any change in
road speed, i.e. V1 , V2 , V3 etc.
; G3 ˆ G2
; G4 ˆ G3
; G5 ˆ G4
ˆ 1st overall gear ratio
ˆ 2nd overall gear ratio
ˆ 3rd overall gear ratio
ˆ 4th overall gear ratio
ˆ 5th overall gear ratio
; G2 ˆ G1
is known as the minimum to maxNH
imum speed range ratio K for a given engine.
Now, gear G2 ˆ G1
Overall ˆ Engine speed (rev/min)
gear ratio Road wheel speed (rev/min)
The ratio
ˆ G1 K,
ˆ k (a constant)
ˆ G2 K ˆ (G1 K)K
ˆ G 1 K2
gear G4 ˆ G3
ˆ G3 K ˆ (G1 K 2 )K
ˆ G 1 K3
gear G5 ˆ G4
ˆ G4 K ˆ (G1 K 3 )K
ˆ G 1 K4 :
Wheel speed when engine is on the high limit NH in
first gear G1 ˆ
Wheel speed when engine is on the low limit NL in
second gear G2 ˆ
These wheel speeds must be equal for true rolling
gear G3 ˆ G2
Hence the ratios form a geometric progression.
Fig. 3.35 Gear ratios selected on geometric progression
The following relationship will also apply for a
five speed gearbox:
Example A transmission system for a vehicle
is to have an overall bottom and top gear ratio
of 20:1 and 4.8 respectively. If the minimum to maximum speeds at each gear changes are 2100 and
3000 rev/min respectively, determine the following:
G2 G3 G4 G5 NL
G1 G2 G3 G4 NH
G5 ˆ G1 K 4
K4 ˆ
Hence K ˆ
a) the intermediate overall gear ratios
b) the intermediate gearbox and top gear ratios.
ˆ 0:7
4 G5
In general, if the ratio of the highest gear (GT )
and that of the lowest gear (GB ) have been determined, and the number of speeds (gear ratios) of
the gearbox nG is known, the constant K can be
determined by:
a) 1st gear ratio G1 ˆ 20:0:1
2nd gear ratio G2 ˆ G1 K ˆ 20 0:7 ˆ 14:0:1
3rd gear ratio G3 ˆ G1 K 2 ˆ 20 0:72 ˆ 9:8:1
4th gear ratio G4 ˆ G1 K 3 ˆ 20 0:73 ˆ 6:86:1
5th gear ratio G5 ˆ G1 K 4 ˆ 20 0:74 ˆ 4:8:1
ˆ K nG
GT ˆ GB K nG
For commercial vehicles, the gear ratios in
the gearbox are often arranged in geometric
progression. For passenger cars, to suit the changing traffic conditions, the step between the ratios
of the upper two gears is often closer than that
based on geometric progression. As a result, this
will affect the selection of the lower gears to some
b) G1 ˆ
ˆ 4:166:1
G2 ˆ
ˆ 2:916:1
G3 ˆ
ˆ 2:042:1
G4 ˆ
ˆ 1:429:1
Top gear G5 ˆ
ˆ 1:0:1
Hydrokinetic fluid couplings and torque converters
A fluid drive uses hydrokinetic energy as a means
of transferring power from the engine to the transmission in such a way as to automatically match
the vehicle's speed, load and acceleration requirement characteristics. These drives may be of a
simple two element type which takes up the drive
smoothly without providing increased torque or
they may be of a three or more element unit
which not only conveys the power as required
from the engine to the transmission, but also multiplies the output torque in the process.
4.1 Hydrokinetic fluid couplings
(Figs 4.1 and 4.2)
The hydrokinetic coupling, sometimes referred to
as a fluid flywheel, consists of two saucer-shaped
discs, an input impeller (pump) and an output
turbine (runner) which are cast with a number of
flat radial vanes (blades) for directing the flow path
of the fluid (Fig. 4.1).
Owing to the inherent principle of the hydrokinetic coupling, there must be relative slip between
the input and output member cells exposed to each
Fig. 4.1 Fluid coupling action
other, and the vortex flow path created by pairs of
adjacent cells will be continuously aligned and
misaligned with different cells.
With equal numbers of cells in the two half
members, the relative cell alignment of all the cells
occurs together. Consequently, this would cause a
jerky transfer of torque from the input to the output
drive. By having differing numbers of cells within
the impeller and turbine, the alignment of each pair
of cells at any one instant will be slightly different
so that the impingement of fluid from one member
to the other will take place in various stages of
circulation, with the result that the coupling torque
transfer will be progressive and relatively smooth.
The two half-members are put together so that
the fluid can rotate as a vortex. Originally it was
common practice to insert at the centre of rotation a
hollow core or guide ring (sometimes referred to as
the torus) within both half-members to assist in
establishing fluid circulation at the earliest moment
of relative rotation of the members. These couplings
had the disadvantage that they produced considerable drag torque whilst idling, this being due mainly
to the effectiveness of the core guide in circulating
fluid at low speeds. As coupling development progressed, it was found that turbine drag was reduced
at low speeds by using only a core guide on the
impeller member (Fig. 4.2). With the latest design
Fig. 4.2 Fluid coupling
these cores are eliminated altogether as this also
reduces fluid interference in the higher speed range
and consequently reduces the degree of slip for a
given amount of transmitted torque (Fig. 4.6).
by the impeller around its axis and secondly it
circulates round the cells in a vortex motion.
This circulation of fluid only continues as long as
there is a difference in the angular speeds of the
impeller and turbine, because only then is the centrifugal force experienced by the fluid in the faster
moving impeller greater than the counter centrifugal force acting on the fluid in the slower moving
turbine member. The velocity of the fluid around
the couplings' axis of rotation increases while it
flows radially outwards in the impeller cells due to
the increased distance it has moved from the centre
of rotation. Conversely, the fluid velocity decreases
when it flows inwards in the turbine cells. It therefore follows that the fluid is given kinetic energy by
the impeller and gives up its kinetic energy to the
turbine. Hence there is a transference of energy
from the input impeller to the output turbine, but
there is no torque multiplication in the process.
4.1.1 Hydrokinetic fluid coupling principle of
operation (Figs 4.1 and 4.3)
When the engine is started, the rotation of the
impeller (pump) causes the working fluid trapped
in its cells to rotate with it. Accordingly, the fluid is
subjected to centrifugal force and is pressurized so
that it flows radially outwards.
To understand the principle of the hydrokinetic
coupling it is best to consider a small particle of
fluid circulating between one set of impeller and
turbine vanes at various points A, B, C and D as
shown in Figs 4.1 and 4.3.
Initially a particle of fluid at point A, when the
engine is started and the impeller is rotated, will
experience a centrifugal force due to its mass and
radius of rotation, r. It will also have acquired some
kinetic energy. This particle of fluid will be forced
to move outwards to point B, and in the process
of increasing its radius of rotation from r to R, will
now be subjected to considerably more centrifugal
force and it will also possess a greater amount of
kinetic energy. The magnitude of the kinetic energy
at this outermost position forces it to be ejected
from the mouth of the impeller cell, its flow path
making it enter one of the outer turbine cells at
point C. In doing so it reacts against one side of the
turbine vanes and so imparts some of its kinetic
energy to the turbine wheel. The repetition of fluid
particles being flung across the junction between the
impeller and turbine cells will force the first fluid
particle in the slower moving turbine member
(having reduced centrifugal force) to move inwards
to point D. Hence in the process of moving inwards
from R to r, the fluid particle gives up most of its
kinetic energy to the turbine wheel and subsequently
this is converted into propelling effort and motion.
The creation and conversion of the kinetic
energy of fluid into driving torque can be visualized
in the following manner: when the vehicle is at rest
the turbine is stationary and there is no centrifugal
force acting on the fluid in its cells. However, when
the engine rotates the impeller, the working fluid
in its cells flows radially outwards and enters the
turbine at the outer edges of its cells. It therefore
causes a displacement of fluid from the inner edges
of the turbine cells into the inner edges of the
impeller cells, thus a circulation of the fluid will
be established between the two half cell members.
The fluid has two motions; firstly it is circulated
4.1.2 Hydrokinetic fluid coupling velocity
diagrams (Fig. 4.3)
The resultant magnitude of direction of the fluid
leaving the impeller vane cells, VR, is dependent
upon the exit velocity, VE, this being a measure of
the vortex circulation flow rate and the relative
linear velocity between the impeller and turbine, VL.
The working principle of the fluid coupling
may be explained for various operating conditions
assuming a constant circulation flow rate by means
of velocity vector diagrams (Fig. 4.3).
When the vehicle is about to pull away, the engine
drives the impeller with the turbine held stationary.
Because the stalled turbine has no motion, the relative forward (linear) velocity VL between the two
members will be large and consequently so will the
resultant entry velocity VR. The direction of fluid
flow from the impeller exit to turbine entrance will
make a small angle 1 , relative to the forward direction of motion, which therefore produces considerable drive thrust to the turbine vanes.
As the turbine begins to rotate and catch up to
the impeller speed the relative linear speed is
reduced. This changes the resultant fluid flow
direction to 2 and decreases its velocity. The net
output thrust, and hence torque carrying capacity,
will be less, but with the vehicle gaining speed there
is a rapid decline in driving torque requirements.
At high turbine speeds, that is, when the output
to input speed ratio is approaching unity, there will
be only a small relative linear velocity and resultant
entrance velocity, but the angle 3 will be large.
This implies that the magnitude of the fluid thrust
will be very small and its direction ineffective in
Fig. 4.4 Relationship of torque capacity efficiency and
speed ratio for fluid couplings
Fig. 4.5 Relationship of engine speed, torque and slip
for a fluid coupling
Fig. 4.3 Principle of the fluid coupling
rotating the turbine. Thus the output member will
slip until sufficient circulating fluid flow imparts
enough energy to the turbine again.
It can be seen that at high rotational speeds the
cycle of events is a continuous process of output
speed almost, but never quite, catching up to input
speed, the exception being when the drive changes
from engine driven to overrun transmission driven
when the operating conditions will be reversed.
in impeller speed, considerably raises the coupling
torque carrying capacity. A further controlling factor which affects the torque transmitted is the
quantity of fluid circulating between the impeller
and turbine. Raising or lowering the fluid level in
the coupling increases or decreases the torque
which can be transmitted to the turbine (Fig. 4.4).
4.3 Fluid friction coupling (Figs 4.6 and 4.7)
A fluid coupling has the take-up characteristics
which particularly suit the motor vehicle but it
suffers from two handicaps that are inherent in
the system. Firstly, idling drag tends to make the
vehicle creep forwards unless the parking brake is
fully applied, and secondly there is always a small
amount of slip which is only slight under part load
(less than 2%) but becomes greater when transmitting anything near full torque.
These limitations have been overcome for large
truck applications by combining a shoe and drum
centrifugally operated clutch to provide a positive
lock-up at higher output speeds with a smaller
coreless fluid coupling than would be necessary if
the drive was only to be through a fluid coupling.
The reduced size and volume of fluid circulation in
the coupling thereby eliminate residual idling drag
(Fig. 4.6).
With this construction there is a shoe carrier
between the impeller and flywheel attached to the
output shaft. Mounted on this carrier are four brake
shoes with friction material facings. They are each
pivoted (hinged) to the carrier member at one end
and a garter spring (coil springs shown on front view
to illustrate action) holds the shoes in their retraction position when the output shaft is at rest.
When the engine is accelerated the fluid coupling
automatically takes up the drive with maximum
smoothness. Towards maximum engine torque
speed the friction clutch shoes are thrown outwards
by the centrifugal effect until they come into contact with the flywheel drum. The frictional grip will
now lock the input and output drives together.
Subsequently the fluid vortex circulation stops
and the fluid coupling ceases to function (Fig. 4.7).
Relative slip between input and output member in
low gear is considerably reduced, due to the automatic friction clutch engagement, and engine braking is effectively retained down to idling speeds.
4.2 Hydrokinetic fluid coupling efficiency and
torque capacity (Figs 4.4 and 4.5)
Coupling efficiency is the ratio of the power available at the turbine to the amount of power supplied
to the impeller. The difference between input and
output power, besides the power lost by fluid shock,
friction and heat, is due mainly to the relative slip
between the two members (Fig. 4.4). A more useful
term is the percentage slip, which is defined as the
ratio of the difference in input and output speeds
divided by the input speed and multiplied by 100.
N n
i:e: % slip ˆ
The percentage slip will be greatly influenced by
the engine speed and output turbine load conditions
(Fig. 4.5). A percentage of slip must always exist to
create a sufficient rate of vortex circulation which is
essential to impart energy from the impeller to the
turbine. The coupling efficiency is at best about 98%
under light load high rotational speed conditions,
but this will be considerably reduced as turbine output
load is increased or impeller speed is lowered. If the
output torque demand increases, more slip will occur
and this will increase the vortex circulation velocity
which will correspondingly impart more kinetic
energy to the output turbine member, thus raising
the torque capacity of the coupling. An additional
feature of such couplings is that if the engine should
tend to stall due to overloading when the vehicle
is accelerated from rest, the vortex circulation will
immediately slow down, preventing further torque
transfer until the engine's speed has recovered.
Fluid coupling torque transmitting capacity for a
given slip varies as the fifth power of the impeller
internal diameter and as the square of its speed.
i:e: T / D5 N 2
D ˆ impeller diameter
N ˆ impeller speed (rev/min)
4.4 Hydrokinetic three element torque converter
(Figs 4.8 and 4.9)
A three element torque converter coupling is comprised of an input impeller casing enclosing the
Thus it can be seen that only a very small
increase in impeller diameter, or a slight increase
Fig. 4.6 Fluid friction coupling
output turbine wheel. There are about 26 and 23
blades for the impeller and turbine elements respectively. Both of these elements and their blades are
fabricated from low carbon steel pressings. The third
element of the converter called the stator is usually
an aluminium alloy casting which may have something in the order of 15 blades (Figs 4.8 and 4.9).
The working fluid within a converter when the
engine is operating has two motions:
1 Fluid trapped in the impeller and turbine vane
cells revolves bodily with these members about
their axis of rotation.
2 Fluid trapped between the impeller and turbine
vane cells and their central torus core rotates in a
circular path in the section plane, this being
known as its vortex motion.
When the impeller is rotated by the engine, it acts
as a centrifugal pump drawing in fluid near the
Fig. 4.7 Relationship of torque carrying capacity, efficiency and output speed for a fluid coupling
centre of rotation, forcing it radially outwards
through the cell passages formed by the vanes to the
impeller peripheral exit. Here it is ejected due to its
momentum towards the turbine cell passages and in
the process acts at an angle against the vanes, thus
imparting torque to the turbine member (Fig. 4.8).
The fluid in the turbine cell passages moves
inwards to the turbine exit. It is then compelled to
flow between the fixed stator blades (Fig. 4.9). The
reaction of the fluid's momentum as it glides over
the curved surfaces of the blades is absorbed by the
casing to which the stator is held and in the process
it is redirected towards the impeller entrance. It
enters the passages shaped by the impeller vanes.
As it acts on the drive side of the vanes, it imparts
a torque equal to the stator reaction in the direction
of rotation (Fig. 4.8).
It therefore follows that the engine torque
delivered to the impeller and the reaction torque
transferred by the fluid to the impeller are both
transmitted to the output turbine through the media
of the fluid.
Output turbine
4.4.1 Hydrokinetic three element torque
converter principle of operation (Fig. 4.8)
When the engine is running, the impeller acts as
a centrifugal pump and forces fluid to flow radially
around the vortex passage made by the vanes and
core of the three element converter. The rotation
of the impeller by the engine converts the engine
power into hydrokinetic energy which is utilized in
Fig. 4.8 Three element torque converter action
Fig. 4.9 Three element torque converter
providing a smooth engine to transmission take-up
and in producing torque multiplication if a third
fixed stator member is included.
An appreciation of the principle of the converter
can be obtained by following the movement and
events of a fluid particle as it circulates the vortex
passage (Fig. 4.8).
Consider a fluid particle initially at the small
diameter entrance point A in the impeller. As the
impeller is rotated by the engine, centrifugal force
will push the fluid particle outwards to the impeller's
largest exit diameter, point B. Since the particle's
circumferential distance moved every revolution
will be increased, its linear velocity will be greater
and hence it will have gained kinetic energy.
Pressure caused by successive particles arriving
at the impeller outermost cell exit will compel the
particle to be flung across the impeller±turbine
junction where it acts against the side of cell vane
it has entered at point C and thereby transfers some
of its kinetic energy to the turbine wheel. Because
the turbine wheel rotates at a lower speed relative
to the impeller, the pressure generated in the impeller will be far greater than in the turbine. Subsequently the fluid particle in the turbine curved
passage will be forced inwards to the exit point D
and in doing so will give up more of its kinetic
energy to the turbine wheel.
The fluid particle, still possessing kinetic energy
at the turbine exit, now moves to the stator blade's
entrance side to point E. Here it is guided by the
curvature of the blades to the exit point F.
From the fixed stator (reactor) blades the fluid
path is again directed to the impeller entrance point
A where it imparts its hydrokinetic energy to the
impeller, this being quite separate to the kinetic
energy produced by the engine rotating the impeller.
Note that with the fluid coupling, the transfer of
fluid from the turbine exit to the impeller entrance
is direct. Thus the kinetic energy gained by the
input impeller is that lost by the output turbine
and there is no additional gain in output turning
effort, as is the case when a fixed intermediate
stator is incorporated.
4.4.2 Hydrokinetic three element torque
converter velocity diagrams (Figs 4.9 and 4.10)
The direction of fluid leaving the turbine to enter
the stator blades is influenced by the tangential exit
velocity which is itself determined by the vortex
circulating speed and the linear velocity due to the
rotating turbine member (Fig. 4.10).
When the turbine is in the stalled condition and
the impeller is being driven by the engine, the direction
of the fluid leaving the impeller will be determined
entirely by the curvature and shape of the turbine
vanes. Under these conditions, the fluid's direction
of motion, 1 , will make it move deep into the concave side of the stator blades where it reacts and is
Fig. 4.10 Principle of the single stage torque converter
made to flow towards the entrance of the impeller in
a direction which provides the maximum thrust.
Once the turbine begins to rotate, the fluid will
acquire a linear velocity so that the resultant
effective fluid velocity direction will now be 2 .
A reduced backward reaction to the stator will be
produced so that the direction of the fluid's
momentum will not be so effective.
As the turbine speed of rotation rises, the fluid's
linear forward velocity will also increase and,
assuming that the turbine's tangential exit velocity
does not alter, the resultant direction of the fluid
will have changed to 3 where it now acts on the
convex (back) side of the stator blades.
Above the critical speed, when the fluid's thrust
changes from the concave to the convex side of the
blades, the stator reaction torque will now act in the
opposite sense and redirect the fluid. Thus its resultant direction towards the impeller entry passages will
hinder instead of assist the impeller motion. The result
of this would be in effect to cancel out some of the
engine's input torque with further speed increases.
The inherent speed limitation of a hydrokinetic
converter is overcome by building into the stator
hub a one way clutch (freewheel) device (Fig. 4.9).
Therefore, when the direction of fluid flow changes
sufficiently to impinge onto the back of the blades,
the stator hub is released, allowing it to spin freely
between the input and output members. The freewheeling of the stator causes very little fluid interference, thus the three element converter now
becomes a two element coupling. This condition
prevents the decrease in torque for high output
speeds and produces a sharp rise in efficiency at
output speeds above the coupling point.
zero. Above this speed the stator is freewheeled.
This offers less resistance to the circulating fluid
and therefore produces an improvement in coupling efficiency (Figs 4.11 and 4.12).
This description of the operating conditions of the
converter coupling shows that if the transmission
is suddenly loaded the output turbine speed will
automatically drop, causing an increase in fluid
circulation and correspondingly a rise in torque
multiplication, but conversely a lowering of efficiency
due to the increased slip between input and output
members. When the output conditions have changed
and a reduction in load or an increase in turbine speed
follows the reverse happens; the efficiency increases
and the output to input torque ratio is reduced.
4.5 Torque converter performance terminology
(Figs 4.11 and 4.12)
To understand the performance characteristics of
a fluid drive (both coupling and converter), it is
essential to identify and relate the following terms
used in describing various relationshipsand conditions.
4.5.1 Fluid drive efficiency (Figs 4.11 and 4.12)
A very convenient method of expressing the energy
losses, due mainly to fluid circulation within a fluid
drive at some given output speed or speed ratio, is
4.4.3 Hydrokinetic torque converter
characteristics (Figs. 4.11 and 4.12)
Maximum torque multiplication occurs when there
is the largest speed difference between the impeller
and turbine. A torque output to input ratio of
about 2:1 normally occurs with a three element
converter when the turbine is stationary. Under
such conditions, the vortex rate of fluid circulation
will be at a peak. Subsequently the maximum
hydrokinetic energy transfer from the impeller to
turbine then stator to impeller again takes place
(Figs 4.11 and 4.12). As the turbine output speed
increases relative to the impeller speed, the efficiency rises and the vortex velocity decreases and
so does the output to input torque ratio until eventually the circulation rate of fluid is so low that it
can only support a 1:1 output to input torque
ratio. At this point the reaction torque will be
Fig. 4.11 Characteristic performance curves for a three
element converted coupling
impeller to turbine speed variation, with the result
that the vortex fluid circulation and correspondingly torque conversion are at a maximum, conversely converter efficiency is zero. Whilst these stall
conditions prevail, torque conversion loading drags
the engine speed down to something like 60±70% of
the engine's maximum torque speed, i.e. 1500±2500
rev/min. A converter should only be held in the
stall condition for the minimum of time to prevent
the fluid being overworked.
4.5.5 Design point (Figs 4.11 and 4.12)
Torque converters are so designed that their internal passages formed by the vanes are shaped so as
to make the fluid circulate with the minimum of
resistance as it passes from one member to another
member at definite impeller to turbine speed ratio,
known as the design point. A typical value might be
Above or below this optimum speed ratio, the
resultant angle and direction of fluid leaving one
member to enter another will alter so that the flow
from the exit of one member to the entry of another
will no longer be parallel to the surfaces of the vanes,
in fact it will strike the sides of the passage vanes
entered. When the exit and entry angles of the vanes
do not match the effective direction of fluid motion,
some of its momentum will be used up in entrance
losses and consequently the efficiency declines as the
speed ratio moves further away on either side of the
design point. Other causes of momentum losses are
internal fabrication finish, surface roughness and
inter-vane or blade thickness interference. If the
design point is shifted to a lower speed ratio, say 0.6,
the torque multiplication will be improved at
stall and lower speed conditions at the expense of
an earlier fall-off in efficiency at the high speed ratio
such as 0.8. There will be a reduction in the torque
ratio but high efficiency will be maintained in the
upper speed ratio region.
Fig. 4.12 Characteristic performance curves for a converter coupling plotted to a base of output (turbine speed)
to input (impeller speed)
to measure its efficiency, that is, the percentage
ratio of output to input work done.
i:e: Efficiency ˆ
Output work done
Input work done
4.5.2 Speed ratio (Fig. 4.12)
It is frequently necessary to compare the output
and input speed differences at which certain events
occur. This is normally defined in terms of a speed
ratio of output (turbine) speed N2 to the input
(impeller) speed N1.
i:e: Speed ratio ˆ
4.5.3 Torque ratio (Fig. 4.12)
The torque multiplication within a fluid drive is
more conveniently expressed in terms of a torque
ratio of output (turbine) torque T2 to the input
(impeller) torque T1.
i:e: Torque ratio ˆ
4.5.6 Coupling point (Figs 4.11 and 4.12)
As the turbine speed approaches or exceeds that of
the impeller, the effective direction of fluid entering
the passages between the stator blades changes
from pushing against the concave face to being
redirected towards the convex (back) side of the
blades. At this point, torque conversion due to
fluid transfer from the fixed stator to the rotating
impeller, ceases. The turbine speed when the direction of the stator reaction is reversed is known as the
coupling point and is normally between 80 and 90%
of the impeller speed. At this point the stator is
released by the freewheel device and is then driven
4.5.4 Stall speed (Figs 4.11 and 4.12)
This is the maximum speed which the engine
reaches when the accelerator pedal is fully down, the
transmission in drive and the foot brake is fully
applied. Under such conditions there is the greatest
in the same direction as the impeller and turbine.
At and above this speed the stator blades will spin
with the impeller and turbine which then simply act
as a fluid coupling, with the benefit of increasing
efficiency as the turbine output speed approaches
but never reaches the input impeller speed.
torque (drive speed) is greater than that of the
output member. If the conditions are reversed and
the output member's applied torque (or speed)
becomes greater than that of the input, the output
member will overrun the input member (rotate
faster). Thus the lock between the two members
will be automatically released. Immediately the
drive will be discontinued which permits the input
and output members to revolve independently to
one another.
Overrun clutches can be used for a number of
applications, such as starter motor pre-engagement
drives, overdrives, torque converter stator release,
automatic transmission drives and final differential
Most overrun clutch devices take the form of
either the roller and wedge or sprag lock to engage
and disengage drive.
4.5.7 Racing or run-away point (Fig. 4.12)
If the converter does not include a stator freewheel
device or if the mechanism is jammed, then the
direction of fluid leaving the stator would progressively change from transferring fluid energy to
assist the impeller rotation to one of opposition
as the turbine speed catches up with that of the
impeller. Simultaneously, the vortex fluid circulation will be declining so that the resultant torque
capacity of the converter rapidly approaches zero.
Under these conditions, with the accelerator pedal
fully down there is very little load to hold back
the engine's speed so that it will tend to race or
run-away. Theoretically racing or run-away should
occur when both the impeller and turbine rotate at
the same speed and the vortex circulation ceases,
but due to the momentum losses caused by internal
fluid resistance, racing will tend to begin slightly
before a 1:1 speed ratio (a typical value might be
4.6.1 Overrun clutch with single diameter rollers
(Fig. 4.13)
A roller clutch is comprised of an inner and outer
ring member and a series of cylindrical rollers
spaced between them (see Fig. 4.13). Incorporated
between the inner and outer members is a cage
which positions the rollers and guides so that they
roll up and down their ramps simultaneously. One
of the members has a cylindrical surface concentric
with its axis, this is usually made the outer member.
The other member (inner one) has a separate wedge
ramp formed for each roller to react against. The
shape of these wedge ramps may be flat or curved
depending upon design. In operation each roller
provides a line contact with both the outer internal
cylindrical track and the external wedge ramp track
of the inner member.
When the input wedge member is rotated
clockwise and the output cylindrical member is
prevented from rotating or rotates anticlockwise
in the opposite direction, the rollers revolve and
climb up the wedge ramps, and thereby squeeze
themselves between the inner and outer member
tracks. Eventually the elastic compressive and
frictional forces created by the rollers against
these tracks prevents further roller rotation.
Torque can now transfer from the input inner
member to the outer ring member by way of these
jammed (locked) rollers.
If the output outer member tries to rotate in the
same direction but faster than the inner member,
the rollers will tend to rotate and roll down their
ramps, thereby releasing (unlocking) the outer
member from that of the input drive.
4.5.8 Engine braking transmitted through
converter or coupling on overrun
Torque converters are designed to maximize their
torque multiplication from the impeller to the turbine in the forward direction by adopting backward swept rotating member circulating passage
vanes. Unfortunately, in the reverse direction
when the turbine is made to drive the impeller on
transmission overrun, the exit and entry vane guide
angles of the members are unsuitable for hydrokinetic energy transference, so that only a limited
amount of engine braking torque can be absorbed
by the converter except at high output overrun
vehicle speeds. Conversely, a fluid coupling with
its flat radial vanes is able to transmit torque in
either drive or overrun direction with equal effect.
4.6 Overrun clutches
Various names have been used for these mechanisms such as freewheel, one way clutch and overrun
clutch, each one signifying the nature of the device
and is therefore equally appropriate.
A freewheel device is a means whereby torque
is transmitted from one stationary or rotating
member to another member, provided that input
Fig. 4.13 Overrun freewheel single diameter roller type clutch
4.6.2 Overrun clutch with triple diameter rollers
(Fig. 4.14)
This is a modification of the single roller clutch in
which the output outer member forms an internal
cylindrical ring, whereas the input inner member
has three identical external inclined plane profiles
(see Fig. 4.14). Situated between the inner and
outer tracks are groups of three different sized
rollers. An anchor block and energizing shoe is
arranged, between each group of rollers; the blocks
are screwed to the inner member while the shoes
(with the assistance of the springs) push the rollers
together and against their converging contact
tracks. The inclined plane profile required to
match the different diameter rollers provides a
variable wedge angle for each size of roller. It is
claimed that the take-up load of each roller will
be progressive and spread more evenly than
would be the case if all the rollers were of the
same diameter.
When the input inner ring takes up the drive, the
rollers revolve until they are wedged between the
inclined plane on the inner ring and the cylindrical
internal track of the outer member. Consequently
the compressive load and the frictional force thus
created between the rollers and tracks locks solid
the inner and outer members enabling them to
transmit torque.
If conditions change and the outer member
overruns the inner member, the rollers will be
compelled to revolve in the opposite direction to
when the drive was established towards the diverging end of the tracks. It thus releases the outer
member and creates the freewheel phase.
4.6.3 Sprag overrun clutch (Fig. 4.15)
A very reliable, compact and large torque-carrying
capacity overrun clutch is the sprag type clutch.
This dispenses with the wedge ramps or inclined
plane formed on the inner member which is
essential with roller type clutches (see Fig. 4.15).
The sprag clutch consists of a pair of inner and
outer ring members which have cylindrical external
and internal track surfaces respectively. Interlinking
the input and output members are circular rows of
short struts known as sprags. Both ends of the
sprags are semicircular with their radius of curvature being offset to each other so that the sprags
appear lopsided. In addition a tapered waste is
formed in their mid-region. Double cages are incorporated between inner and outer members. These
cages have rectangular slots formed to equally
space and locate the sprags around the inner and
outer tracks. During clutch engagement there will
be a slight shift between relative positions of the
two cages as the springs tilt, but the spacing will be
Fig. 4.14
Overrun freewheel triple diameter roller type clutch
accurately kept. This ensures that each sprag
equally contributes its share of wedge action
under all operating conditions. In between the
cages is a ribbon type spring which twists the sprags
into light contact with their respective track when
the clutch is in the overrun position.
When the inner ring member is rotated clockwise
and the outer ring member is held stationary or is
rotated anticlockwise, the spring tension lightly
presses the sprags against their track. This makes
the inner and outer members move in opposite
directions. The sprags are thus forced to tilt anticlockwise, consequently wedging their inclined
planes hard against the tracks and thereby locking
the two drive and driven members together.
As conditions change from drive to overrun and
the outer member rotates faster than the inner one,
the sprags will rotate clockwise and so release the
outer member: a freewheel condition is therefore
turbine and stator members within the converter,
so that there are more stages of conversion
(Fig. 4.16).
Consider the three stage torque converter. As shown
in Fig. 4.17, it is comprised of one impeller, three
interlinked output turbines and two fixed stator
Tracing the conversion vortex circuit starting
from the input rotating member (Fig. 4.18), fluid
is pumped from the impeller P by centrifugal force
to the two velocity components Vt and Vr, making
up the resultant velocity Vp which enters between
the first turbine blades T1 and so imparts some of
its hydrokinetic energy to the output. Fluid then
passes with a velocity VT1 to the first fixed stator,
S1, where it is guided and redirected with a resultant velocity VS1 , made up from the radial and
tangential velocities Vr and Vt to the second set of
turbine blades T2, so that momentum is given to
this member. Fluid is now transferred from the exit
of the second turbine T2 to the entrance of the
second stator S2. Here the reaction of the curved
blades deflects the fluid towards the third turbine
blades T3 which also absorb the fluid's thrust.
Finally the fluid completes its circulation cycle by
again entering the impeller passages.
The limitation of a multistage converter is that
there are an increased number of entry and exit
junctions between various members which raise
4.7 Three stage hydrokinetic torque converter
(Figs 4.16, 4.17 and 4.18)
A disadvantage with the popular three element
torque converter is that its stall torque ratio is
only in the region of 2:1, which is insufficient
for some applications, but this torque multiplication can be doubled by increasing the number of
Fig. 4.15 (a and b)
Overrun freewheel sprag type clutch
the fluid flow resistance around the torus passages.
Subsequently, efficiency drops off fairly rapidly
with higher speed ratios compared to the three
element converter (Fig. 4.16).
4.8 Polyphase hydrokinetic torque converter
(Figs 4.19 and 4.20)
The object of the polyphase converter is to extend
the high efficiency speed range (Fig. 4.20) of the
simple three element converter by altering the vane
or blade shapes of one element. Normally the stator
is chosen as the fluid entrance direction changes
with increased turbine speed. To achieve this, the
stator is divided into a number of separate parts, in
this case three, each one being mounted on its
own freewheel device built into its hub (Fig. 4.19).
The turbine exit and linear velocities VE and VL
produce an effective resultant velocity VR which
changes its direction of entry between stator blades
as the impeller and turbine relative speeds
Fig. 4.16 Characteristic performance curves of a three
stage converter
Fig. 4.17
Multistage (six element) torque converter
approach unity. It is this direction of fluid entering
between the stator blades which in phases releases
the various stator members.
spin in the same direction as the input and output
elements. The two remaining fixed stators now
form the optimum blade curvatures for high
Initial phase
Under stall speed conditions, the fluid flow from
the turbine to the stator is such as to be directed
onto the concave (rear) side of all three sections of
the divided stator blades, thus producing optimum
stator reaction for maximum torque multiplication
Third phase
With higher vehicle and turbine speeds, the fluid's
resultant direction of entry to the two remaining
held stators changes sufficiently to push from the
rear of the second set of stator blades S2. This section
will now be released automatically to enable the
third set of stator blades to operate with optimum
Second phase
As the turbine begins to rotate and the vehicle is
propelled forwards, the fluid changes its resultant
direction of entry to the stator blades so that it
impinges against the rear convex side of the first
stator blades S1. The reaction on this member is
now reversed so that it is released and is able to
Coupling phase
Towards unity speed ratio when the turbine speed
has almost caught up with the impeller, the fluid
entering the third stator blades S3 will have altered
its direction to such an extent that it releases this
Fig. 4.18 Principle of the three stage torque converter
last fixed set of blades. Since there is no more
reaction torque, conversion ceases and the input
and output elements act solely as a fluid coupling.
The disadvantage of the early fall in efficiency
with rising speed may be overcome by incorporating
a friction disc type clutch between the flywheel and
converter which is hydraulically actuated by means
of a servo piston (Fig. 4.21). This lock-up clutch is
designed to couple the flywheel and impeller
assembly directly to the output turbine shaft either
manually, at some output speed decided by the
driver which would depend upon the vehicle load
and the road conditions or automatically, at a definite input to output speed ratio normally in the
region of the design point here where efficiency is
highest (Fig. 4.22).
To overcome the problem of fluid drag between
the input and output members of the torque converter when working in conjunction with either
4.9 Torque converter with lock-up and gear
change friction clutches (Figs 4.21 and 4.22)
The two major inherent limitations with the torque
converter drive are as follows:
Firstly, the rapid efficiency decline once the
relative impeller to turbine speed goes beyond
the design point, which implies higher input speeds
for a given output speed and increased fuel consumption. Secondly, the degree of fluid drag at idle
speed which would prevent gear changing with
constant mesh and synchromesh gearboxes.
Fig. 4.19
Principle of a polystage torque converter
Fig. 4.22 Characteristic performance curves of a three
element converter with lock-up clutch
Fig. 4.20 Relationship of speed ratio, torque ratio and
efficiency for a polyphase stator torque converter
Fig. 4.21 Torque converter with lock-up and gear change function clutches
constant mesh or synchromesh gearboxes, a
conventional foot operated friction clutch can be
utilized between the converter and the gearbox.
When the pedal is depressed and the clutch is in
its disengaged position, the gearbox input primary
shaft and the output main shaft may be unified,
thereby enabling the gear ratio selected to be
engaged both smoothly and silently.
Semi- and fully automatic transmission
5.1 Automatic transmission considerations
Because it is difficult to achieve silent and smooth
gear ratio changes with a conventional constant
mesh gear train, automatic transmissions commonly adopt some sort of epicyclic gear arrangement, in which different gear ratios are selected by
the application of multiplate clutches and band
brakes which either hold or couple various members of the gear train to produce the necessary
speed variations. The problem of a gradual torque
take-up when moving away from a standstill has
also been overcome with the introduction of a
torque converter between the engine and transmission gearing so that engine to transmission slip is
automatically reduced or increased according to
changes in engine speed and road conditions.
Torque converter performance characteristics have
been discussed in Chapter 3.
The actual speed at which gear ratio changes
occur is provided by hydraulic pressure signals
supplied by the governor valve and a throttle
valve. The former senses vehicle speed whereas
the latter senses engine load.
These pressure signals are directed to a hydraulic
control block consisting of valves and pistons which
compute this information in terms of pressure
variations. The fluid pressure supplied by a pressure
pump then automatically directs fluid to the
various operating pistons causing their respective
clutch, clutches or band brakes to be applied.
Consequently, gear upshifts and downshifts are
performed independently of the driver and are so
made that they take into account the condition of
the road, the available output of the engine and the
requirements of the driver.
siderable and the large gear ratio steps of the conventional transmission are reduced and smoothed
out by the converter's response between automatic
gear shifts. Studying Fig. 5.1, whilst in first gear, the
torque converter provides a maximum torque multiplication at stall pull away conditions which progressively reduces with vehicle speed until the
converter coupling point is reached. At this point,
the reaction member freewheels. With further speed
increase, the converter changes to a simple fluid
coupling so that torque multiplication ceases. In
second gear the converter starts to operate nearer
the coupling point causing it to contribute far less
torque multiplication and in third and fourth gear
the converter functions entirely beyond the coupling
point as a fluid coupling. Consequently, there is no
further torque multiplication.
5.2 Four speed and reverse longitudinally
mounted automatic transmission mechanical
power flow (Fig. 5.2)
(Similar gear trains are adopted by some ZF,
Mercedes-Benz and Nissan transmissions)
The epicyclic gear train is comprised of three planetary gear sets, an overdrive gear set, a forward
gear set and a reverse gear set. Each gear set consists of an internally toothed outer annular ring
gear, a central externally toothed sun gear and a
planet carrier which supports three intermediate
planet gears. The planet gears are spaced evenly
between and around the outer annular gear and
the central sun gear.
The input to the planetary gear train is through
a torque converter which has a lock-up clutch.
Different parts of the gear train can be engaged
or released by the application of three multiplate
clutches, two band brakes and one first gear one
way roller clutch.
Table 5.1 simplifies the clutch and brake engagement sequence for each gear ratio.
A list of key components and abbreviations used
are as follows:
5.1.1 The torque converter (Fig. 5.1)
The torque converter provides a smooth automatic
drive take-up from a standstill and a torque multiplication in addition to that provided by the normal
mechanical gear transmission. The performance
characteristics of a hydrokinetic torque converter
incorporated between the engine and the gear train
is shown in Fig. 5.1 for light throttle and full throttle
maximum output conditions over a vehicle speed
range. As can be seen, the initial torque multiplication when driving away from rest is con-
Manual valve
Vacuum throttle valve
Governor valve
Pressure regulating valve
Torque converter
Fig. 5.1 Torque multiplication and transmitted power performance relative to vehicle speed for a typical four speed
automatic transmission
1±2 shift valve
2±3 shift valve
3±4 shift valve
Converter check valve
Drive clutch
High and reverse multiplate clutch
Forward clutch
Overdrive band brake
Second gear band brake
Low and reverse multiplate brake
First gear one way roller clutch
Torque converter one way clutch
Parking lock
put shaft and pinion carrier. Torque is then split
between the overdrive annular gear and the sun
gear, both paths merging due to the engaged direct
clutch. Therefore the overdrive pinion gears are
prevented from rotating on their axes, causing the
overdrive gear set to revolve as a whole without any
gear ratio reduction at this stage. Torque is then
conveyed from the overdrive annular gear to the
intermediate shaft where it passes through the
applied forward clutch plates to the annular gear
of the forward gear set. The clockwise rotation of
the forward annular gear causes the forward planet
gears to rotate clockwise, driving the double sun
gear counter clockwise. The forward planetary carrier is attached to the output shaft so that the planet
gears drive the sun gear instead of walking around
the sun gear. This anticlockwise rotation of the sun
gear causes the reverse planet gears to rotate
(H ‡ R)C
(L ‡ R)B
5.2.1 D drive range Ð first gear
(Figs 5.3(a) and 5.4(a))
With the selector lever in D range, engine torque is
transmitted to the overdrive pinion gears via the out118
Fig. 5.2 Longitudinally mounted four speed automatic transmission layout
Table 5.1 Clutch and brake engagement sequence
High and
reverse clutch
(H ‡ R) C
Second gear
band brake
Low and
brake (L ‡ R)B
One way
P and N
First D
Second D
Third D
Fourth D
Reverse R
Fig. 5.3 (a±e) Four speed and reverse automatic transmission for longitudinally mounted units
Fig. 5.3 contd
clockwise. With the one way roller clutch holding
the reverse planet carrier, the reverse planetary gears
turn the reverse annular gear and output shaft clockwise in a low speed ratio of approximately 2.46:1.
multiplate clutch and the forward multiplate
clutch, both of which are applied. Subsequently,
the high and reverse clutch will rotate the double
sun gear clockwise and similarly the forward clutch
will rotate the forward annular gear clockwise.
This causes both external and internal gears on
the forward gear set to revolve in the same direction at similar speeds so that the bridging planet
gears become locked and the whole gear set therefore revolves together as one. The output shaft
drive via the reverse carrier therefore turns clockwise with no relative speed reduction to the input
shaft, that is as a direct drive ratio 1:1.
5.2.2 D drive range Ð second gear
(Figs 5.3(b) and 5.4(b))
In D range in second gear, both direct and forward
clutches are engaged. At the same time the second
gear band brake holds the double sun gear and
reverse pinion carrier stationary.
Engine torque is transmitted through the locked
overdrive gear set similarly to first gear. It is then
conveyed through the applied forward clutch via
intermediate shaft to the forward annular gear.
With the double sun gear held by the applied second
gear band brake, the clockwise rotation of the
forward annular gear compels the pinion gears to
rotate on their own axes and roll `walk' around the
stationary sun gear in a clockwise direction.
Because the forward pinion gear pins are mounted
on the pinion carrier, which is itself attached to the
output shaft, the output shaft will be driven clockwise at a reduced speed ratio of approximately
5.2.4 D drive range Ð fourth or overdrive gear
(Figs 5.3(d) and 5.4(d))
In D range in fourth gear, the overdrive band brake,
the high and reverse clutch and the forward clutch
are engaged. Under these conditions, torque is conveyed from the input shaft to the overdrive carrier,
causing the planet gears to rotate clockwise around
the held overdrive sun gear. As a result, the overdrive annular gear will be forced to rotate clockwise but at a higher speed than the input overdrive
carrier. Torque is then transmitted via the intermediate shaft to the forward planetary gear set
which are then locked together by the engagement
of the high and reverse clutch and the forward
clutch. Subsequently, the gear set is compelled to
rotate bodily as a rigid straight through drive. The
torque then passes from the forward planet carrier
to the output shaft. Hence there is a gear ratio step
up by the overdrive planetary gear set of roughly
30%, that is, the output to input shaft gear ratio is
about 0.7:1.
5.2.3 D drive range Ð third or top gear
(Figs 5.3(c) and 5.4(c))
With the selector lever in D range, hydraulic line
pressure will apply the direct clutch, high and
reverse clutch and forward clutch.
As for first and second gear operating conditions, the engine torque is transmitted through the
locked overdrive gear set to the high and reverse
Fig. 5.4 (a±e) Four speed and reverse epicycle gear set directional motion
5.2.5 R range Ð reverse gear
(Figs 5.3(e) and 5.4(e))
With the selector lever in reverse position all three
clutches and the low and reverse multiplate brake
are engaged. Subsequently, engine torque will be
transmitted from the input shaft through the locked
overdrive gear set through the locked forward gear
set via the intermediate shaft to the reverse sun gear
in a clockwise direction.
Because the reverse planet carrier is held by the
low and reverse multiplate brake, the planet gears
are forced to rotate counterclockwise on their axes,
and in doing so compel the reverse annular gear to
also rotate counterclockwise. As a result, the output shaft, which is attached to the reverse annular
gear, rotates counterclockwise, that is, in the
reverse direction, to the input shaft at a reduction
ratio of approximately 2.18:1.
the various valves and to energize the clutch and
band servo pistons will vary under different working conditions. Therefore the fluid pressure generated by the pump is unlikely to suit the many
operating requirements. To overcome these difficulties, a pressure regulating valve is used which
automatically adjusts the pump's output pressure
to match the working requirements at any one
time. One of the functions of the pressure regulating valve is to raise the line pressure reaching the
clutch and brake when the vehicle is driven hard
with large throttle opening to prevent the friction
surfaces slipping. Conversely under light loads and
with a small throttle opening, a much lower line
pressure is adequate to clamp the friction plates or
bands. By reducing the line pressure, fierce clutch
and brake engagements are eliminated which promotes smooth and gentle gear changes. Power consumption, which is needed to drive the hydraulic
pump, is also reduced as actuating pressures are
lowered. The pressure regulating valve is normally
a spring-loaded spool type valve, that is, a plunger
with one or more reduced diameter sections
along its length, positioned in a cylinder which
has a number of passages intersecting the cylinder
When the engine speed, and correspondingly
pump pressure, is low, fluid flows via the inlet
port around the wasted section of the plunger and
out unrestricted along a passage leading to the
manual valve where it is distributed to the various
control valves and operating pistons. As the pump
pressure builds up with rising engine speed, line
pressure is conveyed to the rear face of the plunger
and will progressively move the plunger forward
against the control spring, causing the middle land
to uncover an exhaust port which feeds back to the
pump's intake. Hence as the pump output pressure
tends to rise, more fluid is passed back to the suction intake of the pump. It therefore regulates the
output fluid pressure, known as line pressure,
according to the control spring stiffness. To enable
the line pressure to be varied to suit the operating
conditions, a throttle pressure is introduced to the
spring end of the plunger which opposes the line
pressure. Increasing the throttle pressure raises line
pressure and vice versa.
In addition to the main pressure regulating valve
there is a secondary regulating valve which limits
the fluid flowing through to the torque converter.
Raising the torque converter's fluid pressure
increases its torque transmitting capacity which is
desirable when driving in low gear or when the
engine is delivering its maximum torque.
5.3 The fundamentals of a hydraulic control
The effective operation of an automatic transmission relies upon a hydraulic control circuit to
actuate the gear changes relative to the vehicle's
road speed and acceleration pedal demands with
the engine delivering power. Only a very small
proportion of a transmission's operating time is
spent in performing gear changes. In fact, the
hydraulic system is operational for less than 1%
of the driving time. The transition time from one
gear ratio to the next takes roughly one second or
less and therefore the hydraulic control valves must
be designed to direct fluid pressure to the appropriate operating pistons which convert the fluid
pressure into mechanical force and movement to
energize the respective clutches and band brakes
instantly and precisely.
An understanding of a basic hydraulic control
system can best be considered under the four
Pressure supply and regulating valves
Speed and load sensing valves
Gear shift valves
Clutch and brake coupling and hold devices
5.3.1 Pressure supply and regulating valve
(Fig. 5.5)
The essential input to the hydraulic control system
is fluid pressure generated by a pump and driven by
the engine. The pump's output pressure will
increase roughly in proportion to the engine's
speed. However, the pressure necessary to actuate
5.3.2 Speed and load sensing valves
(Figs 5.5 and 5.6)
For gear changes to take place effectively at the
optimum engine and road speed, taking into
account the driver's demands expressed in throttle
opening, some means of sensing engine load and
vehicle road speed must be provided. Engine output torque is simply monitored by a throttle valve
which is linked to the accelerator pedal, either
directly or indirectly, via a vacuum diaphragm
operated linkage which senses the change in induction depression, which is a measure of the engine
load. The amount the accelerator pedal or manifold vacuum alters is relayed to the throttle valve
which accordingly raises or lowers the output pressure. This is then referred to as throttle pressure.
Road speed changes are measured by a centrifugal force-sensitive regulating valve which senses
transmission output shaft speed and transmits this
information in the form of a fluid pressure, referred
to as governor pressure, which increases or decreases
according to a corresponding variation in road speed.
Both throttle pressure and governor pressure are
signalled to each gear shift valve so that these may
respond to the external operating conditions
(i.e. engine torque developed and vehicle speed)
by permitting fluid pressure to be either applied or
released from the various clutch and brake actuating
piston chambers.
opposing end forces acting on the spool valve end
pressure load
pressure load
But PA ˆ F
hence FS ‡ FT ˆ FG
where FS
ˆ Spring load
ˆ Throttle pressure load
ˆ Throttle pressure
ˆ Governor pressure load
ˆ CSA of plunger at throttle pressure
PG ˆ Governor pressure
AG ˆ CSA of plunger at governor pressure end
Thus increasing or decreasing the spring stiffness
or enlarging or reducing the diameter of the spool
valve at one end considerably alters the condition
when the shift valve moves from one end to
the other to redirect line pressure to and from the
various clutch and brakes and so produce the
necessary gear change.
Each shift valve control spring will have a particular stiffness so that different governor pressures,
that is, road speeds, are required to cause either
a gear upshift or downshift for a given opposing
throttle pressure. Conversely, different engine
power outputs will produce different throttle pressures and will alter the governor pressure accordingly when a particular gear shift occurs. Large
engine loads (high throttle pressure) will delay
gear upshifts whereas light engine load demands
(low throttle pressure) and high vehicle speeds
(high governor pressure) will produce early
upshifts and prevent early downshift.
To improve the quality of the time sequence
of up or down gear shift, additional valves and
components are included to produce a smooth
transition from one gear to the next. Some of
these extra devices are described in Section 5.6.
5.3.3 Gear shift valves (Fig. 5.5)
Shift valves are of the spool plunger type, taking
the form of a cylindrical plunger reduced in diameter in one or more sections so as to divide its
length into a number of lands. When operating,
these valves shift from side to side and cover or
uncover passages leading into the valve body so
that different hydraulic circuits are switched on
and off under various operating conditions.
The function of a shift valve is to direct the fluid
pressure to the various clutch and brake servo pistons to effect gear changes when the appropriate
load and speed conditions prevail. Shift valves are
controlled by line or throttle pressure, which is
introduced into the valve at the spring end, and
governor pressure, which is introduced directly to
the valve at the opposite end. Generally, the governor valve end is of a larger diameter than the spring
end so that there will be a proportionally greater
movement response due to governor pressure variation. Sometimes the shift valve plunger at the governor pressure end is referred to as the governor plug.
The position of the shift valve at any instant
depends upon the state of balance between the
5.3.4 Clutch and brake coupling and hold devices
(Figs 5.5 and 2.16)
Silent gear change synchronization is made possible by engaging or locking out various members of
the epicyclic gear train gear sets with the engine's
power being transmitted continuously. It therefore
requires a rapid and accurate gear change which is
achieved by utilizing multiplate clutches and band
brakes. A gear up- or downshift therefore occurs
with the almost simultaneous energizing of one
Fig. 5.5 (a and b)
Basic multiplate clutch and band brake transmission hydraulic control system
clutch or brake and a corresponding de-energizing
of another clutch or brake.
one shaft or member to another quickly and
smoothly. The rotating and fixed friction plates
can be energized by an annular shaped, hydraulically operated piston either directly or indirectly by
a dished washer which acts also as a lever to multiply the operating clamping load. Return springs
are used to separate the pairs of rubbing faces
when the fluid pressure is released. Wear and
Multiplate clutch (Figs 2.16 and 5.5) Wet multiplate type clutches are very compact for their torque transmitting and heat dissipating capacity.
They are used to lock any two members of a planetary gear set together or to transfer drive from
adjustment of the friction plate pack is automatically compensated by the piston being free to move
further forward (see Chapter 2, Fig. 2.16).
With rising output shaft speed, the centrifugal
force acting through the primary valve is sufficient
to overcome the hydraulic line pressure, which is
acting against the shouldered groove face area and
will therefore progressively move outwards as the
rotational speed increases until the valve borders
on an end stop. The opening of the governor valve
outlet passage now allows fluid to flow out from
the governor, where it is then directed to the large
diameter end of the shift valve. This output pressure is known as governor pressure. With even
higher rotational output shaft speed (vehicle
speed), greater centrifugal force will be imposed
on the secondary valve until it is able to overcome
the much larger hydraulic inward load imposed on
the large shoulder of this valve. The secondary
valve will start to move out from the centre of
rotation, uncovering the secondary valve outlet
passage so that increased governor pressure passes
to the shift valve.
This two stage governor valve action enables the
governor to be more sensitive at the very low
speeds but not oversensitive at the higher speeds
(Fig. 5.5(c)). Sensitivity refers to the amount of
fluid pressure increase or decrease for a unit change
in rotational speed. If there is a large increase or
decrease in governor pressure per unit charge in
speed, then the governor is sensitive. If there is
very little variation in governor pressure with a change
in rotational speed (i.e. vehicle speed), then the
governor is insensitive and therefore not suitable
for signalling speed changes to the hydraulic
control systems.
The reason a single stage governor would not perform satisfactorily over the entire output shift speed
range is due to the centrifugal force square law: at
low speeds the build-up in centrifugal force for
a small increase in rotational speed is very small,
whereas at higher speeds only a small rise in speed
produces a considerable increase in centrifugal
force. If the governor has the correct sensitivity at
high speed it would be insensitive at low speed or if it
has the desired sensitivity at low speed it would be
far too responsive to governor pressure changes in
the higher speed range.
Once the governor pressure end load (PG AG )
equals the spring and throttle pressure load
(FS ‡ PT AT ) with rising vehicle speed, any
further speed increase will push the shift valve
plunger towards the spring end to the position
shown in Fig. 5.5(a). The fluid on the applied side
of the band brake servo piston will now exhaust
(drain) through the shift valve to the inlet side of
the oil pump. Simultaneously, line pressure from
Band brake (Fig. 5.5) This form of brake consists
of a friction band encompassing an external drum
so that when the brake is applied the band contracts, thereby wrapping itself tightly around the
drum until the drum holds. The application of the
band is achieved through a double acting stepped
servo cylinder and piston. Fluid line pressure is
introduced to the small diameter end of the piston
to energize the band brake. To release the band,
similar line pressure is directed to the spring chamber side of the cylinder. Band release is obtained due
to the larger piston area side producing a greater
force to free the band. This method of applying and
releasing the band enables a more prolonged and
controllable energizing and de-energizing action to
be achieved. This class of brake is capable of absorbing large torque reactions without occupying very
much space, which makes the band brake particularly suitable for low gear high torque output gear
sets. Band wear slackness can be taken up by externally adjusting the anchor screw.
5.4 Basic principle of a hydraulically controlled
gearshift (Fig. 5.5)
Selecting the drive D range positions the manual
valve spool so that line pressure from the pressure
regulator valve passes through to the shift valve,
throttle valve and governor valve (Fig. 5.5(a)).
Throttle pressure will be introduced to the spring
end of the shift valve via the throttle valve. Depressing the accelerator pedal allows the spool valve to
move outwards. This increases the valve opening so
that a high throttle pressure will be delivered to the
shift valve. Conversely, depressing the accelerator
pedal partially restricts the flow of fluid and therefore reduces the throttle pressure reaching the shift
valve (Fig. 5.5(b)).
At the same time, line pressure enters the governor valve, flows between the wasted region of both
primary and secondary spool valves and reacts
against the difference in the annular adjacent face
areas of each spool valve. Both valves are forced
inwards, covering up the two exits from the governor valve housing. As the vehicle moves forwards,
the rotation of the governor causes a centrifugal
force to act through the mass of each governor
valve so that it tends to draw the valve spools outwards in opposition to the hydraulic pressure
which is pushing each valve inwards (Fig. 5.5(a)).
the manual valve is directed via the shift valve to
both the release side of the band servo piston and
to the multiplate clutch piston which then energizes
the friction plates.
Supply fluid to the spring side of the servo piston
(known as the release side), provides a more progressive and controllable transition from one gear
change to another which is not possible when
relying only on the return spring.
When the vehicle's speed is reduced or the throttle pressure is raised sufficiently, the shift valve
plunger will move to the governor pressure end of
the valve (Fig. 5.5(a)). The line pressure transmitted to the shift valve is immediately blocked
and both the multiplate clutch and the band
brake hydraulic feed passages are released of fluid
pressure by the middle plunger land uncovering the
exhaust part. Simultaneously, as the same middle
land covers the right hand exhaust port and
uncovers the line pressure passage feeding from
the manual valve, fluid will flow to the applied
side of the band servo piston, causing the band to
contract and so energize the brake.
shift valve spring end is subjected to line pressure
from the manual valve.
Whilst the transmission is in drive first gear the
one way clutch will engage, so preventing the
reverse planetary carrier from rotating (not shown
in hydraulic system).
5.5.2 Second gear (Fig. 5.7)
With the manual valve still in D, drive position,
hydraulic conditions will be similar to first gear,
that is, the overdrive and forward clutches are
engaged, except that rising vehicle speed increases
the governor pressure sufficiently to push the 1±2
shift valve against both spring and line pressure end
loads. As a result, the 1±2 shift valve middle land
uncovers the line pressure supply passage feeding
from the manual valve. Line pressure is now directed to the second gear band servo on the applied
side, energizing the second gear brake and causing
both the forward and reverse sun gears to hold.
If there is a reduction in vehicle speed or if the
engine load is increased sufficiently, the resulting
imbalance between the spring and throttle pressure
load as opposed to governor pressure acting on the
1±2 shift valve at opposite ends causes the shift
valve to move against the governor pressure. Consequently the hydraulic circuitry will switch back
to first gear conditions, causing the transmission
to shift down from second to first gear again.
5.5 Basic four speed hydraulic control system
A simplified hydraulic control system for a four
speed automatic transmission will now be examined
for the reader to obtain an appreciation of the overall
function of the hydraulic computer (control) system.
5.5.1 First gear (Fig. 5.6)
With the manual valve in D, drive position, fluid is
delivered from the oil pump to the pressure regulating valve. It then divides, some being delivered
to the torque converter, the remainder passing out
to the manual valve as regulating pressure (more
commonly known as line pressure). Line pressure
from the manual valve is then channelled to the
forward clutch, which is energized, and to the
overdrive band servo on the applied side. At the
same time, line pressure from the pressure regulating valve passes through the 3±4 shift valve where it
is directed to energize the drive clutch and to the
released side of the overdrive band servo, thus preventing the engagement of the band. Line pressure
is also directed to both the governor valve and to
the vacuum throttle valve. The reduced pressure
output from the governor valve which is known
as governor pressure is directed to the end faces
of each of the three shaft valves, whereas the output
pressure from the throttle valve, known as throttle
pressure, is conveyed to the spring end of the 2±3
and 3±4 shift valves. On the other hand, the 1±2
5.5.3 Third gear (Fig. 5.8)
At even higher road speeds in D, drive position, the
governor pressure will have risen to a point where it
is able to overcome the spring and throttle pressure
load of the 2±3 shift valve. This causes the spool
valve to shift over so that the line pressure passage
feed from the manual valve is uncovered. Line
pressure will now flow through the 2±3 shift valve
where it is directed to the high and reverse clutch to
energize the respective fixed and rotating friction
plates. At the same time, line pressure passes to
the second gear band servo on the release side to
disengage the band. Consequently both overdrive
and forward planetary gear sets lock-up, permitting the input drive from the torque converter to be
transmitted directly through to the transmission's
output shaft.
The actual vehicle speed at which the 2±3 shift
valve switches over will be influenced by the throttle opening (throttle pressure). A low throttle pressure will cause an early gear upshift whereas a large
engine load (high throttle pressure) will raise the
upshift speed.
Fig. 5.6 Hydraulic control system (D) range first gear
5.5.4 Fourth gear (Fig. 5.9)
With still higher road speeds in D, drive position,
the increased governor pressure will actuate the 3±4
shift valve, forcing it to shift across so that it covers
up the line pressure supply passage and at the same
time uncovers the exhaust or drain port. As a result,
the line pressure exhausts from the release side of
the overdrive band servo which then permits the
band to be energized. At the same time the drive
clutch will be de-energized because of the collapse
of line pressure as it is released through the 3±4
shift valve exhaust port.
Fig. 5.7 Hydraulic control system (D) range second gear
Under these operating conditions the overdrive
shaft planetary gear set reduces the intermediate
shift speed and, since the forward clutch is in a
state of lock-up only, this speed step up is transmitted through to the output shaft.
5.5.5 Reverse gear (Fig. 5.10)
With the manual valve in R, reverse position, line
pressure from the manual valve is directed via the
2±3 shift valve to the release side of the second gear
band servo, causing the band to disengage. At the
Fig. 5.8 Hydraulic control system (D) range third gear
same time line pressure from the same supply passage engages the high and reverse clutch. The manual valve also supplies line pressure to the low and
reverse band brake via the 1±2 shift valve to hold
the reverse planetary carrier. In addition, line pressure from the pressure regulating valve output side
is directed via the 3±4 shift valve to the release side
of the overdrive brake servo to disengage the band
and to the drive clutch piston to engage the friction
plates. Note that both band brake servos on the
applied sides have been exhausted of line pressure
and so has the forward clutch piston chamber.
Fig. 5.9 Hydraulic control system (D) range fourth gear
5.5.6 Lock-up torque converter (Fig. 5.11)
between the input pump impeller and the turbine
output shaft. The benefits of this lock-up can only
be realised if the torque converter is allowed to
operate when light torque demands are made on
the engine and only when the converter is operating
above its torque multiplication range that is
beyond the coupling point. Consequently, converter
Introduction To overcome the inherent relative
slip which always occurs between the torque converter's pump impeller and the turbine runner, even
driving at moderate speeds under light load conditions, a lock-up friction clutch may be incorporated
Fig. 5.10 Hydraulic control system (D) reverse gear
lock-up is only permitted to be implemented when
the transmission is in either third or fourth gear.
The advantages of bypassing the power transfer
through the circulating fluid and instead transmitting the engine's output directly to the transmission input shaft eliminates drive slippage,
thereby increasing the power actually propelling
the vehicle. Due to this net gain in power output,
fuel wastage will be reduced.
Lock-up clutch description The lock-up clutch
consists of a sliding drive plate which performs
two functions; firstly to provide the friction coupling device and secondly to act as a hydraulic con132
Fig. 5.11 (a and b)
Lock-up torque converter
trolled piston to energize or de-energize the clutch
engagement facings. The lock-up drive plate/piston
is supported by the turbine hub which is itself
mounted on the transmission input shaft. A transmission damper device, similar to that used on a
conventional clutch drive-plate, is incorporated in
the lock-up plate to absorb and damp shock
impacts when the lock-up clutch engages.
low speeds, low governor pressure permits the
speed cut and lock-up control valve return springs
to push their respective plunger to the right. Under
these conditions, pressurized fluid from the torque
converter flows into the space separating the lockup plate/piston from the turbine. At the same time,
fluid from the oil pump is conveyed to the space
formed between the torque converter's casing and
the lock-up plate/piston via the lock-up control
valve and the central axial passage in the turbine
input shaft. Consequently, the pressure on both sides
of the lock-up plate will be equalized and so the
lock-up plate/piston cannot exert an engagement
load to energize the friction contact faces.
Lock-up control The automatic operation of the
converter lock-up is controlled by a speed cut valve
and a lock-up control valve. The function of these
valves is to open and close fluid passages which
supply and discharge fluid from the space formed
between the torque converter casing and the lockup drive plate/piston.
Lock-up engaged (Fig. 5.11(b)) As the speed of
the vehicle rises, increased governor pressure will
force the speed cut valve plunger against its spring
until it uncovers the line pressure passage leading
into the right hand end of the lock-up control
Lock-up disengaged (Fig. 5.11(a)) With the vehicle driven in either first or second gear at relatively
valve. Line pressure fed from the high and reverse
clutch is directed via the speed cut valve to the right
hand end of lock-up control valve, thereby pushing
its plunger to the left to uncover the lock-up clutch
drain port. Instantly, pressurized fluid from the
chamber created between the torque converter casing and lock-up plate/piston escapes via the central
input shaft passage through the wasted region of the
lock-up control valve plunger back to the inlet side
of the oil pump. As a result, the difference of pressure
across the two sides of the lock-up plate/piston
causes it to slide towards the torque converter casing
until the friction faces contact. This closes the exit
for the converter fluid so that full converter fluid
pressure is exerted against the lock-up plate/piston.
Hence the input and output shafts are now locked
together and therefore rotate as one.
automatic transmission in each gear ratio will
now be considered in some depth, see Fig. 5.12.
The planetary gear train consists of two sun
gears, two sets of pinion gears (three in each set),
two sets of annular (internal) gears and pinion
carriers which support the pinion gears on pins.
Helical teeth are used throughout.
For all forward gears, power enters the gear train
via the forward annular gear and leaves the gear
train by the reverse annular gear. In reverse gear,
power enters the gear train by the reverse sun gear
and leaves the gear train via the reverse annular gear.
First gear compounds both the forward gear set
and the reverse gear set to provide the necessary
low gear reduction. Second gear only utilizes the
forward planetary gear set to produce the intermediate gear reduction. Third gear is achieved by
locking the forward planetary gear set so that
a straight through drive is obtained. With planetary gear trains the gears are in constant mesh and
gear ratio changes are effected by holding, releasing or rotating certain parts of the gear train by
means of a one way clutch, two multiplate clutches,
one multiplate brake and one band brake.
The operation of the automatic transmission
gear train can best be explained by referring to
Table 5.2 which shows which components are
engaged in each manual valve selection position.
Speed cut valve function The purpose of the speed
cut valve is to prevent fluid draining from the space
formed between the converter casing and lock-up
plate/piston via the lock-up control valve if there is
a high governor pressure but the transmission has
not yet changed to third or fourth gear. Under these
conditions, there is no line pressure in the high and
reverse clutch circuit which is controlled by the shift
valve. Therefore when the speed cut valve plunger
moves to the left there is no line pressure to actuate
the lock-up control valve so that the lock-up plate/
piston remains pressurized on both sides in the disengaged position.
5.6.1 Selector lever (Table 5.2)
The selector lever has a number of positions
marked P R N D 2 1 with definite functions as
5.6 Three speed and reverse transaxle automatic
transmission mechanical power flow
(Gear train as adopted by some Austin-Rover, VW
and Audi 1.6 litre cars)
The operating principle of the mechanical power
or torque flow through a transaxle three speed
P Ð park When selected, there is no drive
through the transmission. A mechanical lock actuated by a linkage merely causes a parking pawl to
engage in the slots around a ring gear attached to
the output shaft (Fig. 5.2). Thus the parking pawl
Table 5.2 Manual valve selection position
P and N
D ± 1st
2 ± 1st
1 ± 1st
D ± 2nd
2 ± 2nd
D ± 3rd
Drive and
reverse clutch
(D ‡ R)C
First and
reverse brake
(I ‡ R)B
gear band
One way
Fig. 5.12
Transaxle three speed automatic transmission layout
locks the output shaft to the transmission casing so
that the vehicle cannot roll backwards or forwards.
This pawl must not be engaged whilst the vehicle is
moving. The engine may be started in this position.
The reverse position must only be selected when the
vehicle is stationary. The engine will not start in
reverse position.
N Ð neutral When selected, all clutches and band
brake are disengaged so that there is no drive
through the transmission. The engine may be
started in N Ð neutral range.
R Ð reverse When selected, the output shaft from
the automatic transmission is made to rotate in the
opposite direction to produce a reverse gear drive.
Fig. 5.13 (a±d)
Three speed and reverse automatic transmission transaxle units
Fig. 5.13
D Ð drive This position is used for all normal
driving conditions, automatically producing 1±2,
2±3 upshifts and 3±2, 2±1 downshifts at suitable
road speeds or according to the position of the
accelerator pedal. The engine will not start in
D Ð drive range.
instead of rolling `walking' around the sun gears.
This counterclockwise rotation of the sun gears
causes the reverse planet gears to rotate clockwise.
With the one way clutch holding the reverse planet
carrier stationary, the reverse planetary gears turn
the reverse annular gear and output shaft clockwise
in a reduction ratio of something like 2.71:1.
When first gear is selected in the D range, a very
smooth transmission take-up is obtained when the
one way clutch locks, but on vehicle overrun the
one way clutch is released so that the transmission
2 Ð First and second This position is selected
when it is desired to restrict gear changes automatically from 1±2 upshift and 2±1 downshifts only.
The selector must not be positioned in 2 range
above 100 km/h (70 mph). The engine will not
start in this range position.
5.6.3 First gear manual (1 Ð 1st) (Fig. 5.13(a))
The power flow in first gear manual differs from
the D range in that the first and reverse brake are
applied to hold the reverse planet carrier stationary. Under these conditions on vehicle overrun,
engine braking is provided.
1 Ð First gear When this range is selected, the
transmission is prevented from shifting into second
and third gear. A friction clutch locks out the one
way roller clutch so that better control may be
obtained when travelling over rough or wet ground
or icy roads. Engine braking on overrun is available when descending steep hills.
5.6.4 Second gear (D Ð 2nd) (Fig. 5.13(b))
In D range in second gear, the forward clutch and
the second gear band brake are applied. The forward clutch then transmits the engine torque from
the input shaft to the forward annular gear and the
second gear band brake holds the double sun gear
stationary. Thus engine torque is delivered to the
annular gear of the forward planetary train in a
clockwise rotation. Consequently, the planet gears
are compelled to revolve on their axes and roll
`walk' around the stationary sun gear in a clockwise direction. As a result the output shaft, which is
splined to the forward planet carrier, is made to
turn in a clockwise direction at a slower speed
5.6.2 First gear (D Ð 1st) (Fig. 5.13(a))
With the manual selector valve in D range, engine
torque is transmitted from the converter through the
applied forward clutch to the annular gear of the
forward planetary gear train. The clockwise rotation
of the forward annular gear causes the forward
planet gears to rotate clockwise, driving the double
(compound) sun gear anticlockwise. The forward
planetary carrier is splined to the output shaft. This
causes the planet gears to drive the double sun gear
relative to the input shaft with a reduction ratio of
approximately 1.50:1.
5.6.5 Third gear (D Ð 3rd) (Fig. 5.13(c))
In D range engine torque is transmitted through
both forward clutch and drive and reverse clutch.
The drive and reverse clutch rotate the sun gear of
the forward gear train clockwise and similarly the
forward clutch turns the annular gear of the same
gear set also clockwise. With both the annular gear
and sun gear of the forward gear train revolving in
the same direction at the same speed, the planet
gear becomes locked in position, causing the forward gear train to revolve as a whole. The output
shaft, which is splined to the forward planet carrier,
therefore rotates at the same speed as the input
shaft, that is as a direct drive ratio 1:1.
5.6.6 Reverse gear (R) (Fig. 5.13(d))
With the manual selector valve in the R position,
the drive and reverse multiplate brake is applied to
transmit clockwise engine torque to the reverse
gear set sun gear. With the first and reverse brake
applied, the reverse planet gear carrier is held stationary. The planet gears are compelled to revolve
on their own axes, thereby turning the reverse
annular gear which is splined to the output shaft
in an anticlockwise direction in a reduction ratio of
about 2.43:1.
Manual valve
Kickdown valve
Throttle pressure valve
Valve for first gear manual
V(1 ‡ R)GB
V(D ‡ R)C
(D ‡ R)CP
(1 ‡ R)BP
5.7.1 The pressure supply system
This consists of an internal gear crescent oil pump
driven by the engine via a shaft splined to the
torque converter impeller. The oil pressure generated by the oil pump is directed to the pressure
regulating valve. By introducing limited throttle
pressure into the regulator valve spring chamber,
the thrust acting on the left hand end of the valve is
increased during acceleration. This prevents the
regulator valve being pushed back and spilling oil
into the intake side of the oil pump. As a result,
the line pressure will rise as the engine speed
5.7 Hydraulic gear selection control components
(Fig. 5.24)
(Three speed and reverse transaxle automatic
An explanation of how the hydraulic control system is able to receive pressure signals which correspond to vehicle speed, engine load and the driver's
requirements, and how this information produces
the correct up or down gear shift through the
action of the control system's various plunger
(spool) valves will now be considered by initially
explaining the function of each component making
up the control system.
A list of key components and abbreviations used
in the description of the hydraulic control system is
as follows:
1±2 shift valve
1±2 governor plug
Throttle pressure limiting valve
Main pressure limiting valve
Main pressure regulating valve
Valve for first and reverse gear
Converter pressure valve
Soft engagement valve
2±3 shift valve
2±3 governor plug
Valve for direct and reverse
3±2 control valve
3±2 kickdown valve
Governor valve
Forward clutch piston
Oil pump
Converter check valve
Second gear band servo
Forward clutch piston
Direct and reverse clutch piston
First and reverse brake piston
One way clutch
5.7.2 Main pressure regulator valve (MPRV)
(Fig. 5.14(a and b))
This valve controls the output pressure which is
delivered to the brake band, multiplate brake and
clutch servos. Oil pressure from the pump acts on
the left hand end of the valve and opposes the
return spring. This oil pressure moves the valve to
the right, initially permitting oil to pass to the converter pressure valve and its circuit, but with
further valve movement oil will be exhausted back
to the pump intake passage. The line pressure
build-up is also controlled by introducing limited
throttle pressure into the regulator spring chamber
Fig. 5.14 (a and b)
Main pressure regulating valve (MPRV)
which assists the spring in opposing the valve moving to the right. In addition, oil pressure from the
manual valve passage, indirectly controlled by the
governor, is imposed on the left hand end of the
regulator valve. This modifies the valve movement
to suit the various gear train and road condition
which acts on the other end of the shift valves
controlling upshift and downshift speeds.
5.7.4 Main pressure limiting valve (MPLV)
(Fig. 5.16)
This valve is designed to limit or even cut off the
variable throttle pressure passing through to the
main regulating valve and the soft engagement
valve. The pressure passing out from the valve to
the main pressure regulator valve is known as
limited throttle pressure. As the pressure passes
through the valve it reacts on the left hand end of
the main pressure limiting valve so that the valve
will progressively move to the right, until at some
predetermined pressure the valve will close the
throttle pressure port feeding the main pressure
regulating valve circuit.
When the throttle pressure port closes, the high
pressure in the regulator spring chamber is
permitted to return to the throttle pressure circuit
via the non-return ball valve.
5.7.3 Throttle pressure valve (TPV)
(Fig. 5.15(a and b))
The throttle pressure valve transmits regulated
pressure based on engine throttle position. Opening or closing the engine throttle moves the kickdown valve spool so that the throttle valve spring
tension is varied. The amount of intermediate pressure allowed through the throttle pressure valve is
determined by the compression of the spring. The
reduced pressure on the output side of the throttle
valve is then known as throttle pressure. Throttle
pressure is directed to the main pressure limiting
valve, the kickdown valve, and to one end of the
shift valves in opposition to governor pressure,
Fig. 5.15 (a and b)
Kickdown valve (KDV), throttle pressure valve (TPV) and valve for first gear manual range (1GMR)
Fig. 5.16 Main pressure limiting valve
5.7.5 Converter pressure valve (CPV) (Fig. 5.17)
This valve shuts off the oil supply to the torque
converter once the delivery pressure reaches 6 bar.
Line pressure from the main pressure regulator
valve passes through the valve to the torque converter and acts on its right hand end until the preset pressure is reached. At this point the valve is
pushed back against its spring, closing off the oil
supply to the torque converter until the converter
pressure is reduced again.
The force on the output side of the converter
pressure valve feeding into the converter is known
as converter pressure.
ate pressure then passes to the throttle pressure
5.7.8 Kickdown valve (KDV)
(Fig. 5.15(a and b))
This valve permits additional pressure to react on
the shift valves and governor plugs when a rapid
acceleration (forced throttle) response is required
by the driver so that the governor pressure is compelled to rise to a higher value before a gear upshift
occurs. When the throttle is forced wide open, the
kickdown valve is moved over to the right, thus
allowing throttle pressure to pass through the
valve. The output pressure is known as kickdown
pressure. The kickdown pressure feeds in between
both 1±2 and 2±3 shift valves and governor plug
combinations. As a result, this kickdown pressure
opposes and delays the governor pressure movement of the governor plug and shift valve, thereby
preventing a gear upshift occurring until a much
higher speed is reached.
5.7.6 Converter check valve (CCV)
This valve, which is located inside the stator support, prevents the converter oil drainage when the
vehicle is stationary with the engine switched off.
This valve is not shown in the diagrams.
5.7.7 Throttle pressure limiting valve (TPLV)
(Fig. 5.18)
This valve converts line pressure, supplied by the
pump and controlled by the main pressure regulator valve, into intermediate pressure. The pressure
reduction is achieved by line pressure initially
passing through the diagonal passage in the valve
so that it reacts against the left hand end of the
valve. Consequently the valve shifts over and partially reduces the line pressure port opening. The
reduced output pressure now known as intermedi-
5.7.9 1±2 Shift valve and governor plug (1±2)SV
and (1±2)GP (Fig. 5.19)
This valve combination automatically controls and
shifts the transmission from first to second or from
second to first depending upon governor and
throttle pressure. When governor pressure on the
right hand governor plug side overcomes throttle
pressure on the left hand 1±2 shift valve side, both
Fig. 5.17 Converter pressure valve (CPV)
Fig. 5.18
Throttle pressure limiting valve (TPLV)
Fig. 5.19
1±2 shift valve (1±2)SV, and 1±2 governor plug (1±2)GP in 1±2 upshift condition
1±2 governor plug and 1±2 shift valve move to the
left thereby opening the line pressure port which
delivers oil from the pump. Line pressure will now
pass unrestricted through the valve to feed into the
brake band servo. As a result an upchange occurs.
If, in addition to the throttle pressure, kickdown
pressure is introduced to the valve combination,
gear upshifts will be prolonged. If `1' manual
valve is selected, line pressure will be supplied to
the governor plug chamber (large piston area) and
the throttle spring chamber, preventing a 1±2
upshift. `1' manual position cannot be engaged at
speeds above 72 km/h because the 1±2 shift valve
cannot move across, due to the governor pressure.
5.7.10 2±3 Shift valve and governor plug
(2±3)SV and (2±3)GP (Fig. 5.20(a and b))
The 2±3 shift valve and governor plug control the
gear change from second to top gear or from top to
second depending upon governor and throttle pressure. As governor pressure exceeds throttle pressure, the shift valve and governor plug are pushed
over to the left. This permits line pressure to pass
through the valve so that it can supply pressure to
the drive and reverse clutch piston, so that an
upchange can now take place. When `2' manual
valve position is selected, there is no pressure feeding to the shift valve which therefore prevents a 2±3
Fig. 5.20 (a and b) 2±3 shift valve (2±3)SV, 2±3 governor plug (2±3)GP, 3±2 control valve (3±2)CV, 3±2 kickdown valve
(3±2)KDV and valve for direct and reverse clutch V(D ‡ R)C
5.7.11 3±2 Kickdown valve (3±2)KDV
(Fig. 5.20(a and b))
This valve is provided to prolong the downshift
from third to second gear during rapid acceleration
from above 90 km/h so that the change takes place
relatively smoothly. With rising output shaft speed,
the governor pressure acting on the right hand end
of the valve moves it to the right, thus practically
restricting the oil outflow from the servo spring
chamber and therefore extending the second gear
band engagement time.
to the 3±2 control valve. As the vehicle speed
approaches 60 km/h the governor pressure rises sufficiently to force back the 3±2 control valve piston,
thus causing the wasted (reduced diameter) part of
the control valve to complete the exhaustion of oil.
5.7.13 Valves for direct and reverse clutch
V(D ‡ R)C (Fig. 5.21(a and b))
When the manual selector valve is moved to reverse
position the left hand ball valve drops onto its seat
so that line pressure oil from the manual selector is
compelled to move through a restriction. At the
same time, the right hand ball valve is pushed to
the right, immediately closing off the second gear
servo piston spring side from the line pressure. The
right hand ball valve is dislodged to the left when
third gear is selected so that the manual reverse line
pressure is prevented feeding the drive and reverse
clutch piston. During the time third gear is selected
the left hand ball serves no purpose.
5.7.12 3±2 Control valve (3±2)CV
(Fig. 5.20(a and b))
This valve controls the expulsion of oil from the
spring side of the second gear band servo piston at
speeds in the region of 60 km/h. The time period for
oil to exhaust then depends upon the governor
pressure varying the effective exhaust port restriction. Line pressure oil from the spring side of the
second gear band servo piston passes through a
passage leading to the 3±2 kickdown valve annular
groove and from there to the 2±3 shift valve annular groove. Here some oil exhausts out from a fixed
restriction while the remainder passes via a passage
Fig. 5.21 (a and b)
5.7.14 Valve for first gear manual range
V(1G)MR (Figs 5.15(a and b) and 5.22)
The selection of first gear manual supplies line pressure to the underside passage to the ball valve,
Soft engagement valve (SEV), valve for first and reverse clutch, V(1 ‡ R)GC
Fig. 5.22 Manual valve (MV), kickdown valve (KDV), throttle pressure valve (TPV), 1±2 shift valve (1±2)SV and 1±2
governor plug (1±2)GP in first gear ± manual selection
causing it to move to the right. Line pressure then
fills the throttle pressure lines leading to the left
hand end of the 1±2 shift valve and therefore a 1±2
upshift is prevented.
against the opposing variable throttle pressure.
The result of this movement is to restrict and slow
down the pressure build-up on the first and reverse
gear brake piston.
5.7.15 Valves for first and reverse gear brake
V(1 ‡ R)GB (Fig. 5.20(a and b))
The selection of first gear manual position causes
line pressure to dislodge the right hand ball valve to
the left, thereby closing the reverse line passage
from the selector valve. At the same time the left
hand ball valve closes so that line pressure flow for
the engagement of the first and reverse multiplate is
slowed down.
The selection of reverse gear causes the right
hand ball valve to be pushed by line pressure to
the right and so the first gear line passage from the
selector valve is closed. Similarly the left hand ball
valve closes so that line pressure flow to the first
and reverse brake is restricted, thus prolonging the
clutch engagement period.
5.7.17 Clutches and brakes
Front clutch piston (FCP) (Fig. 5.24) This is an
annular shaped piston which directs a clamping
load to a multiplate clutch via a diaphragm type
spring when line pressure is introduced behind the
piston. The engagement of the clutch couples the
output shaft from the torque converter turbine to
the forward annular gear ring. The forward clutch
is applied in all forward drive gear ranges.
First and reverse brake piston (1 ‡ R)CP (Fig.
5.24) Introducing line pressure to the first and
reverse clutch piston cylinder engages the multiplate brake which locks the reverse planetary carrier to the transmission casing. The first and reverse
brake is applied only in first and reverse range.
5.7.16 Soft engagement valve (SEV)
(Fig. 5.21(a and b))
This valve provides a cushioning effect for the
engagement of first and reverse gear brake. This
effect is achieved by line pressure acting on the left
hand valve end pushing the valve to the right
Drive and reverse clutch piston (D ‡ R)CP
(Fig. 5.24) Directing line pressure behind the
drive and reverse clutch piston applies the clutch,
thereby transmitting torque from the torque converter turbine output shaft to the forward sun gear.
When the forward clutch is also applied, the forward planetary gears (annular, planet and sun
gears) are locked together and they rotate bodily,
thus producing a straight through 1:1 third gear
drive. However, when the first and reverse clutch
is applied instead of the forward clutch, the reverse
planetary carrier is held stationary causing the
reverse gear reduction ratio to be engaged.
The timing of the release of one set of gears and
the engagement of another to produce smooth up
and down gear shifts between second and third
gears is achieved by carefully controlling the delivery and exhaustion of hydraulic fluid from the
clutch and band brake servo.
These operating conditions are explained under
second gear band servo.
operates in place of the first and reverse multiplate
brake to prevent the rotation of the reverse pinion
carrier. This one way clutch enables the gear set to
freewheel on overrun and to lock-up on drive, therefore preventing a jerky gear ratio in 1±2 upshift and
2±1 downshift.
5.7.18 The governor valve (GV)
(Figs 5.23 and 5.24)
The governor revolving with the transmission output shaft is basically a pressure regulating valve
which reduces line pressure to a value that varies
with output vehicle speed. This variable pressure is
known as governor pressure and is utilized in the
control system to effect up and down gear shifts
from 1±2 and 2±3 shift valves. Governor pressure
opposes shift valve spring force, throttle pressure
and kickdown pressure, and the resulting force
acting on the governor plug and shift valve determines the vehicle's gear change speeds. The governor drive is achieved through a skew gear meshing
with a ring gear mounted on the reverse annular
gear carrier which is attached to the output pinion
The two types of governor valves used for this
class of automatic transmission are the ball and
pivot flyweight and the plunger and flyweight.
These governors are described below.
Second gear band servo (2GBS) (Fig. 5.24) This is
a double acting piston servo which has a small
piston area to apply the band brake and a large
piston area which is on the release spring chamber
side of the servo.
Directing line pressure to the small piston area
chamber of the servo applies the band brake against
the resistance of the return spring and thereby holds
stationary both sun gears. Introducing line pressure
on the large piston area spring chamber side of the
servo produces an opposing force which releases the
grip on the band brake. The piston returns to the
`off' position and the relaxing of the band brake is
made possible by the difference in piston area on
each side, both sides being subjected to the same line
pressure. The band brake is applied only in the
second gear forward speed range.
During upshift from 2±3 it is important that the
second gear band brake does not release too
quickly relative to the drive and reverse clutch
engagement, in order to avoid run-up (rapid engine
speed surge) during the transition from 2nd to 3rd
gear. During downshift it is also important that the
second gear band brake does not engage before the
drive and rear clutch releases in order to avoid
tie-up (gear jamming) on the 3±2 shift.
The 3±2 control valve and the 3±2 kickdown
valves therefore affect the timing relationship
between the second gear band servo and the drive
and reverse clutch to provide correct shift changes
under all operating conditions.
Plunger and flyweight type governor (Fig. 5.23)
Rotation of the governor at low speed causes the
governor weight and valve to produce a centrifugal
force. This outward force is opposed by an equal
and opposite hydraulic force produced by governor
pressure acting on the stepped annular area of the
governor valve. Because the governor valve is a
regulating valve, and will attempt to remain in
equilibrium, governor pressure will rise in accordance with the increase in centrifugal force caused
by increased rotational speed. As the output shaft
speed increases, the governor weight moves outwards (due to the centrifugal force) to a stop in
the governor body, when it can move no further.
When this occurs, the governor spring located
between the weight and the governor valve
becomes effective. The force of this spring then
combines with centrifugal force of the governor
valve to oppose the hydraulic pressure, thus
making the pressure less sensitive to output shaft
speed variation. Therefore the governor provides
two distinct phases of regulation, the first being
used for accurate control of the low speed shift
First gear one way clutch (OWC) (Fig. 5.13(a))
When in drive range, the one way roller type clutch
Fig. 5.23
Plunger and governor valve
Ball and pivot flyweight type governor (Fig. 5.24)
This type of governor consists of a ball valve
controlled by a hinged flyweight and a pressure
relief ball valve. Fluid from the oil pump at line
pressure is introduced via a restriction into an axial
passage formed in the governor drive shaft. When
the transmission output shaft stops rotating (vehicle stationary) with the engine idling, fluid pressure
forces the governor ball valve off its seat, permitting fluid to escape back to the sump. Rotation of
the output shaft as the vehicle accelerates from a
standstill causes the flyweight centrifugal force to
close the ball valve. Therefore fluid trapped in the
governor drive shaft passage, known as governor
pressure, has to reach a higher pressure before fluid
exhausts through the valve. By these means the line
pressure is regulated to a valve that varies with the
output shaft and vehicle speed. A pressure relief
valve is also included to safeguard the system
from excessively high pressure if the governor
valve malfunctions.
5.7.19 Hydraulic accumulator (Fig. 5.24)
This is a cylinder and spring loaded piston which is
used to store a small amount of pressure energy to
enable a rapid flow of fluid under pressure to one
of the operating components or to absorb and
smooth fluctuating fluid delivery. The piston is
pushed back when the fluid pressure exceeds the
spring load and fluid enters and fills up the space
left behind by the displaced piston.
With the transmission in neutral or park, line
pressure from the pressure pump enters the accumulator at the opposite end to the spring, thereby
displacing the piston and compressing the spring.
When the hydraulic control shifts into the second
gear phase, line pressure from the 1±2 shift valve is
directed to the second gear band servo applied end
and the spring end of the accumulator.
When the accumulator spring is compressed,
fluid from the supply can flow rapidly to the
applied side of the band servo piston. As soon as
the servo piston meets resistance (starts to apply its
load), the fluid pressure increases and the accumulator piston spring is extended as the piston is
pushed back by the spring. This is because there is
equal line pressure acting on either side of the
accumulator piston and so the spring is able to
apply its load and extend. As a result, the supply
of fluid is reduced to the applied side of the second
gear band servo piston. The accumulator therefore
smooths and times the application of the second
gear band brake in order to reduce the risk of shock
and a jerky operation. In addition, the extra quantity of fluid in the system due to the accumulator
leads to a slow rate of release of the servo piston
and band.
5.8.2 Fluid flow in park (Fig. 5.24)
With the manual valve in park position, the
hydraulic flow is the same as in neutral, except
there is no line pressure to the 2±3 shift valve and
the main pressure regulating valve provides maximum increase in line pressure.
5.8.3 Fluid flow in drive range Ð first gear
(Figs 5.24 and 5.19)
Both main pressure regulating valve and throttle
pressure valve operate as for neutral and park.
With the manual selector valve in D, line pressure is directed to the 1±2 shift valve, 2±3 shift valve
and to the forward clutch which it engages.
The shift valve is subjected to governor pressure
at one end which opposes the spring tension and
throttle pressure at the opposite end. As the car
speed increases governor pressure will overcome
throttle pressure causing a 1±2 upshift to take place.
Throughout this period a reduced line pressure
reacts against the left hand end of the main pressure
regulating valve. The valve movement then allows
more oil to pass to the torque converter, thereby
causing a reduction in line pressure to occur.
5.8 Hydraulic gear selection control operation
5.8.1 Fluid flow in neutral (Fig. 5.24)
Pressurized oil from the pump flows to the main
pressure regulating valve. The valve shifts over due
to the oil pressure, thus opening a passage supplying the torque converter via the converter pressure
valve. Increased pump pressure moves the valve
further until it uncovers the exhaust port dumping
the oil back into the oil pump suction intake
The oil pressure generated between the pump
and main pressure limiting valve is known as line
pressure and is directed to the throttle pressure
limiting valve, accumulator, manual selection
valve and through the latter valve to the 2±3 shift
When the throttle foot released, the intermediate
pressure exhausts so that there will be no throttle
pressure. Depressing the throttle pedal increases
the throttle spring tension and creates a throttle
pressure which is then directed to the kickdown
valve, main pressure limiting valve, 1±2 shift valve
and the 2±3 shift valve.
At the same time, a limited throttle pressure is
created between the main pressure limiting valve
and main pressure regulating valve. This pressure is
also directed to the soft engagement valve.
With the manual gear selector valve in neutral,
there is no line pressure to the rear of the main
regulating valve and therefore the trapped line
pressure will be at a maximum.
5.8.4 Fluid flow in drive range Ð second gear
(Figs 5.24 and 5.19)
As for drive range Ð first gear, the main regulator
valve and throttle pressure valve function as for
neutral and park.
When the manual selector valve is positioned in
D, line pressure is directed to the 1±2 shift valve,
2±3 shift valve and to the forward clutch which
it applies.
The rising governor pressure imposes itself
against one end of the 1±2 governor plug counteracting throttle pressure until at some predetermined pressure difference the valve shifts across.
This then permits line pressure to flow to the accumulator and the second gear servo piston thus
causing the second gear brake band to be applied.
At the same time, full line pressure is applied to
the left hand end of the main pressure regulating
valve so that there will be a further decrease in line
pressure relative to drive range ± first gear operating conditions.
5.8.5 Fluid flow in drive range Ð third gear
(Figs 5.24 and 5.20(a))
As for drive range ± first and second gears, the
main regulator valve and throttle pressure valve
perform as for neutral and park.
The manual selector valve will still be in D position so that line pressure is directed to both 1±2 and
Fig. 5.24
Three speed automatic transmission hydraulic control system in neutral position
2±3 shift valves and to the forward clutch piston
which clamps the friction clutch plates.
Increased governor pressure acting on the 2±3
governor plug moves the adjacent 2±3 shift valve
over. This allows line pressure to flow around the
wasted region of the shift valve to the 3±2 kickdown valve and from there to the second gear band
servo spring chamber side (release) of the piston.
Simultaneously, line pressure passes from the 2±3
shift valve through the right hand ball valve orifice
for direct and reverse clutch to the direct and
reverse clutch piston which is accordingly engaged.
The main pressure regulating valve will be subjected to full line pressure acting on its left hand
end so that line pressure reduction will be as in
drive range Ð second gear.
reduced line pressure to be created. Likewise in
second gear full line pressure will act behind the
main pressure regulating valve so that the line
pressure is further reduced.
5.8.8 Fluid flow in drive range Ð second gear Ð
forced kickdown (Figs 5.24 and 5.15(b))
Similarly as for all other manual selector positions,
the main pressure regulating valve and throttle
pressure valve operate in the same way as for neutral and park.
With the manual selector valve in D and the
accelerator pedal fully depressed, the kickdown
valve reduced waste aligns with the kickdown line
outlet port thus causing throttle pressure to flow
into the kickdown lines.
Kickdown pressure now flows to 1±2 and 2±3
governor plugs assisting the throttle pressure
applied on the 1±2 and 2±3 shift valves. When the
governor pressure is low enough, throttle pressure
and kickdown pressure overcomes governor pressure, causing the 2±3 governor plug to move to the
right. As a result, the 2±3 shift valve moves to
exhaust oil from the drive and reverse clutch piston
chamber and the second gear band servo piston
spring chamber causing a 3±2 downshift to occur.
5.8.6 Fluid flow in first gear Ð manual selection
(Figs 5.24 and 5.15)
With the manual selector in `1' position, line pressure
passes to the forward clutch piston and accordingly
applies the clutch plates. Line pressure from the
manual selector valve moves the ball valve for the
first gear manual range so that it cuts off throttle
pressure to the 1±2 shift valve. Line pressure is therefore able to pass through the ball valve for first and
reverse gear to the 1±2 governor plug, the soft
engagement valve and finally passing to the first
and reverse brake piston to engage the brake plates.
Consequently, line pressure will fill the normal
throttle pressure lines of 1±2 shift valve and will
react against the left hand end of the valve. This
then prevents governor pressure at the opposite
end acting on the governor plug moving the valve
for a 1±2 upshift.
5.8.9 Fluid flow in reverse gear
(Figs 5.24, 5.21(a) and 5.14(a))
In reverse gear the pressure regulator valve and
throttle pressure valve operate in a similar way to
With the manual selection valve in the R position, line pressure is directed to the drive and
reverse clutch piston by way of the ball valves for
direct and reverse clutch. Similarly, line pressure is
directed through the ball valves for first and reverse
gear brake to the soft engagement valve and from
there to the first and reverse gear brake piston
thereby engaging the clutch plates.
Whilst in reverse gear, the main pressure regulating valve provides maximum line pressure increase,
this being due to there being no pressure acting on
the left hand end of the valve.
5.8.7 Fluid flow in second gear Ð manual
selection (Figs 5.24 and 5.19)
During this phase of gear change, the main pressure regulating valve and throttle pressure valve
operate as for neutral and park.
When the manual selector valve is in `2' position,
line pressure passes to the forward clutch piston to
apply the clutch plates. Similarly, line pressure is
also directed to the middle of the 1±2 shift valve.
When road speed is high enough, governor pressure
will be sufficient to push the valve to one side, thus
uncovering the port feeding the accumulator and the
second gear band servo piston on the applied side.
In the second gear Ð manual selection position,
line pressure passing to the 2±3 shift valve is
blocked so that a 2±3 upshift is prevented.
In first gear a reduced pressure is applied at the
end of the main pressure regulating valve, causing a
5.8.10 Transmission power train operating faults
The effective operation of an automatic transmission depends greatly upon clutch, band and one
way clutch holding ability, torque converter one way
clutch operation and engine performance. The method
used in diagnosing faults in the engagement components of the transmission is known as the stall
test. This test entails accelerating the engine with
the throttle wide open to maximum speed while the
torque converter turbine is held stationary.
a) Slip in R can be the drive and reverse clutch or
first and reverse brake. Engage `1' range. If slip
still occurs, first and reverse brake must be slipping.
b) Slip in D can be forward clutch or one way
clutch. Engage `1'. If slip still occurs, forward
clutch must be slipping.
c) Slip in R can be the drive and reverse clutch or
first and reverse brake. Engage `1' range. If there
is no slip the drive and reverse clutch could be
Stall test procedure
1 Drive car or run engine until engine and transmission has attained normal working temperatures.
2 Check the level of fluid in the transmission box
and correct if necessary.
3 Apply the hand brake and chock the wheels.
4 Connect a tachometer via leads to the coil ignition terminals.
5 Apply the foot brake, select D range and fully
press the accelerator pedal down for a period not
exceeding 10 seconds to avoid overheating the
transmission fluid (this is very important).
6 Quickly observe the highest engine speed
reached on the tachometer and immediately
release the throttle pedal.
7 Shift selector lever to N and allow transmission
fluid to cool at least two minutes or more before
commencing next test.
8 Repeat tests 5 and 6 in `1' and R range.
Road test for defective torque converter A road test
enables a seized or slipping stator to be engaged,
whereas the stall test can only indicate a possible
slipping stator. The symptoms for a faulty stator
one way clutch are shown in Table 5.5.
If the converter one way clutch has seized, the
vehicle will have poor high speed performance
because the stator reaction above the coupling
point speed hinders the circulation of fluid if it is
not able to freewheel. Conversely, if the converter
one way clutch is slipping there will be no stator
reaction for the fluid and therefore no torque multiplication so that the acceleration will be sluggish
up to about 50 km/h.
Interpreting stall test results A typical stall test
maximum engine speed could be 2300 rev/min
100. If the actual stall speed differs from the
recommended value (i.e. 2300 rev/min), Table 5.3
should be used as a guide to trace the fault. The
stall test therefore helps to determine if the fault is
due to the engine, the converter or the transmission
Note The reason why a slipping torque converter stator drags down the engine's maximum
speed is because the spinning stator makes the
converter behave as a fluid coupling (no torque
multiplication), causing the fluid to have a retarding effect on the impeller.
By performing the stall test in D, `1' and R range,
observing in which range or ranges the slippage
occurs and comparing which clutch or band operates in the slipping range enables the effective
components to be eliminated and the defective
components to be identified (see Table 5.4).
5.9 The continuously variable belt and pulley
The continuously variable transmission CVT, as
used by Ford and Fiat, is based simply on the
Table 5.4
Selector position
Possible causes
Below 1600 rev/min
Approximately 2100 rev/min
Above 2500 rev/min
Stator slip
Poor engine performance
Transmission slip
Possible fault
(D ‡ R)C or (1 ‡ R)B
FC, OWC or (1 ‡ R)C
Table 5.5 Table of symptoms for a faulty one way clutch
Vehicle response
Table 5.3 Table of stall tests
Test results
Table of possible faults
0±50 km/h
above 50 km/h
Slipping stator
Very sluggish
No hill start
Drives normally
Drives normally
Seized stator
Loss of power
principle of a belt running between two V-shaped
pulleys which is designed so that the effective belt
contact diameter settings can be altered to produce
a stepless change in the input to output pulley shaft
Van Doorne Transmissie in Holland has been
mainly responsible for the development of the
steel belt which is the key component in the transmission. At present the steel belt power output
capacity is suitable for engine sizes up to 1.6 litres
but there does not appear to be any reason why
uprated steel belts cannot be developed.
This type of transmission does not suffer from
the limitations of the inefficient torque converter
which is almost universely used by automatic transmissions incorporating epicyclic gear trains operated
by multiplate clutches and band brakes.
drive) is achieved if the belt contact with the
primary pulley is at a smaller diameter relative to
the secondary pulley wrap diameter (Fig. 5.26).
In the case of the Ford Fiesta, the pulleys
provide a continuously variable range of ratios
from bottom 2.6:1 to a super overdrive top of
An intermediate gear reduction of about 1.4:1
between the belt output pulley shaft and the final
drive crownwheel is also provided so that the transmission can be made to match the engine's power
output and the car's design expectation.
5.9.2 Belt design (Figs 5.26 and 5.27)
Power is transmitted from the input to the output
pulley through a steel belt which resembles a steel
necklace of thin trapezoidal plates strung together
between two multistrip bands made from flexible
high strength steel (Fig. 5.27). There are 300 plates,
each plate being roughly 2 mm thick, 25 mm wide
and 12 mm deep, so that the total length of the
endless belt is approximately 600 mm. Each band
is composed of 10 continuous strips 0.18 mm thick.
Also made of high strength steel, they fit into location slots on either side of the plates, their purpose
being to guide the plates, whereas it is the plates'
function to transmit the drive by pushing. Another
feature of the plates is that they are embossed in
a dimple form to assist in the automatic alignment
of the plates as they flex around the pulley.
Contact between the belt plates and pulley is
provided by the tapered edges of the plates which
match the inclination of the pulley walls. When in
drive, both primary and secondary sliding half
pulleys are forced against the belt so that different
plates are in contact and are wedged between the
vee profile of the pulley at any one time.
Consequently, the grip produced between the
plates and pulley walls also forces the plates
together so that in effect they become a continuous
strut which transmits drive in compression (unlike
the conventional belt which transfers power under
tension (Fig. 5.26)). The function of the non-drive
side of the belt, usually referred to as the slack side,
is only to return the plate elements back to the
beginning of the drive (compressive) side of the
The relative movement between the band strips
and the plates by this design is very small, therefore
frictional losses are low. Nevertheless the transmission efficiency is only 92% with a one to one speed
ratio dropping to something like 86% at pull-away,
when the speed ratio reduction is 2.6:1.
5.9.1 Stepless speed ratios (Figs 5.25 and 5.26)
The transmission consists basically of a pair of
variable width vee-shaped pulleys which are interconnected by a composite steel belt. Each pulley
consists of two shallow half cones facing each other
and mounted on a shaft, one being rigidly attached
to it whereas the other half is free to slide axially on
linear ball splines (Fig. 5.25). The variable speed
ratios are obtained by increasing or decreasing the
effective wrap contact diameter of the belt with the
primary input pulley producing a corresponding
reduction or enlargement of the secondary output
pulley working diameter. The belt variable wrap
contact diameter for both primary and secondary
pulleys is obtained by the wedge shaped belt being
supported between the inclined adjacent walls of
the two half pulleys.
When the primary input half pulleys are brought
axially closer, the wedge or vee-shaped belt running
between them is squeezed and is forced to ride up
the tapered walls to a larger diameter. Conversely,
since the belt is endless and inextensible, the secondary output half pulleys are compelled to separate, thus permitting the belt wrap to move inwards
to a smaller diameter.
Alternatively, drawing the secondary output half
pulleys closer to each other enlarges the belt's running diameter at that end. Accordingly it must
reduce the primary input pulley wrap diameter at
the opposite end. A one to one speed ratio is
obtained when both primary and secondary pulleys
are working at the same belt diameter (Fig. 5.26).
A speed ratio reduction (underdrive) occurs when
the primary input pulley operates at a larger belt
contact diameter than the secondary output pulley
(Fig. 5.26). Conversely, a speed ratio increase (over
Fig. 5.25
Section view of a transverse continuously variable transmission
Fig. 5.26 Illustration of pulley and belt under- and overdrive speed ratios
Fig. 5.27 Steel belt construction
5.9.3 Hydraulic control system (Fig. 5.28)
The speed ratio setting control is achieved by a
spur type hydraulic pump and control unit which
supplies oil pressure to both primary and secondary sliding pulley servo cylinders (Fig. 5.28). The
ratio settings are controlled by the pressure
exerted by the larger primary servo cylinder
which accordingly moves the sliding half pulley
axially inwards or outwards to reduce or increase
the output speed setting respectively. This primary
cylinder pressure causes the secondary sliding
pulley and smaller secondary servo cylinder to
move proportionally in the opposite direction
against the resistance of both the return spring
and the secondary cylinder pressure, this being
necessary to provide the correct clamping loads
between the belt and pulleys' walls. The cylinder
pressure necessary to prevent slippage of the belt
varies from around 22 bar for the pull away lowest
ratio setting to approximately 8 bar for the highest
overdrive setting.
The speed ratio setting and belt clamping load
control is achieved via a primary pulley position
senser road assembly.
Fig. 5.28
Transaxle continuously variable belt and pulley transmission layout
However, engine and road speed signals are provided by a pair of pitot tubes which sense the rate of
fluid movement, this being a measure of speed, be it
either under the influence of fluid flow caused by
the engine's input or by the output drive relating to
vehicle speed.
5.9.4 Epicyclic gear train construction and
description (Figs 5.25 and 5.28)
Drive in both forward and reverse direction is
obtained by a single epicyclic gear train controlled
by a forward multiplate clutch and a reverse multiplate brake, both of which are of the wet type
(immersed in oil) (Fig. 5.25). The forward clutch is
not only used for engagement of the drive but also
to provide an initial power take-up when driving
away from rest.
The epicyclic gear train consists of an input planetary carrier, which supports three sets of double
planetary gears, and the input forward clutch
plates. Surrounding the planetary gears is an
internally toothed annulus gear which also supports the rotating reverse brake plates. In the centre
of the planetary gears is a sun gear which is
attached to the primary pulley drive shaft.
opposite sense anticlockwise, that is in the reverse
direction to the input drive from the engine.
5.9.5 Performance characteristics (Fig. 5.29)
With D drive selected and the car at a standstill
with the engine idling, the forward clutch is just
sufficiently engaged to produce a small amount of
transmission drag (point 1). This tends to make the
car creep forwards which can be beneficial when on
a slight incline (Fig. 5.29). Opening the throttle
slightly fully engages the clutch, causing the car to
move positively forwards (point 2). Depressing the
accelerator pedal further sets the speed ratio according to the engine speed, road speed and the driver's
requirements. The wider the throttle is opened the
lower the speed ratio setting will be and the higher
the engine speed and vice versa. With a light constant throttle opening at a minimum of about
1700 rev/min (point 3) the speed ratio moves up to
the greatest possible ratio for a road speed of
roughly 65 km/h which can be achieved on a level
road. If the throttle is opened still wider (point 4) the
speed ratio setting will again change up, but at a
higher engine speed. Fully depressing the accelerator
pedal will cause the engine speed to rise fairly
rapidly (point 5) to about 4500 rev/min and will
remain at this engine speed until a much higher
road speed is attained. If the engine speed still continues to rise the pulley system will continue to
change up until maximum road speed (point 6) has
been reached somewhere near 5000 rev/min.
Partially reducing the throttle open then causes
the pulley combination to move up well into the
overdrive speed ratio setting, so that the engine
speed decreases with only a small reduction in the
car's cruising speed (point 7).
Even more throttle reduction at this road speed
causes the pulley combination speed ratio setting to
go into what is known as a backout upshift (point
8), where the overdrive speed ratio reaches its maximum limit. Opening the throttle wide again brings
about a kickdown downshift (point 9) so that there
is a surplus of power for acceleration. A further
feature which provides engine braking when
driving fast on winding and hilly slopes is through
the selection of L range; this changes the form of
driving by preventing an upshift when the throttle
is eased and in fact causes the pulley combination
to move the speed ratio towards an underdrive
situation (point 10), where the engine operates
between 3000 and 4000 rev/min over an extensive
road speed range.
The output torque developed by this continuously variable transmission approaches the ideal
Neutral or park (N or P position) (Fig. 5.28) When
neutral or park position is selected, both the multiple clutch and brake are disengaged. This means
that the annulus gear and the planetary gears driven
by the input planetary carrier are free to revolve
around the sun gear without transmitting any
power to the primary pulley shaft.
The only additional feature when park position
is selected is that a locking pawl is made to engage
a ring gear on the secondary pulley shaft, thereby
preventing it from rotating and causing the car to
creep forward.
Forward drive (D or L position) (Fig. 5.28) Selecting D or L drive energizes the forward clutch so
that torque is transmitted from the input engine
drive to the right and left hand planetary carriers
and planet pins, through the forward clutch
clamped drive and driven multiplates. Finally it is
transferred by the clutch outer casing to the primary pulley shaft. The forward gear drive is a direct
drive causing the planetary gear set to revolve
bodily at engine speed with no relative rotational
movement of the gears themselves.
Reverse drive (R position) (Fig. 5.28) Selecting
reverse gear disengages the forward clutch and
energizes the reverse multiplate brake. As a result,
the annular gear is held stationary and the input
from the engine rotates the planetary carrier (Fig.
The forward clockwise rotation of the carrier
causes the outer planet gears to rotate on their
own axes as they are compelled to roll round the
inside of internally toothed annular gear in an
anticlockwise direction.
Motion is then transferred from the outer planet
gears to the sun gear via the inner planet gears.
Because they are forced to rotate clockwise, the
meshing sun gear is directionally moved in the
Fig. 5.29
CVT speed and torque performance characteristics
constant power curve (Fig. 5.29) in which the
torque produced is inversely proportional to the
car's road speed.
5.10.2 Gear train power flow for individual gear
D drive range Ð first gear (Fig. 5.31) With the
position selector lever in D drive range, the one way
clutch (OWC) holds the front planet carrier while
multiplate clutch (B) and the multiplate brake (G)
are applied. Power flows from the engine to the
torque converter pump wheel, via the fluid media
to the output turbine wheel. It is then directed by
way of the input shaft and the applied multiplate
clutch (B) to the front planetary large sun gear (SL).
With the front planet carrier (CF ) held stationary
by the locked one way clutch (OWC), power passes
from the large sun gear (SL) to the long planet gears
(PL) in an anticlockwise direction. The long planet
gear (PL) therefore drives the short planet gears
(PS) in a clockwise direction thus compelling the
front annular ring gear (AF ) to move in a clockwise
direction. Power thus flows from the front annular
ring gear (AF ) though the rear intermediate shaft
to the rear planetary gear annular ring gear (AR) in
a clockwise direction. With the rear sun gear (SR)
held stationary by the applied multiplate brake
(G), the rear planet gears (PR) are forced to roll
around the fixed sun gear in a clockwise direction,
this in turn compels the rear planet carrier (CR) and
the output shaft also to rotate in a clockwise direction at a much reduced speed. Thus a two stage
speed reduction produces an overall underdrive
5.10 Five speed automatic transmission with
electronic-hydraulic control
5.10.1 Automatic transmission gear train system
(Fig. 5.30)
This five speed automatic transmission system is
broadly based on a ZF design. Power is supplied
though a hydrodynamic three element torque converter incorporating an integral disc type lock-up
clutch. The power drive is then directed though
a Ravigneaux type dual planetary gear train which
provides five forward gears and one reverse gear; it
then passes to the output side via a second stage
single planetary gear train. The Ravigneaux planetary gear train has both large and small input sun
gears, the large sun gears mesh with three long
planet gears whereas the small sun gears mesh
with three short planet gears; both the long and
the short planet gears are supported on a single
gear carrier. A single ring-gear meshing with the
short planet gear forms the output side of the planetary gear train. Individual gear ratios are selected
by applying the input torque to either the pinion
carrier or one of the sun gears and holding various
other members stationary.
Front planetary
gear train
(2+3+5)B E
Input shaft
Rear planetary
gear train
Fig. 5.30
Five speed and reverse automatic transmission (transaxle/longitudinal) layout
Input shaft
Fig. 5.31
Front intermediate
Five speed and reverse automatic transmission power flow first gear
first gear. If the `2' first gear is selected multiplate
brake (F) is applied in addition to the multiplate
clutch (B) and multiplate brake (G). As a result
instead of the one way clutch (OWC) allowing
the vehicle to freewheel on overrun, the multiplate brake (F) locks the front planetary carrier
(CF ) to the casing. Consequently a positive drive
exists between the engine and transmission on
both drive and overrun: it thus enables engine
braking to be applied to the transmission when
the transmission is overrunning the engine.
the stationary small sun gear (SS). Consequently,
the annular ring gear (AF ) is also forced to rotate in
a clockwise direction but at a reduced speed to that
of the input large sun gear (SL). The drive is then
transferred from the front planetary annular ring
gear (AF ) to the rear planetary annular ring gear
(AR) via the rear intermediate shaft. With the multiplate clutch (D) applied the rear planetary sun
gear (SR) and rear annular ring gear (AR) are
locked together, thus preventing the rear planet
gears from rotating independently on their axes.
The drive therefore passes directly from the rear
annular ring gear (AR) to the rear carrier (CR) and
output shaft via the jammed rear planet gears.
Thus it can be seen that the overall gear reduction
is obtained in the front planetary gear train,
whereas the rear planetary gear train only provides
a one-to-one through drive.
D drive range Ð second gear (Fig. 5.32) With the
position selector lever in D drive range, multiplate
clutch (C) and multiplate brakes (B) and (G) are
Power flows from the engine via the torque converter to the input shaft, it then passes via the
multiplate clutch (B) to the first planetary large
sun gear (SL). With the multiplate brake (E)
applied, the front planetary small sun gear (SS) is
held stationary. Consequently the large sun gear
(SL) drives the long plant gears (PL) anticlockwise
and the short planet gears (PS) clockwise, and at
the same time, the short planet gears (PS) are compelled to roll in a clockwise direction around the
stationary small sun gear (SS).
The drive then passes from the front planetary
annular ring gear (AF ) to the rear planetary annular ring gear (AR) via the rear intermediate shaft.
With the rear sun gear (SR) held stationary by the
applied multiplate brake (G) the clockwise rotation of the rear annular ring gear (AR) compels
the rear planet gears (PR) to roll around the held
rear sun gear (SR) in a clockwise direction taking
with it the rear carrier (CR) and the output shaft
at a reduced speed. Thus the overall gear reduction takes place in both front and rear planetary
gear trains.
D drive range Ð fourth gear (Fig. 5.34) With the
positive selector lever in D drive range, multiplate
clutches (B), (C) and (D) are applied. Power flows
from the engine via the torque converter to the
input shaft, it then passes via the multiplate clutch
(B) to the front planetary large sun gear (SL) and
via the multiplate clutch (C) to the front planetary
planet-gear carrier (CF ). Consequently both the
large sun gear and the planet carrier rotate at the
same speed thereby preventing any relative planetary gear motion, that is, the gears are jammed.
Hence the output drive speed via the annular ring
gear (AF ) and the rear intermediate shaft is the
same as that of the input shaft speed. Power is then
transferred to the rear planetary gear train by way
of the front annular ring gear (AF ) and rear intermediate shaft to the rear planetary annular ring
gear (AR) and rear intermediate shaft to the rear
planetary annular ring gear (AR). However, with
the multiplate clutch (D) applied, the rear annular
ring gear (AR) becomes locked to the rear sun gear
(SR); the drive therefore flows directly from the rear
annular ring gear to the rear planet carrier (CR) and
output shaft via the jammed planet gears. Thus
there is no gear reduction in both front and rear
planetary gear trains, hence the input and output
rotary speeds are similar.
D drive range Ð third gear (Fig. 5.33) With the
position selector lever in D drive range, multiplate
clutches (B) and (D), and multiplate brake (E) are
Power flows from the engine via the torque converter to the input shaft, it then passes via the
multiplate clutch (B) to the front planetary large
sun gear (SL). With the multiplate brake (E)
applied, the front planetary small sun gear (SS) is
held stationary. This results in the large sun gear
(SL) driving the long planet gears (PL) anticlockwise and the short planet gears (PS) clockwise, and
simultaneously, the short planet gears (PS) are
compelled to roll in a clockwise direction around
D drive range Ð fifth gear (Fig. 5.35) With the
position selector lever in D drive range, multiplate
clutches (C) and (D) and multiplate brake (E) are
applied. Power flows from the engine via the torque
converter to the input shaft, it then passes via the
multiplate clutch (C) to the front planetary planet
Fig. 5.32
Second gear
Input shaft
Front intermediate
Third gear
Front intermediate
Fig. 5.33
Input shaft
Input shaft
Fig. 5.34
Fourth gear
Front intermediate
Input shaft
Fifth gear
Front intermediate
Fig. 5.35
Rear intermediate
Rear planetary
gear train
carrier (CF ). With the multiplate brake (E) applied
the front planetary short sun gear (SS) remains
stationary. As a result the planet gear carrier (CF )
and both long and short planet gear pins are driven
around in a clockwise direction; it thus compels the
short planet gears (PS) to roll clockwise around the
fixed small sun gear (SS). It also causes the annulus
ring gear (AF ) to revolve around its axis; however,
this will be at a speed greater than the input planet
carrier (CF ). Note that the long planet gears (PL)
and large sun gear (SL) revolve but are both
inactive. The drive then passes from the front
planetary annular ring gear (AF ) to the rear planetary annular ring gear (AR) via the rear intermediate
shaft. With the multiplate clutch (D) applied both
rear annular ring gear (AR) and rear sun gear (SR)
are locked together. Hence the rear planet gears
sandwiched between both the sun and the annular
gears also jam; the drive therefore is passed directly
though the jammed rear planetary gear train cluster
to the output shaft without a change in speed. An
overall speed step-up is thus obtained, that is, an
overdrive fifth gear is achieved, the step-up taking
place only in the first stage planetary gear train, the
second planetary gear train providing only a
through one-to-one drive.
5.10.3 Gear change±hydraulic control (Fig. 5.30)
The shifting from one gear ratio to another is
achieved by a sprag type one way clutch (when
shifting from first to second gear and vice versa),
four rotating multiplate clutches `A, B, C and D'
and three held multiplate brakes `E, F and G'. The
multiplate clutches and brakes are engaged by electro-hydraulic control, hydraulic pressure being
supplied by the engine driven fluid pump. To
apply a clutch or brake, pressurized fluid from the
hydraulic control unit is directed to an annular
shaped piston chamber causing the piston to
clamp together the drive and driven friction disc
members of the multiplate clutch. Power therefore
is able to be transferred from the input to the output clutch members while these members rotate at
different speeds. Shifting from one ratio to another
takes place by applying and releasing various
multiplate clutches/brakes. During an up or down
gear shift such as 2±3, 3±4, 4±5 or 5±4, 4±3, 3±2 one
clutch engages while another clutch disengages. To
achieve an uninterrupted power flow, the disengaging clutch remains partially engaged but at a
much reduced clamping pressure, whereas the
engaging clutch clamping pressure rise takes place
in a phased pattern.
R reverse gear (Fig. 5.36) With the position selector in reverse R position, the multiplate brakes (F)
and (G) and the multiplate clutch (A) are applied.
Power flows from the engine to the torque converter to the input shaft, it then passes via the multiplate clutch (A) to the front planetary small sun
gear (SS). With the multiplate clutch (F) applied,
the front planet gear carrier (CF ) is held stationary,
and the drive passes from the clockwise rotating
small sun gear (SS) to the short planet gears (PS)
making the latter rotate anticlockwise. As a result
the internal toothed front annular ring gear (AF )
will also be compelled to rotate anticlockwise. The
drive then passes from the front planetary annular
ring gear (AF ) to the rear planetary annular ring
gear (AR) via the rear intermediate shaft. With the
rear sun gear (SR) held stationary by the applied
multiplate brake (G) the anticlockwise rotation of
the rear annular ring gear (AR) compels the rear
planet gear (PR) to roll around the held rear sun
gear (SR) in an anticlockwise direction taking with
it the rear carrier (CR) and the output shaft at a
reduced speed.
Thus the direction of drive is reversed in the first
planetary gear train, and there is an under-drive
gear reduction in both planetary gear trains.
5.10.4 Upshift clutch overlap control
characteristics (Fig. 5.37 (a±c))
The characteristics of a gear ratio upshift is shown
in Fig. 5.37(a), it can be seen with the vehicle accelerating, and without a gear change the engine speed
steadily rises; however, during a gear ratio upshift
transition phase, there is a small rise in engine
speed above that of the speed curve when there is
no gear ratio change taking place. This slight speed
upsurge is caused by a small amount of slip overlap
between applying and releasing the clutches. Immediately after the load transference phase there is
a speed decrease and then a steady speed rise, this
being caused by the full transmitted driving load
now pulling down the engine speed, followed by an
engine power recovery which again allows the
engine speed to rise.
When a gear upshift is about to commence the
engaging clutch pressure Fig. 5.37(b) rises sharply
from residual to main system pressure for a short
period of time, it then drops rapidly to just under
half the main system pressure and remains at this
value up to the load transfer phase. Over the load
transfer phase the engaging clutch pressure rises
fairly quickly; however, after this phase the pressure
rise is at a much lower rate. Finally a small pressure
Input shaft
Input from
Fig. 5.36
Front intermediate
Rear intermediate
Reverse gear
Rear planetary
gear train
Engine speed (rev/min)
Engine speed increasing
without gear change
Engine speed variation
with gear change
Time (seconds)
Pressure (bar)
Engaging clutch
pressure rise
Disengaging clutch
pressure decrease
Time (seconds)
Torque (Nm)
Engaging clutch
Disengaging clutch
Regulating phase
Post regulating phase
Load (torque) transfer phase
Fig. 5.37 (a±c)
Upshift clutch overlap control characteristics
jump brings it back to the main system pressure.
Between the rise and fall of the engaging clutch
pressure, the disengaging clutch pressure falls to
something like two thirds of the main systems pressure, it then remains constant for a period of time.
Near the end of the load transfer phase the pressure
collapses to a very low residual pressure where it
remains during the time the clutch is disengaged.
Fig. 5.37(b) therefore shows a pressure overlap
between the disengaging clutch pressure decrease
and the engaging clutch pressure increase over the
load transfer period. The consequence of too much
pressure overlap would be to cause heavy binding
of the clutch and brake multiclutch plate members
and high internal stresses in the transmission power
line, whereas insufficient pressure overlap causes
the engine speed to rise when driving though the
load transfer period. Fig. 5.37(c) shows how the
torque load transmitted by the engaging and disengaging clutches changes during a gear ratio upshift.
It shows a very small torque dip and recovery for
the disengaging clutch after the initial disengaging
clutch pressure drop, then during the load transfer
phase the disengaging clutch output torque
declines steeply while the engaging clutch output
torque increases rapidly. The resultant transmitted
output torque over the load transfer phase also
shows a dip but recovers and rises very slightly
above the previous maximum torque, this being
due to the transmission now being able to deliver
the full engine torque.
Finally the transmitted engine torque drops a small
amount at the point where the engine speed has
declined to its minimum, it then remains constant as
the engine speed again commences to rise.
valves are energized by the electronic transmission
control unit `ETCU' which in turn receives input
signals from various speed, load, temperature and
accelerator pedal sensors; all of these sensors are
simultaneously and continuously monitoring the
changing parameters. In addition a position selector lever or button operated by the driver relays the
different driving programs to the electronic transmission control unit `ETCU'.
Electronic transmission control unit The function
of the electronic transmission control unit `ETCU'
is to collect, analyse and process all the input
signals and to store program data so that the
appropriate hydraulic circuit pressures will bring
about transmission gear changes to match the
engine speed and torque, vehicle's weight and
load, driver's requirements and road conditions.
Control program The stored program provides
data to give favourable shift characteristics for
gears and the torque converter lock-up clutch, it
co-ordinates parameters for pressure calculations,
engine manipulation and synchronizing gear
change phases, it provides regulation parameters
for smooth gear shifts and the converter lock-up
and finally it has built-in parameters for fault
Transmission input signal sensors The various signals which activate the electronic transmission control unit `ETCU' can be divided into three groups:
(1) transmission, (2) engine and (3) vehicle:
1 Transmission
(a) input turbine speed sensor
(b) output drive speed sensor
(c) transmission temperature sensor
(d) position switch signalling-selector lever
position to the electronic transmission control unit
5.10.5 Description of major hydraulic and
electronic components
Hydraulic control unit (Fig. 5.38) The hydraulic
control unit is housed in the oil-pan position underneath the transmission gears. A fluid pump operates
the hydraulic circuitry; it is driven directly from the
engine via the torque converter casing, fluid is
directed by way of a pressure regulation valve
to the interior of the torque converter and to the
various clutches and brakes via passages and
valves. The hydraulic control circuit which operates
the gear shifts are activated by three electromagnetic operated open/close valves (solenoid
valves) and four electro-magnetic progressive
opening and closing regulation valves (electronic
pressure regulation valve `EPRV'). Both types of
2 Engine
(a) engine speed sensor
(b) engine load-injector opening duration
(c) throttle valve opening-potentiometer
(d) engine temperature sensor
3 Vehicle
(a) kickdown switch
(b) position PRND432 indicator
(c) manual gear selection program
(d) brake light switch
Selector lever
control unit
Selector lever position
Output speed sensor
Electronic transmission
control unit
Input speed sensor
Oil temperature
transmission sensor
Diagnosis connection
Throttle valve potentiometer
Electronic injection
control unit
Cruise control
Gear change
position switch
Brake light
and switch
Fig. 5.38
Fuel injection
Basic electronic control system layout
5.10.6 Description and function of the electro/
hydraulic valves
valve SV-3. Solenoid valve MV3 activates the traction/coasting valve (T/C)V.
Solenoid (electro-magnetic) valves (MV1, MV2 and
MV3) (Fig. 5.39) The solenoid valves MV1, MV2
and MV3 are electro-magnetic disc armature operated ball-type valves which are energized by current
supplied by the electronic transmission control unit.
The ball-valve is either in the open or closed position,
when the valve is de-energized the ball-valve closes
the inlet port and vents the outlet port whereas when
energized the ball-valve blocks the vent port and
opens the inlet port. Solenoid valve MV1 when
energized activates shift valve SV-1 and SV-3.
Solenoid valve MV2 activates shift valve SV-2
and shuts down the switch over function of shift
Electronic pressure regulating valves (EPRV-1,
EPRV-2, EPRV-3 and EPRV-4) (Fig. 5.39) The
electronic pressure regulating valves EPRV-(1±4) are
variable pressure electro-magnetic cylindrical
armature operated needle-type valves, the output
pressure delivered being determined by the magnitude of the current supplied at any one time by the
electronic transmission control unit to the electronic pressure regulating valves. With increasing
current the taper-needle valve orifice is enlarged,
this increases the fluid spill and correspondingly
reduces the actuating control pressure delivered
to the various valves responsible for gear shifts.
TV (5-4)
TV (4-5)
Gear change position switch
Electronic transmission control
Fuel injection
quantity (torque)
Fig. 5.39
Throttle valve
position (acceleration)
Hydraulic/electronic transmission control system ± Neutral position
Speed sensors
Hydraulic/electronic automatic transmission control system abbreviation key for Figs 5.39±5.46.
A list of key components and abbreviations used in the description of the electro/hydraulic control system is as follows:
1 fluid pressure pump
2 selector position valve
3 main pressure valve
4 pressure reducing valves
PRV-1 and PRV-2
5 modulation pressure valve
6 shift valve
SV-1, SV-2 and SV-3
7 reverse gear valve
8 clutch valves
CV-A, CV-B, CV-C and CV-D
9 brake valves
BV-E, BV-F and BV-G
10 retaining valves
RV-E and RV-G
11 traction/coasting valve
12 traction valves
TV (4±5) and TV (5±4)
13 converter pressure valve
14 converter pressure control valve
15 converter lock-up clutch valve
16 lubrication pressure valve
17 solenoid (electro-magnetic) valves
MV1, MV2 and MV3
18 electronic pressure regulating valves
EPRV-1, EPRV-2, EPRV-3 and EPRV-4
19 multiplate clutch/brake
20 pressure relief valve
21 non-return valve
Table 5.6 Hydraulic/electronic automatic transmission control system solenoid valve, clutch and brake engagement sequence for different gear ratios for
Figs 5.39±5.46
Solenoid valve ± clutch ± brake engagement sequence
Gear range
Solenoid valve logic
Clutch and brake logic
Solenoid valves
N/P neutral/park
D 1st gear
D 2nd gear
D 3rd gear
D 4th gear
D 5th gear
2 1st gear
R ˆ reverse
Pressure regulating valves
Selector position valve (SPV) (Fig. 5.39) This
valve is indirectly operated by the driver to select
the forward and reverse direction of drive and the
neutral or park positions.
Control pressure (bar)
Main pressure valve (MPV) (Fig. 5.39) The main
pressure valve `MPV' regulates the fluid pressure
supply produced by the internal gear crescent
pump; it is a variable pressure limiting valve
which relates to driving conditions and the driver's
Pressure reduction valve (PRV-1) (Fig. 5.39) The
pressure reduction valve `PRV-1' reduces the main
fluid pressure supply to an approximate constant 5
bar output pressure which is the necessary fluid
pressure supply to operate the solenoid valves
MV1, MV2 and MV3.
Control current (A)
Pressure reduction valve (PRV-2) (Fig. 5.39) The
pressure reduction valve `PRV' reduces the main
fluid pressure supply to an approximate constant 5
bar output pressure which is the necessary fluid
pressure supply to operate the electronic pressure
regulation valves EPRV-1, EPRV-2, EPRV-3 and
Fig. 5.40 Electronic pressure regulating valve currentpressure characteristics
Thus control pressure delivered is inversely proportional to the amount of current supplied, that is, as
the current rises the pressure decreases and vice
versa. The characteristics of control pressure versus
control current is shown in Fig. 5.40.
Modulation pressure value (MOD-PV) (Fig.
5.39) The modulation pressure valve is actuated
by the electronic pressure regulator valve EPRV-1,
it produces an output pressure which rises proportional to engine torque. The modulation pressure is
conveyed to the main pressure valve and to each of
the clutch valves, its purpose being to raise the
system's pressure and to maximize the opening of
the clutch valves with increased engine load so that
a higher supply pressure reaches the appropriate
multiplate clutch or/and brake.
5.10.7 Description and function of pump and
hydraulic valves
Pump (P) (Fig. 5.39) This internal gear crescenttype pump consists of an internal toothed-spur
ring gear which runs outside but in mesh with a
driving external toothed-spur gear, so that its
axis of rotation is eccentric to that of the driving
gear. Due to their eccentricity, there is a space
between the external and internal gears which is
occupied by a fixed spacer block known as the
crescent whose function is to separate the inletoutput port areas. The rotation of the gears creates a low pressure area at the inlet suction end
of the crescent which draws in fluid. As the gear
wheels rotate, oil will be trapped between the
teeth of the inner driver gear and the inside crescent side walls, and between teeth of the outer
gear and the outside crescent side wall. These
teeth will then carry this fluid around to the
other end of the crescent where it will then be
discharged at pressure by both set of teeth into
the output port.
Shift valves (SV-1, SV-2 and SV-3) (Fig. 5.39) The
shift valves are actuated by the various solenoid valves MV1, MV2 and MV3: the function of
a shift valve is to convey system pressure to the
relevant operating circuit controlling the application or release of the various multiplate
Reverse gear valve (RGV) (Fig. 5.39) The reverse
gear valve functions as a shift valve for selecting
reverse gear; it also acts as safety valve for the
forward gears by interrupting system pressure
reaching clutch `A', thus preventing the reverse
gear being accidentally engaged whenever the
vehicle is moving in a forward direction.
Clutch/brakes (CV-A, CV-B, CV-C, CV-D/BV-E,
BV-F and BV-G) (Fig. 5.39) The clutch valves
control the engagement and disengagement of the
multiplate clutches and brakes. These valves are
variable pressure reduction valves which are actuated by the appropriate solenoid valves, electronic
pressure regulator valves, traction valves and shift
valves and are responsible for producing the
desired clutch pressure variations during each
gear shift phase. Clutch valves CV-B, CV-C and
CV-F are influenced by modulation pressure which
resists the partial closure of the clutch valves, hence
it permits relatively high fluid pressure to reach
these multiplate clutches and brake when large
transmission torque is being transmitted.
plate side of the lock-up clutch when the torque
converter pressure control valve is actuated.
Converter pressure control valve (CPCV)
(Fig. 5.39) The converter pressure control valve
`CPCV' is actuated by the electronic pressure regulation valve `EPRV-4', its purpose being to prevent the converter pressure valve `CPV' from
supplying reduced system pressure to the chamber
formed between the drive-plate and lock-up clutch
and to vent this space. As a result the fluid pressure
on the torque converter side of the lock-up clutch is
able to clamp the latter to the drive-plate.
Converter lock-up clutch valve (CLCV)
(Fig. 5.39) The converter lock-up clutch valve
`CLCV' is actuated with the converter pressure
control valve `CPCV' by the electronic pressure
regulation valve `EPRV-4'. The converter lock-up
clutch valve `CLCV' when actuated changes the
direction of input flow at reduced system pressure
from the drive-plate to the turbine wheel side of the
lock-up clutch. Simultaneously the converter pressure valve `CPV' is actuated, this shifts the valve so
that the space between the drive-plate and lock-up
clutch face is vented. As a result the lock-up clutch
is forced hard against the drive-plate thus locking
out the torque converter function and replacing
it with direct mechanical drive via the lock-up
Retaining valves (RV-E and RV-G) (Fig. 5.39) In
addition to the electronic pressure regulator valve
which actuates the clutch valves, the retaining
valves RV-E and RV-G modify the opening and
closing phases of the clutch valves in such a way as
to cause a progressive build-up or a rapid collapse
of operating multiplate clutch/brake fluid pressure
during engagement or disengagement respectively.
Traction/coasting valve (T/C-V) (Fig. 5.39) The
traction coasting valve T/C-V cuts out the regulating action of the traction valve TV (5±4) and shifts
the traction valve TV (4±5) into the shut-off position when required.
Traction valve (TV) (4±5) (Fig. 5.39) The traction
valve TV (4±5) controls the main system fluid pressure to the multiplate-clutch MPC-B via the traction valve TV (5±4) and clutch valve CV-B and
hence blocks the fluid pressure reaching the multiplate clutch CV-B when there is a upshift from
fourth to fifth gear.
Lubrication pressure valve (LPV) (Fig. 5.39) The
lubrication pressure valve `LPV' supplies fluid
lubricant at a suitable reduced system pressure to
the internal rubbing parts of the transmission gear
5.10.8 Operating description of the electro/
hydraulic control unit
To simplify the various solenoid valve, clutch and
brake engagement sequences for each gear ratio
Table 5.6 has been included.
Traction valve (TV) (5±4) (Fig. 5.39) The traction
valve TV (5±4) is another form of clutch valve, its
function being to supply system pressure to the
multiplate clutch MPC-B via clutch valve CV-B
when there is a downshift from fifth to fourth gear.
Neutral and park position (Fig. 5.39) With the
selector lever in neutral or park position, fluid is
delivered from the oil-pump to the selector position
valve `SPV', modulation pressure valve `MOD-V',
pressure reduction valves `PRV-1' and `PRV-2',
shift valve `SV-1', traction/coasting valve `(T/C)V'
and clutch valve `CV-G'. Regulating fluid pressure
is supplied to the torque converter `TC' via the
converter pressure valve `CPV' and to the lubrication system by way of the lubrication pressure valve
`LPV'. At the same time regulated constant fluid
Converter pressure valve (CPV) (Fig. 5.39) The
converter pressure valve `CPV' supplies the torque
converter with a reduced system pressure to match
the driving demands, that is, driving torque under
varying driving conditions, it also serves as a pressure limiting valve to prevent excessive pressure
build-up in the torque converter if the system
pressure should become unduly high. The valve in
addition vents the chamber formed on the drive172
pressure (5 bar) is supplied to the solenoid valves
`MV1, MV2 and MV3' via the pressure reduction
valve `PRV-1', and the electronic pressure regulating valves `EPRV-(1±4)' via the pressure reduction
valve `PRV-2'. In addition controlling modulation
pressure is relayed to the spring chamber of clutch
valves `CV-B, CV-C and CV-D' and brake valve
`CV-F' via the modulation pressure valve `MODPV'. Neutral and parking position has the following multiplate clutch solenoid valves and electronic
pressure regulator valves activated:
valves `BV-G and RV-G' respectively, enabling
fluid pressure to apply the multiplate brake
Second gear (Fig. 5.42) Engagement of second
gear is obtained by applying multiplate clutch
`MPC-B' and the multiplate brakes `MPB-E and
MPB-G'. This is achieved in the following manner
with the selector position valve in the D drive
1 Multiplate clutch and brake `MPC-B and
MPB-G' respectively applied as for first gear.
2 Solenoid valves `MV1 and MV2' are energized,
thus opening both valves. Fluid pressure from
`MV1' is applied to the left-hand side of both
`SV-1 and SV-3'; however, only valve SV-1 shifts
over to the right-hand side. At the same time
fluid pressure from solenoid valve `MV2' shifts
valve `SV-2' over against the return-spring
tension and also pressurizes the spring end of
shift valve `SV-3'. This prevents shift valve
`SV-3' moving over to the right-hand side when
fluid pressure from solenoid valve `MV-1' is
simultaneously applied at the opposite end.
3 The electronic pressure regulating valves
`EPRV-1 and EPRV-3' have their controlling
current reduced, thereby causing an increase in
line pressure to the modulation valve MOD-PV
and to both brake and retaining valves `BV-G
and RV-B' respectively. Consequently line pressure continues to apply the multiplate brake
4 The electronic pressure regulating valve `EPRV-2'
has its controlling current reduced, thus progressively closing the valve, consequently there
will be an increase in fluid pressure acting on the
right-hand side of both brake and retaining
valves `BV-E and RV-E' respectively. As a result
the brake valve `BV-E' opens to permit line pressure to actuate and apply the multiplate brake
1 multiplate brake `MPB-G'.
2 solenoid valves `MV1 and MV3'.
3 electronic pressure regulating valves `EPRV-1
and EPRV-2'.
First gear (Fig. 5.41) Engagement of first gear is
obtained by applying the one way clutch `OWC' and
multiplate clutch and multiplate brake `MPC-B
and MPB-G' respectively. This is achieved in the
following manner:
1 Moving selector position valve `SPV' into D drive
range. Fluid pressure from the selector position
valve `SPV' then passes via the traction valves `TV
(4±5) and TV (5±4)' respectively to clutch valve
`CV-B', it therefore permits fluid pressure to
apply the multiplate clutch `MPC-B'.
2 Energizing solenoid valves `MV1 and MV2' opens
both valves. Solenoid valve `MV1' applies a
reduced constant fluid pressure to the left-hand
side of shift valves `SV-1 and SV-3'. Shift valve
`SV-1' shifts over to the right-hand side against
the tension of the return spring blocking the fluid
pressure passage leading to clutch valve `CV-D',
however shift valve `SV-3' cannot move over
since a similar reduced constant pressure is introduced to the spring end of the valve via solenoid
valve `MV2'. Solenoid valve `MV2' applies
reduced constant pressure to the left-hand side
of shift valve `SV-2' and the right-hand side of
shift valve `SV-3'; this pushes the shift valve
`SV-2' to the right and so prevents shift valve
`SV-3' also being pushed to the right by fluid
pressure from solenoid valve `MV1' as previously mentioned.
3 Electronic pressure regulator valve `EPRV-1'
supplies a variable regulated fluid pressure to
the modulation pressure valve `MOD-PV', this
pressure being continuously adjusted by the electronic transmission control unit `ETCU' to suit
the operating conditions. Electronic pressure
regulating valve `EPRV-3 supplies a variable
controlling fluid pressure to brake and retaining
Third gear (Fig. 5.43) Engagement of third gear is
obtained by applying the multiplate clutches
`MPC-B and MPC-D' and the multiplate brake
The shift from second to third gear is achieved in
the following manner with the selector position
valve in the D drive range:
1 Multiplate clutch `MPC-B' and multiplate brake
`MPB-E' are applied as for second gear.
2 Solenoid valve `MV2' remains energized thus
keeping the valve open as for first and second
TV (5-4)
TV (4-5)
Fig. 5.41
Hydraulic/electronic transmission control system ± first gear
TV (5-4)
TV (4-5)
Fig. 5.42
Hydraulic/electronic transmission control system ± second gear
TV (5-4)
TV (4-5)
Fig. 5.43
Hydraulic/electronic transmission control system ± third gear
TV (5-4)
TV (4-5)
Fig. 5.44
Hydraulic/electronic transmission control system ± fourth gear
TV (5-4)
TV (4-5)
Fig. 5.45
Hydraulic/electronic transmission control system ± fifth gear
TV (5-4)
TV (4-5)
Fig. 5.46
Hydraulic/electronic transmission control system ± reverse gear
3 Solenoid valve `MV3' is in the de-energized state,
it therefore blocks line pressure reaching traction/
coasting valve `(T/C)V' via passage `Y-Y'.
4 Electronic pressure regulating valves `EPRV-1
and EPRV-2' de-energized, this closes the valves
and increases their respective regulating fluid
pressure as for second gear.
5 Electronic pressure regulating valve `EPRV-3'
control current is increased, this causes the
valve to open and the regulating fluid pressure
to collapse. The returning spring now moves the
clutch and retaining valves `CV-G and RV-G'
respectively over to the right-hand side. Brake
valve `BV-G' now blocks the line pressure reaching the multiplate clutch MPB-G and releases
(exhausts) the line pressure imposed on the
annular shaped brake piston; the multiplate
brake `MPB-G' is therefore disengaged.
6 Solenoid valve `MV1' is de-energized, this permits the shift valve `SV-1' to return to the lefthand side. Subsequently line pressure now passes
via the shift valve `SV-1' to the clutch valve `CV-D'
and hence applies the multiplate clutch `MPC-D'.
valve `CPCV' to the left-hand side and converter
lock-up clutch `CLCV' to the right-hand side.
Fluid pressure is thus supplied via the converter
lock-up clutch valve `CLCV' to the torque converter `TC', whereas fluid pressure reaching the
left-hand side of the torque converter lock-up
clutch chamber is now blocked by the converter
lock-up clutch valve `CLCV' and exhausted by
the converter pressure valve `CPV'. As a result
fluid pressure within the torque converter pushes
the lock-up clutch hard against the impeller
rotor casing. Subsequently the transmission
drive, instead of passing via fluid media from
the impeller-rotor casing to the turbine-rotor
output shaft, is now diverted directly via the
lock-up clutch from the impeller-rotor casing to
the turbine-rotor output shaft.
Fifth gear (Fig. 5.45) Engagement of fifth gear is
obtained by applying the multiplate clutches
`MPC-C and MPC-D' and the multiplate brake
The shift from fourth to fifth gear is achieved in
the following manner with the selector position
valve `SPV' in the D drive range:
1 Multiplate clutches `MPC-C and MPC-D'
applied as for fourth gear.
2 Solenoid valve `MV2' de-energized as for fourth
3 Solenoid valve `MV3' is energized, this allows
fluid pressure via passage `Y-Y' to shift traction/coasting valve `(T/C)V' over to the righthand side. As a result fluid pressure is released
(exhausts) from the spring side of the traction
valve `TV (5±4)', hence fluid pressure acting on
the left-hand end of the valve now enables it to
shift to the right-hand side.
4 Solenoid valve `MV1' is energized, this pressurizes the left-hand side of the shift valves `SV-1
and SV-3'. However, `SV-1' cannot move over
due to the existing fluid pressure acting on the
spring end of the valve, whereas `SV-3' is free to
shift to the right-hand end. Fluid pressure from
the clutch valve `CV-E' now passes via shift valve
`SV-3' and traction/coasting valve `(T/C)V' to
the traction valve `TV (4±5)' causing the latter
to shift to the right-hand side. Consequently
traction valve `TV (4±5)' now blocks the main
fluid pressure passing through the clutch valve
`CV-B' and simultaneously releases the multiplate clutch `MPC-B' by exhausting the fluid
pressure being applied to it.
5 Electronic pressure regulating valve `EPRV-2'
de-energized and partially closed. Controlled
Fourth gear (Fig. 5.44) Engagement of fourth
gear is obtained by applying the multiplate clutches
`MPC-B, MPC-C and MPC-D'.
The shift from third to fourth gear is achieved in
the following manner with the selector position
valve in the D drive range:
1 Multiplate clutches `MPC-B and MPC-D'
applied as for third gear.
2 Solenoid valves `MV1 and MV3' de-energized
and closed as for third gear.
3 Electronic pressure regulating valve `EPRV-1'
de-energized and partially closed, whereas
`EPRV-3' remains energized and open, both
valves operating as for third gear.
4 Electronic pressure regulating valve `EPRV-2'
now progressively energizes and opens, this
removes the control pressure from brake and
retaining valves `BV-E and RV-E' respectively.
Line pressure to brake valve `BV-E' is now
blocked causing the release (exhausting) of fluid
pressure via the brake valve `BV-E' and the disengagement of the multiplate brake `MPB-E'.
5 Fluid pressure now passes though to the multiplate clutch `MPC-C' via shift valves `SV-1 and
SV-2', and clutch-valve `CV-C'. Subsequently,
the multiplate clutch `MPC-C' is applied to
complete the gear shift from third to fourth gear.
6 Electronic pressure regulating valve `EPRV-4'
de-energizes and progressively closes. Control
pressure now shifts converter pressure control
fluid pressure now passes to the right-hand end
of the clutch valve `CV-E' and retaining valve
`RV-E', thus causing both valves to shift to the
left-hand end. Fluid pressure is now permitted to
apply the multiplate brake `MPB-E' to complete
the engagement of fifth gear.
6 Electronic pressure regulating valve `EPRV-4'
de-energized as for fourth gear. This causes
the converter lock-up clutch `CLC' to engage
thereby by-passing the torque converter `TC'
fluid drive.
with a front mounted two speed `splitter' gear
change and a rear positioned single stage two
speed epicyclic gear `range' change; however, the
basic concept has been modified and considerably
simplified in this text.
Gear changes are achieved by four pneumatically operated power cylinders and pistons which
are attached to the ends of the three selector rods,
there being one power cylinder and piston for each
of the splitter and range selector rods and two for
the three speed and reverse constant mesh two
piece selector rod. Gear shifts are actuated by
inlet and exhaust solenoid control valves which
supply and release air to the various shift power
cylinders as required (see Fig. 5.48).
A multiplate transmission brake with its inlet
and exhaust solenoid control valves are provided
to shorten the slow down period of the clutch,
input shaft and twin countershaft assembly during
the gear change process.
A single plate dry friction clutch is employed but
instead of having a conventional clutch pedal to control the engagement and disengagement of the power
flow, a pneumatic operated clutch actuator with inlet
and exhaust solenoid control valves are used. Thus
the manual foot control needed for driving away
from rest and changing gear is eliminated.
Gradual engagement of the power flow via the
clutch when pulling away from a standstill and
smooth gear shift changes are achieved via the
wheel speed and engine speed sensors, air pressure
sensors and the electronic diesel control unit
(EDCU): this being part of the diesel engine management equipment, they all feed signals to the
electronic transmission control unit (ETCU). This
information is then processed so that commands to
the various solenoid control valves can be made
to produce the appropriate air pressure delivery
and release to meet the changing starting and
driving conditions likely to be experienced by a
transmission system. A gear selector switch control
stick provides the driver with a hand control which
instructs the electronic transmission control unit
(ETCU) to make an up and down gear shift when
prevailing engine torque and road resistance conditions are matched.
Reverse gear (Fig. 5.46) Engagement of reverse
gear is obtained by applying the multiplate clutch
`MPC-A' and the multiplate brakes `MPB-F and
The shift from neutral to reverse gear is achieved
in the following manner with the selector position
valve `SPV' moved to reverse drive position `R'.
1 Multiplate brake `MPB-G' applied as for neutral
and park position.
2 Solenoid valve `MV1' energized thus opening the
valve. Constant fluid pressure now moves shift
valves `SV-1 and SV-3' over to the right-hand
3 Electronic pressure regulating valve `EPRV-1'
de-energized as for neutral position.
4 Electronic pressure regulating valves `EPRV-3'
de-energized and closed. Controlling fluid pressure is relayed to the brake valve `BV-G' and
retaining valve `RV-G'. Both valves shift to
the left-hand side thus permitting fluid pressure
to reach and apply the multiplate brake
5 Selector position valve `SPV' in reverse position
diverts fluid pressure from the fluid pump,
directly to multiplate clutch `MPC-A' and
indirectly to multiplate brake `MPB-F' via the
selector position valve `SPV', reverse gear
valve `RGV', shift valve `SV-2' and the clutch
valve `CV-F'. Both multiplate clutch `MPC-A'
and multiplate brake `MPB-F' are therefore
5.11 Semi-automatic (manual gear change two
pedal control) transmission system
5.11.2 Splitter gear change stage (Fig. 5.47)
Power flows via the clutch and input shaft to the
splitter synchromesh dog clutch. The splitter synchromesh dog clutch can engage either the left or
right hand matching dog clutch teeth on the central
splitter gear mounted on the input shaft to obtain a
low splitter gear ratio, or to the central third gear
5.11.1 Description of transmission system
(Fig. 5.48)
The system being described is broadly based on the
ZF Man Tip Matic/ZF AS Tronic 12 speed twin
countershaft three speed constant mesh gearbox
Power flow path
Range shift
power cylinder
single stage
range gear box
Constant mesh
three speed and
reverse gear box
Selector rod
plunger & spring
power cylinder
Splitter shift
power cylinder
Fig. 5.47
rod and
selector rod
and fork
rod and
H selector
rod and
1 and R
Twin countershaft 12 speed constant mesh gearbox with synchromesh two speed splitter and range changes
mounted on the mainshaft to obtain the high splitter gear ratio. Power is now able to pass via the
twin countershafts to each of the mainshaft constant mesh central gears by way of the constant
mesh gears 1, 2, 3 and R.
load by the electronic diesel control unit `EDCU'
which is part of the diesel engine's fuel injection
equipment, and engine speed is also monitored by
the EDCU, whereas vehicle speed or wheel speed is
monitored by the wheel brake speed sensors. These
three factors are continuously being monitored, the
information is then passed on to the electronic
transmission control unit `ETCU' which processes
it so that commands can be transferred in the form
of electric current to the inlet and exhaust clutch
actuator solenoid control valves.
5.11.3 Constant mesh 1-2-3 and R gear stage
(Fig. 5.47)
The selection and engagement of one of the sliding
dog clutch set of teeth either with R, 1, 2 or 3
floating mainshaft central constant mesh gears permits the drive path to flow from the twin countershaft gears via the mainshaft to the epicylic range
change single stage gear train.
Engagement and disengagement of clutch when
pulling away from a standstill (Fig. 5.48) With
the vehicle stationary, the ignition switched on
and first gear selected, the informed ETCU energizes and opens the clutch solenoid inlet control
valve whereas the exhaust control valve remains
closed (see Fig. 5.48). Compressed air now enters
the clutch cylinder actuator, this pushes the piston
and rod outwards causing the clutch lever to pivot
and to pull back the clutch release bearing and
sleeve. As a result the clutch drive disc plate and
input shaft to the gearbox will be disengaged from
the engine. As the driver depresses the accelerator
pedal the engine speed commences to increase
(monitored by the engine speed sensor), the
ETCU now progressively de-energizes the solenoid controlled clutch inlet valve and conversely
energizes the solenoid controlled exhaust valve.
The steady release of air from the clutch actuator
cylinder now permits the clutch lever, release
bearing and sleeve to move towards the engagement
position where the friction drive plate is progressively squeezed between the flywheel and the clutch
pressure plate. At this stage the transmission drive
can be partially or fully taken up depending upon
the combination of engine speed, load and wheel
As soon as the engine speed drops below some
predetermined value the ETCU reacts by de-energizing and closing the clutch exhaust valve and
energizing and opening the clutch inlet valve, thus
compressed air will again enter the clutch actuator
cylinder thereby causing the friction clutch drive
plate to once more disengage.
Note a built-in automatic clutch re-adjustment
device and wear travel sensor is normally incorporated within the clutch unit.
5.11.4 Range change gear stage (Fig. 5.47)
Low range gear selection With the synchromesh
dog clutch hub moved to the left-hand side, the
internal toothed annular gear (A) will be held stationary; the drive from floating mainshaft is therefore compelled to pass from the central sun gear (S)
to the output shaft via the planet gear carrier (CP)
(see Fig. 5.47). Now since the annular gear is held
stationary, the planet gears (P) are forced to rotate
on their axes and also to roll around the internal
teeth of the annular gear (A), consequently the
planet carrier (CP) and output shaft will now rotate
at a lower speed than that of the sun gear (S) input.
High range gear selection With the syncromesh
dog clutch hub moved to the right-hand side, the
annular gear (A) becomes fixed to the output shaft,
therefore the drive to the planet gears (P) via the
floating mainshaft and sun gear (S) now divides
between the planet gear carrier (CP) and the annular gear carrier (CA) which are both fixed to the
output shaft (see Fig. 5.47). As a result the planet
gears (P) are prevented from rotating on their axes
so that while the epicyclic gear train is compelled to
revolve as one rigid mass, it therefore provides a
one-to-one gear ratio stage.
5.11.5 Clutch engagement and disengagement
(Fig. 5.48)
With the ignition switched on and the first gear
selected the clutch will automatically and progressively take up the drive as the driver depresses the
accelerator pedal. The three basic factors which
determine the smooth engagement of the transmission drive are vehicle load, which includes pulling away from a standstill and any road gradient,
vehicle speed and engine speed. Thus the vehicle's
resistance to move is monitored in terms of engine
Engagement and disengagement of the clutch during
a gear change (Fig. 5.48) When the driver moves
the gear selector stick into another gear position
Exhaust valve open Inlet valve closed (IVC)
Range shift
solenoid control valves
Low gear
Constant mesh 3–2 shift
solenoid control valves
Constant mesh 1-R shift
solenoid control valves
1-R shift
power cylinder
Splitter shift
solenoid control
Splitter shift
power cylinder
3 2
Low gear
1 R
Inlet valve
Exhaust valve
speed sensor
Inlet valve
Single disengaged
plate dry
Fig. 5.48
Clutch actuator cylinder
A simplified electro/pneumatic gear shift and clutch control
Exhaust valve
with the vehicle moving forwards, the ETCU
immediately signals the clutch solenoid control
valves to operate so that the compressed air can
bring about the disengagement and then engagement of the clutch drive plate for sufficient time
(programmed time setting) for the gear shift to take
place (see Fig. 5.48). This is achieved in the first
phase by de-energizing and closing the clutch solenoid exhaust valve and correspondingly energizing
and opening the inlet valve, thus permitting the
compressed air to enter the clutch actuator cylinder and to release the clutch. The second phase
de-energizes and closes the inlet valve and then
energizes and opens the exhaust valve so that the
clutch release mechanism allows the clutch to
engage the transmission drive.
close and open the inlet and exhaust valves respectively for the high splitter gear solenoid control,
and at the same time to close and open the exhaust
and inlet valves respectively for the low splitter gear
solenoid control (see Fig. 5.48). The splitter shift
power cylinder will now operate, compressed air
will be released from the left-hand side and
simultaneously compressed air will be introduced to the right-hand side of the splitter
shift power cylinder; the piston and selector
rod now smoothly shift to the low splitter gear
position. Conversely if high splitter gear was to
be selected, the reverse would happen to the
solenoid control valves with regards to their
opening and closing so that the piston and selector
rod would in this case move to the right.
5.11.6 Transmission brake (Figs 5.47 and 5.48)
This is a compressed air operated multiplate brake.
Its purpose is to rapidly reduce the free spin speed
of the driven disc plate, input shaft and twin countershaft masses when the clutch is disengaged thus
enabling fast and smooth gear shifts to be made.
When a gear shift change is about to be made the
driver moves the gear selector stick to a new position. This is signalled to the ETCU, and one outcome is that the transmission brake solenoid
control inlet valve is energized to open (see Fig.
5.48). It thus permits compressed air to enter the
piston chamber and thereby to squeeze together the
friction disc plate so that the freely spinning countershafts are quickly dragged down to the main shaft's
speed, see Fig. 5.47. Once the central gears wedged
in between the twin countershafts have unified
their speed with that of the mainshaft, then at this
point the appropriate constant mesh dog clutch can
easily slide into mesh with it adjacent central gear
dog teeth. Immediately after the gear shift the
transmission brake inlet valve closes and the
exhaust valve opens to release the compressed air
from the multiplate clutch cylinder thereby preventing excessive binding and strain imposed to
the friction plates and assembly.
5.11.8 Range gear shifts (Figs 5.47 and 5.48)
The range gear shift takes place though a single
stage epicyclic gear train and operates also via a
synchromesh type dog clutch mechanism.
Going though the normal gear change sequence
from 1 to 12 the first six gear ratios one to six are
obtained with the range gear shift in the low position and from seventh to twelfth gear in high range
shift position, see Fig. 5.47.
With the ignition switched on and the gear selector stick moved to gear ratios between one and six
the low range gear shift will be selected, the ETCU
activates the range shift solenoid control valves
such that the high range inlet and exhaust valves
are closed, and opened respectively, whereas the
low range inlet valve is opened and exhaust valve
is closed, see Fig. 5.48. Hence compressed air is
exhausted from the left hand side of the range
shift power cylinder and exposed to fresh compressed air on the right-hand side. Subsequently
the piston and selector rod is able to quickly shift
to the low range position.
A similar sequence of events takes place if the
high range gear shift is required except the opening
and closing of the valves will be opposite to that
needed for the low range shift.
5.11.7 Splitter gear shifts (Fig. 5.48)
The splitter gear shift between low and high gear
ratio takes place though a synchromesh type dog
clutch device. Note for all the gear changes taking
place in the gearbox, the splitter gears are constantly shifted from low to high going up the gear
ratios or from high to low going down the gear
ratios. With ignition switched on and the gear
selector stick positioned say in low gear, the
ETCU signals the splitter solenoid control to
5.11.9 Constant mesh three speed and reverse
gear shift (Figs 5.47 and 5.48)
These gear shifts cover the middle section of the
gearbox which involves the engagement and disengagement of the various central mainshaft constant
mesh gears via a pair of sliding dog clutches. There
is a dog clutch for engagement and disengagement
for gears 1-R and similarly a second dog clutch for
gears 2±3.
To go though the complete gear ratio steps,
the range shift is put initially into `low', then the
splitter gear shifts are moved alternatively into low
and high as the constant mesh dog clutch gears are
shifted progressively up; this is again repeated but
the second time with the range shift in high (see
Fig. 5.47). This can be presented as range gear
shifted into `low', 1 gear constant mesh low and
high splitter, 2 gear constant mesh low and high
splitter, and 3 gear constant mesh low and high
splitter gear; at this point the range gear is shifted
into `high' and the whole sequence is repeated,
1 constant mesh gear low and high splitter,
2 constant mesh gear low and high splitter and
finally third constant mesh gear low and high splitter; thus twelve gear ratios are produced thus:
where OGR = overall gear ratio
CM = constant mesh gear ratio
LS/HS = low or high splitter gear ratio
LS/HR = low or high range gear ratio
Assume that the ignition is switched on and the
vehicle is being driven forwards in low splitter and
low range shift gear positions (see Fig. 5.48). To
engage one of the three forward constant mesh
gears, for example, the second gear, then the gear
selector stick is moved into 3 gear position (low
splitter, low range 2 gear). Immediately the
ETCU signals the constant mesh 3±2 shift solenoid
control valves by energizing the 2 constant mesh
solenoid control so that its inlet valve opens and its
exhaust valve closes; at the same time, the 3 constant mesh solenoid control is de-energized so that
its inlet valve closes and the exhaust valve opens
(see Fig. 5.48). Accordingly, the 2±3 shift power
cylinder will be exhausted of compressed air on
the right-hand side, while compressed air is delivered to the left-hand side of the cylinder, the difference in force between the two sides of the piston
will therefore shift the 3±2 piston and selector rod
into the second gear position. It should be remembered that during this time period, the clutch will
have separated the engine drive from the transmission and that the transmission brake will have
slowed the twin countershafts sufficiently for the
constant mesh central gear being selected to equalize its speed with the mainshaft speed so that a
clean engagement takes place. If first gear was
then to be selected, the constant mesh 3±2 shift
solenoid control valves would both close their
exhaust valves so that compressed air enters from
both ends of the 2±3 shift power cylinder, it therefore moves the piston and selector rod into the
neutral position before the 1-R shift solenoid control valves are allowed to operate.
First six overall gear ratios = splitter gear (L and
H)S constant mesh gears (1, 2 and 3) range
gear low (LR)
Second six overall gear ratios = splitter gear
(L and H) constant mesh gears (1, 2 and 3) range gear high (HR).
OGR ˆ LS CM …1† LR
OGR ˆ HS CM …1† LR
OGR ˆ LS CM …2† LR
OGR ˆ LS CM …2† LR
OGR ˆ LS CM …3† LR
OGR ˆ HS CM …3† LR
Low range
OGR ˆ LS CM …1† HR
OGR ˆ HS CM …1† HR
OGR ˆ LS CM …2† HR
OGR ˆ HS CM …2† HR
OGR ˆ LS CM …3† HR
OGR ˆ HS CM …3† HR
High range
Transmission bearings and constant velocity joints
6.1 Rolling contact bearings
Bearings which are designed to support rotating
shafts can be divided broadly into two groups; the
plain lining bearing, known as the journal bearing,
and the rolling contact bearing. The fundamental
difference between these bearings is how they
provide support to the shaft. With plain sleeve or
lining bearings, metal to metal contact is prevented
by the generation of a hydrodynamic film of lubricant (oil wedge), which supports the shaft once
operating conditions have been established. However, with the rolling contact bearing the load is
carried by balls or rollers with actual metal to metal
contact over a relatively small area.
With the conventional journal bearing, starting
friction is relatively high and with heavy loads the
coefficient of friction may be in the order of 0.15.
However, with the rolling contact bearing the starting friction is only slightly higher than the operating friction. In both groups of bearings the
operating coefficients will be very similar and may
range between 0.001 and 0.002. Hydrodynamic
journal bearings are subjected to a cyclic projected
pressure loading over a relatively large surface area
and therefore enjoy very long life spans. For example, engine big-ends and main journal bearings may
have a service life of about 160 000 kilometres
(100 000 miles). Unfortunately, the inherent nature
of rolling contact bearing raceway loading is of a
number of stress cycles of large magnitude for each
revolution of the shaft so that the life of these
bearings is limited by the fatigue strength of the
bearing material.
Lubrication of plain journal bearings is very
important. They require a pressurized supply of
consistent viscosity lubricant, whereas rolling contact bearings need only a relatively small amount of
lubricant and their carrying capacity is not sensitive to changes in lubricant viscosity. Rolling contact bearings have a larger outside diameter and are
shorter in axial length than plain journal bearings.
Noise levels of rolling contact bearings at high
speed are generally much higher than for plain
journal bearings due mainly to the lack of a hydrodynamic oil film between the rolling elements and
their tracks and the windage effects of the ball or
roller cage.
6.1.1 Linear motion of a ball between two flat
tracks (Fig. 6.1)
Consider a ball of radius rb placed between an
upper and lower track plate (Fig. 6.1). If the
upper track plate is moved towards the right so
that the ball completes one revolution, then the
ball has rolled along the lower track a distance
2rb and the upper track has moved ahead of the
ball a further distance 2rb. Thus the relative movement, L, between both track plates will be
2rb ‡ 2rb , which is twice the distance, l, travelled
forward by the centre of the ball. In other words,
the ball centre will move forward only half that of
the upper to lower relative track movement.
2rb 1
L 4rb 2
6.1.2 Ball rolling contact bearing friction
(Fig. 6.2(a and b))
When the surfaces of a ball and track contact under
load, the profile a±b±c of the ball tends to flatten out
and the profile a±e±c of the track becomes concave
(Fig. 6.2(a)). Subsequently the pressure between the
contact surfaces deforms them into a common elliptical shape a±d±c. At the same time, a bulge will be
established around the contact edge of the ball due to
the displacement of material.
If the ball is made to roll forward, the material in
the front half of the ball will be subjected to
increased compressive loading and distortion whilst
that on the rear half experiences pressure release
(Fig. 6.2(b)). As a result, the stress distribution
over the contact area will be constantly varying.
The energy used to compress a perfect elastic
material is equal to that released when the load is
removed, but for an imperfect elastic material (most
materials), some of the energy used in straining the
material is absorbed as internal friction (known as
elastic hysteresis) and is not released when the load is
removed. Therefore, the energy absorbed by the ball
and track when subjected to a compressive load,
causing the steel to distort, is greater than that
released as the ball moves forward. It is this missing
Fig. 6.1 Relationship of rolling element and raceway movement
Fig. 6.2 (a and b)
Illustration of rolling ball resistance against motion
energy which creates a friction force opposing the
forward motion of the ball.
Owing to the elastic deformation of the contact
surfaces of the ball and track, the contact area will
no longer be spherical and the contact profile arc
will therefore have a different radius to that of the
ball (Fig. 6.2(b)). As a result, the line a±e±c of the
undistorted track surface is shorter in length than
the rolling arc profile a±d±c. In one revolution the
ball will move forward a shorter distance than if the
ball contact contour was part of a true sphere. Hence
the discrepancy of the theoretical and actual forward
movement of the ball is accommodated by slippage
between the ball and track interfaces.
ally grooved ring races (tracks). Lodged between
these inner and outer members are a number of
balls which roll between the grooved tracks when
relative angular motion of the rings takes place
(Fig. 6.3(a)). A fourth important component
which is not subjected to radial load is the ball
cage or retainer whose function is to space the
balls apart so that each ball takes its share of load
as it passes from the loaded to the unloaded side of
the bearing. The cage prevents the balls piling up
and rubbing together on the unloaded bearing side.
Contact area The area of ball to track groove contact will, to some extent, determine the load carrying capacity of the bearing. Therefore, if both ball
and track groove profiles more or less conform, the
bearing load capacity increases. Most radial ball
bearings have circular grooves ground in the inner
6.1.3 Radial ball bearings (Fig. 6.3)
The essential elements of the multiball bearing is
the inner externally grooved and the outer intern194
and outer ring members, their radii being 2±4%
greater than the ball radius so that ball to track
contact, friction, lubrication and cooling can be
controlled (Fig. 6.3(a)). An unloaded bearing produces a ball to track point contact, but as the load
is increased, it changes to an elliptical contact area
(Fig. 6.3(a)). The outer ring contact area will be
larger than that of the inner ring since the track
curvature of the outer ring is in effect concave and
that of the inner ring is convex.
be at a pitch circle radius rp and revolving at Nc
Linear speed of ball
ˆ 2rb Nb (m=s)
Linear speed of inner race ˆ 2ri Ni (m=s)
Linear speed of cage
ˆ 2rp Nc (m=s)
ri ‡ ro
Pitch circle radius rp
But the linear speed of the cage is also half
speed of the inner race
2ri Ni
Hence Linear speed of
Bearing failure The inner ring face is subjected to
a lesser number of effective stress cycles per revolution of the shaft than the corresponding outer ring
race, but the maximum stress developed at the
inner race because of the smaller ball contact area
as opposed to the outer race tends to be more
critical in producing earlier fatigue in the inner
race than that at the outer race.
ˆ ri Ni (m=s)
If no slip takes place,
Linear speed of ball ˆ Linear speed of inner
2rb Nb ˆ 2ri Ni
; Nb ˆ Ni (rev=min) (6)
Lubrication Single and double row ball bearings
can be externally lubricated or they may be prepacked with grease and enclosed with side covers to
prevent the grease escaping from within and at the
same time stopping dust entering the bearing
between the track ways and balls.
Linear speed of cage ˆ Half inner speed of
inner race
2rp Nc ˆ ri Ni
Hence Nc ˆ
2 rp
; Nc ˆ
6.1.4 Relative movement of radial ball bearing
elements (Fig. 6.3(b))
The relative movements of the races, ball and cage
may be analysed as follows:
Consider a ball of radius rb revolving Nb rev/min
without slip between an inner rotating race of
radius ri and outer stationary race of radius ro
(Fig. 6.3(b)). Let the cage attached to the balls
Fig. 6.3 (a and b)
ri N i
(rev=min) (7)
rp 2
Example A single row radial ball bearing has
an inner and outer race diameter of 50 and 70 mm
If the outer race is held stationary and the inner
race rotates at 1200 rev/min, determine the following information:
Deep groove radial ball bearing terminology
(a) The number of times the balls rotate for one
revolution of the inner race.
(b) The number of times the balls rotate for them
to roll round the outer race once.
(c) The angular speed of balls.
(d) The angular speed of cage.
(a) Diameter of balls ˆ ro
ˆ 35 25 ˆ 10 mm
Assuming no slip,
Number of Ball
Inner race
ball rotations circumference circumference
Number of ball
rotations, 2rb
ˆ 2ri
; Number of ball
2ri ri
2rb rb
ˆ 5 revolutions
(b) Number of Ball
Outer race
ball rotations circumference circumference
Fig. 6.4
Number of ball
rotations, 2rb ˆ2ro
2ro ro
; Number of rotations ˆ
2rb rb
ˆ ˆ 7 revolutions
force which is known as a radial force. This
kind of loading could be caused by pairs of
meshing spur gears radially separating from
each other when transmitting torque (Fig. 6.4).
2 A load (force) applied parallel to the shaft and
bearing axis. This produces an end thrust which
is known as an axial force. This kind of loading
could be caused by pairs of meshing helical gears
trying to move apart axially when transmitting
torque (Fig. 6.4).
ˆ 1200
ˆ 6000 rev=min
(c) Ball angular speed Nb ˆ
Illustration of radial and axial bearing loads
When both radial and axial loads are imposed on
a ball bearing simultaneously they result in a single
combination load within the bearing which acts
across the ball as shown (Fig. 6.6).
ri ‡ ro 25 ‡ 35
ˆ 30 mm
6.1.6 Ball and roller bearing internal clearance
Internal bearing clearance refers to the slackness
between the rolling elements and the inner and
outer raceways they roll between. This clearance is
measured by the free movement of one raceway ring
relative to the other ring with the rolling elements in
between (Fig. 6.5). For ball and cylindrical (parallel) roller bearings, the radial or diametrical clearance is measured perpendicular to the axis of the
bearing. Deep groove ball bearings also have axial
clearance measured parallel to the axis of the bearing. Cylindrical (parallel) roller bearings without
inner and outer ring end flanges do not have axial
clearance. Single row angular contact bearings and
ri N1
rp 2
25 1200
ˆ 500 rev=min
Cage angular speed Nc ˆ
6.1.5 Bearing loading
Bearings used to support transmission shafts are
generally subjected to two kinds of loads:
1 A load (force) applied at right angles to the shaft
and bearing axis. This produces an outward
cast iron housing or a high strength steel housing?
Is it a solid or hollow shaft; are the inner and outer
ring member sections thin, medium or thick?
b) The type of housing or shaft fit; is it a light,
medium or heavy interference fit?
The diametric clearance reduction when an inner
ring is forced over a solid shaft will be a proportion
of the measured ring to shaft interference.
The reductions in diametric clearance for a heavy
and a thin sectioned inner raceway ring are roughly
50% and 80% respectively. Diametric clearance
reductions for hollow shafts will of course be less.
Working bearing clearances are affected by the
difference in temperature between the outer and
inner raceway rings which arise during operation.
Because the inner ring attached to its shaft is not
cooled so effectively as the outer ring which is
supported in a housing, the inner member expands
more than the outer one so that there is a tendency
for the diametric clearance to be reduced due to the
differential expansion of the two rings.
Another reason for having an initial diametric
clearance is it helps to accommodate any inaccuracies in the machining and grinding of the bearing
The diametric clearance affects the axial clearance of ball bearings and in so doing influences
their capacity for carrying axial loads. The greater
the diametric clearance, the greater the angle of ball
contact and therefore the greater the capacity for
supporting axial thrust (Fig. 6.6).
Bearing internal clearances have been so derived
that under operating conditions the existing clearances provide the optimum radial and axial load
carrying capacity, speed range, quietness of running and life expectancy. As mentioned previously,
the diametric clearance is greatly influenced by the
type of fit between the outer ring and its housing
and the inner ring and its shaft, be they a slip, push,
light press or heavy press interference.
The tightness of the bearing fit will be determined
by the extremes of working conditions to which the
bearing is subjected. For example, a light duty application will permit the bearing to be held with a
relatively loose fit, whereas for heavy conditions
an interference fit becomes essential.
To compensate for the various external fits and
applications, bearings are manufactured with
different diametric clearances which have been
standardized by BSI and ISO. Journal bearings
are made with a range of diametrical clearances,
these clearances being designated by a series of
codes shown below in Table 6.1.
Fig. 6.5 Internal bearing diametric clearance
taper roller bearings do have clearance slackness or
tightness under operating conditions but this cannot be measured until the whole bearing assembly
has been installed in its housing.
A radial ball bearing working at operating temperature should have little or no diametric clearance,
whereas roller radial bearings generally operate
more efficiently with a small diametric clearance.
Radial ball and roller bearings have a much
larger initial diametric clearance before being fitted
than their actual operating clearances.
The difference in the initial and working diametric clearances of a bearing, that is, before and
after being fitted, is due to a number of reasons:
1 The compressive interference fit of the outer
raceway member when fitted in its housing
slightly reduces diameter.
2 The expansion of the inner raceway member
when forced over its shift minutely increases its
The magnitude of the initial contraction or
expansion of the outer and inner raceway members
will depend upon the following:
a) The rigidity of the housing or shaft; is it a low
strength aluminium housing, moderate strength
Fig. 6.6 Effects of diametric clearance and axial load on angle of contact
Group 2 These bearings have the least diametric
clearance. Bearings of this group are suitable when
freedom from shake is essential in the assembled
bearing. The fitting interference tolerance prevents
the initial diametric clearance being eliminated.
Very careful attention must be given to the bearing
housing and shaft dimensions to prevent the expansion of the inner ring or the contraction of the outer
ring causing bearing tightness.
Table 6.1 Journal bearing diametrical clearances
Group 2
Normal group
Group 3
Group 4
Note The lower the number the smaller is the
bearing's diametric clearance. In the new edition of
BS 292 these designations are replaced by the ISO
groups. For special purposes, bearings with a smaller
diametric clearance such as Group 1 and larger
Group 5 are available.
The diametrical clearances 0, 00, 000 and 0000
are usually known as one dot, two dot, three dot or
four dot fits. These clearances are identified by the
appropriate code or number of polished circles on
the stamped side of the outer ring.
The applications of the various diametric clearance groups are compared as follows:
Normal group Bearings in this group are suitable
when only one raceway ring has made an interference fit and there is no appreciable loss of clearance
due to temperature differences. These diametric
clearances are normally adopted with radial ball
bearings for general engineering applications.
Group 3 Bearings in this group are suitable when
both outer and inner raceway rings have made an
interference fit or when only one ring has an interference fit but there is likely to be some loss of
clearance due to temperature differences. Roller
bearings and ball bearings which are subjected to
axial thrust tend to use this diametric clearance
member rotated relative to the stationary outer
race members (Fig. 6.8(a)).
The design geometry of the taper roller bearing is
therefore based on the cone principle (Fig. 6.8(b))
where all projection lines, lines extending from the
cone and cup working surfaces (tracks), converge
at one common point on the axis of the bearing.
With the converging inner and outer raceway
(track) approach, the track circumferences at the
large and small ends of each roller will be greater
and smaller respectively. The different surface velocities on both large and small roller ends will be
accommodated by the corresponding change in
track circumferences. Hence no slippage takes
place, only pure rolling over the full length of
each roller as they revolve between their inner and
outer tracks.
Group 4 Bearings in this group are suitable when
both outer and inner bearing rings are an interference fit and there is some loss of diametric clearance
due to temperature differences.
6.1.7 Taper roller bearings
Description of bearing construction (Fig. 6.7) The
taper roller bearing is made up of four parts; the
inner raceway and the outer raceway, known
respectively as the cone and cup, the taper rollers
shaped as frustrums of cones and the cage or roller
retainer (Fig. 6.8). The taper rollers and both inner
and outer races carry load whereas the cage carries
no load but performs the task of spacing out the
rollers around the cone and retaining them as an
Angle of contact (Fig. 6.7) Taper roller bearings
are designed to support not only radial bearing
loads but axial (thrust) bearing loads.
The angle of bearing contact , which determines the maximum thrust (axial) load, the bearing
can accommodate is the angle made between the
perpendiculars to both the roller axis and the inner
cone axis (Fig. 6.7). The angle of contact is also
half the pitch cone angle. These angles can range
from as little as 7‰ to as much as 30 . The standard or normal taper roller bearing has a contact
angle of 12±16 which will accommodate moderate
thrust (axial) loads. For large and very large thrust
loads, medium and steep contact angle bearings
are available, having contact angles in the region
of 20 and 28 respectively.
Taper roller bearing true rolling principle (Fig.
6.8(a and b)) If the axis of a cylindrical (parallel)
roller is inclined to the inner raceway axis, then the
relative rolling velocity at the periphery of both the
outer and inner ends of the roller will tend to be
different due to the variation of track diameter
(and therefore circumference) between the two
sides of the bearing. If the mid position of the roller
produced true rolling without slippage, the portion
of the roller on the large diameter side of the tracks
would try to slow down whilst the other half of the
roller on the smaller diameter side of the tracks
would tend to speed up. Consequently both ends
would slip continuously as the central raceway
Area of contact (Fig. 6.7) Contact between roller
and both inner cone and outer cup is of the line form
without load, but as the rollers become loaded the
elastic material distorts, producing a thick line contact area (Fig. 6.7) which can support very large
combinations of both radial and axial loads.
Cage (Fig. 6.7) The purpose of the cage container
is to equally space the rollers about the periphery of
the cone and to hold them in position when the bearing is operating. The prevention of rolling elements
touching each other is important since adjacent rollers move in opposite directions at their points of
closest approach. If they were allowed to touch they
would rub at twice the normal roller speed.
The cage resembles a tapered perforated sleeve
(Fig. 6.7) made from a sheet metal stamping which
Fig. 6.7 Taper rolling bearing terminology
Fig. 6.8 Principle of taper rolling bearing
has a series of roller pockets punched out by a single
impact of a multiple die punch.
Although the back cone rib contributes most to
the alignment of the rollers, the bearing cup and
cone sides furthest from the point of bearing loading may be slack and therefore may not be able to
keep the rollers on the unloaded side in their true
plane. Therefore, the cage (container) pockets are
precisely chamfered to conform to the curvature of
the rollers so that any additional corrective alignment which may become necessary is provided by
the individual roller pockets.
between the tapered faces of the cup and cone
without the guidance of the cage. The magnitude
of the roller-to-rib end thrust, known as the seating
force, will depend upon the taper roller contact
Positive roller alignment (Fig. 6.9) Both cylindrical parallel and taper roller elements, when rolling
between inner and outer tracks, have the tendency to
skew (tilt) so that extended lines drawn through
their axes do not intersect the bearing axis at the
same cone and cup projection apex. This problem
has been overcome by grinding the large end of each
roller flat and perpendicular to its axis so that all the
rollers square themselves exactly with a shoulder or
rib machined on the inner cone (Fig. 6.9). When
there is any relative movement between the cup
and cone, the large flat ends of the rollers make
contact with the adjacent shoulder (rib) of the cone,
compelling the rollers to positively align themselves
Fig. 6.9
Roller self-alignment
be made from either a case hardening alloy steel or
a through hardened alloy steel.
a) The case hardened steel is usually a low alloy
nickel chromium or nickel-chromium molybdenum steel, in which the surface only is hardened
to provide a wear resistance outer layer while
the soft, more ductile core enables the bearing to
withstand extreme shock and overloading.
b) The through hardened steel is generally a high
carbon chromium steel, usually about 1.0%
carbon for adequate strength, together with
1.5% chromium to increase hardenability. (This
is the ability of the steel to be hardened all the
way through to a 60±66 Rockwell C scale.)
The summary of the effects of the alloying
elements is as follows:
Nickel increases the tensile strength and toughness and also acts as a grain refiner. Chromium
considerably hardens and raises the strength with
some loss in ductibility, whilst molybdenum
reduces the tendency to temper-brittleness in low
nickel low chromium steel.
Bearing inner and outer raceways are machined
from a rod or seamless tube. The balls are produced by closed die forging of blanks cut from
bar stock, are rough machined, then hardened
and tempered until they are finally ground and
lapped to size.
Some bearing manufacturers use case-hardened
steel in preference to through-hardened steel
because it is claimed that these steels have hard
fatigue resistant surfaces and a tough crack-resistant core. Therefore these steels are able to withstand impact loading and prevent fatigue cracks
spreading through the core.
Fig. 6.10 Force diagram illustrating positive roller alignment seating force
Self-alignment roller to rib seating force (Fig. 6.10)
To make each roller do its full share of useful work,
positive roller alignment is achieved by the large
end of each roller being ground perpendicular to its
axis so that when assembled it squares itself exactly
with the cone back face rib (Fig. 6.10).
When the taper roller bearing is running under
operating conditions it will generally be subjected
to a combination of both radial and axial loads.
The resultant applied load and resultant reaction
load will be in apposition to each other, acting
perpendicular to both the cup and cone track
faces. Since the rollers are tapered, the direction
of the perpendicular resultant loads will be slightly
inclined to each other, they thereby produce a third
force parallel to the rolling element axis. This third
force is known as the roller-to-rib seating force and
it is this force which provides the rollers with their
continuous alignment to the bearing axis. The magnitude of this roller-to-rib seating force is a function
of the included taper roller angle which can be
obtained from a triangular force diagram (Fig.
6.10). The diagram assumes that both the resultant
applied and reaction loads are equal and that their
direction lies perpendicular to both the cup and cone
track surface. A small roller included angle will
produce a small rib seating force and vice versa.
6.1.9 Bearing friction
The friction resistance offered by the different kinds
of rolling element bearings is usually quoted in terms
of the coefficient of friction so that a relative comparison can be made. Bearing friction will vary to
some extent due to speed, load and lubrication.
Other factors will be the operating conditions
which are listed as follows:
1 Starting friction will be higher than the dynamic
normal running friction.
2 The quantity and viscosity of the oil or grease; a
large amount of oil or a high viscosity will
increase the frictional resistance.
3 New unplanished bearings will have higher
coefficient of friction values than worn bearings
which have bedded down.
6.1.8 Bearing materials
Bearing inner and outer raceway members and
their rolling elements, be they balls or rollers, can
4 Preloading the bearing will initially raise the
coefficient of friction but under working conditions it may reduce the overall coefficient value.
5 Pre-lubricated bearings may have slightly
higher coefficients than externally lubricated
bearings due to the rubbing effect of the seals.
around the bearing (Fig. 6.11(e)). Deep groove
radial ball bearings can tolerate light end thrust.
Angular contact ball bearings are capable of supporting medium axial loads. Taper roller bearings,
be they normal or steep angled, can operate continuously under heavy and very heavy end loads respectively. Only if the shaft being supported deflects will
the end load distribution become uneven.
Coefficient of friction Ð average values for various
bearing arrangements
Self-alignment ball bearings
= 0.001
Cylindrical roller bearing
= 0.0011
Thrustball bearings
= 0.0013
Single row deep grooveball bearings = 0.0015
Taper and spherical roller bearings = 0.0018
6.1.11 Bearing fatigue
Fatigue in ball or roller bearings is caused by
repeated stress reversals as the rolling elements
move around the raceways under load. The periodic elastic compression and release as the rolling
elements make their way around the tracks will
ultimately overwork and rupture the metal just
below the surface. As a result, tiny cracks propagate almost parallel to the surface but just deep
enough to be invisible. With continuous usage the
alternating stress cycles will cause the cracks to
extend, followed by new cracks sprouting out
from the original ones. Eventually there will be
a network of minute interlinking cracks rising and
merging together on the track surface. Subsequently, under further repeating stress cycles,
particles will break away from the surface, the size
of material leaving the surface becoming larger and
larger. This process is known as spalling of the
bearing and eventually the area of metal which
has come away will end the effective life of the
bearing. If bearing accuracy and low noise level is
essential the bearing will need to be replaced, but if
bearing slackness and noise can be accepted, the
bearing can continue to operate until the rolling
elements and their tracks find it impossible to support the load.
6.1.10 Ball and roller bearing load distribution
(Fig. 6.11)
When either a ball or roller bearing is subjected to
a radial load, the individual rolling elements will not
be loaded equally but will be loaded according to
their disposition to the direction of the applied load.
Applying a radial load to a bearing shaft pushes the
inner race towards the outer race in the direction of
the load so that the balls or rollers in one half of the
bearing do not support any load whereas the other
half of the bearing reacts to the load (Fig. 6.11(a)).
The distribution of load on the reaction side of the
bearing will vary considerably with the diametrical
rolling element clearance and the mounting rigidity
preventing deformation of the bearing assembly.
If the internal radial clearances of the rolling
elements are zero and the inner and outer bearing
races remain true circles when loaded, the load
distribution will span the full 180 so that approximately half the balls or rollers will, to some extent,
share the radial load (Fig. 6.11(b)). Conversely
slackness or race circular distortion under load will
reduce the projected load zone so that the rolling
elements which provide support will be very much
more loaded resulting in considerably more shaft
deflection under load. Lightly preloaded bearings
may extend the radial load zone to something
greater than 180 but less than 360 (Fig. 6.11(c)).
This form of initial bearing loading will eliminate
gear mesh teeth misalignment due to shaft deflection under operating conditions. Heavy bearing
preloading may extend the load zone to 360
(Fig. 6.11(d)); this degree of preloading should only
be used for severe working conditions or where large
end thrust is likely to be encountered and must be
absorbed without too much axial movement.
End thrusts (axial loads), unlike radial loads,
produce a uniform load distribution pattern
6.1.12 Rolling contact bearing types
Single row deep grooved radial ball bearing
(Fig. 6.12) These bearings are basically designed
for light to medium radial load operating conditions. An additional feature is the depth of the
grooves combined with the relatively large size of
the balls and the high degree of conformity between
balls and grooves which gives the bearing considerable thrust load carrying capacity so that the bearing will operate effectively under both radial and
axial loads.
These bearings are suitable for supporting gearbox primary and secondary shafts etc..
Single row angular contact ball bearing (Fig. 6.13)
Bearings of this type have ball tracks which are so
Fig. 6.11 (a±e)
Bearing radial and axial load distribution
Fig. 6.12 Single row deep groove radial ball bearing
Fig. 6.13
Single row deep angular contact ball bearing
Fig. 6.14
Double row angular contact ball bearing
disposed that a line through the ball contact forms
an acute angle with the bearing shaft axes. Ball to
track ring contact area is elliptical and therefore
with the inclined contact angle this bearing is particularly suitable for heavy axial loads. Adjustment
of these bearings must always be towards another
bearing capable of dealing with axial loads in the
opposite direction. The standard contact angle is
20 , but for special applications 12, 15, 25 and 30
contact angle bearings are available. These bearings
are particularly suited for supporting front and rear
wheel hubs, differential cage housings and steering
box gearing such as the rack and pinion.
Double row angular contact ball bearings (Fig. 6.14)
With this double row arrangement, the ball tracks
are ground so that the lines of pressure through the
balls are directed towards two comparatively
widely separated points on the shaft. These bearings are normally preloaded so that even when
subjected to axial loads of different magnitudes,
axial deflection of the shaft is minimized. End
thrust in both axial directions can be applied and
at the same time very large radial loads can be
carried for a relatively compact bearing assembly.
A typical application for this type of bearing
would be a semi- or three-quarter floating outer
propellor shaft support, half shafts and wherever
excessive shaft deflection is likely to occur.
Single row cylindrical roller bearing (Fig. 6.16) In
this design of roller bearing, the rollers are guided
by flanges, one on either the inner or outer track
ring. The other ring does not normally have a
flange. Consequently, these bearings do not take
axial loads and in fact permit relative axial deflection of shafts and bearing housing within certain
limits. These bearings can carry greater radial loads
than the equivalent size groove bearing and in some
applications both inner and outer ring tracks are
flanged to accommodate very light axial loads.
Bearings of this type are used in gearbox and
final drive transmissions where some axial alignment may be necessary.
Fig. 6.15
Single row taper roller bearing (Fig. 6.17) The
geometry of this class of bearing is such that the
axes of its rollers and conical tracks form an angle
with the shaft axis. The taper roller bearing is therefore particularly adaptable for applications where
large radial and axial loads are transmitted simultaneously. For very severe axial loads, steep taper
angle bearings are available but to some extent this
is at the expense of the bearing's radial load carrying capacity. With taper bearings, adjustment must
always be towards another bearing capable of
dealing with axial forces acting in the opposite
direction. This is a popular bearing for medium
and heavy duty wheel hubs, final drive pinion
shafts, the differential cage and crownwheel bearings, for heavy duty gearbox shaft support and
in-line injection pump camshafts.
Double row self-alignment ball bearing
half shaft bearing, gearbox secondary output shaft
bearing etc.
Double row self-aligning ball bearing (Fig. 6.15)
This double row bearing has two rows of balls
which operate in individual inner raceway grooves
in conjunction with a common spherical outer raceway ring. The spherical outer track enables the inner
ring and shaft to deflect relative to the outer raceway member, caused by the balls not only rolling
between and around their tracks but also across the
common outer circular track. Thus the self-aligning
property of the bearing automatically adjusts any
angular deflection of the shaft due to mounting
errors, whip or settlement of the mounting. It also
prevents the bearing from exerting a bending influence on the shaft. The radial load capacity for a
single row self-aligning bearing is considerably less
than that for the deep groove bearing due to the
large radius of the outer spherical race providing
very little ball to groove contact. This limitation was
solved by having two staggered rows of balls to
make up for the reduced ball contact area.
Note that double row deep groove bearings are
not used because radial loads would be distributed
unevenly between each row of balls with a periodic
shaft deflection. They are used for intermediate
Double row taper roller bearing (Fig. 6.18) These
bearings have a double cone and two outer single
cups with a row of taper rollers filling the gap
between inner and outer tracks on either side. The
compactness of these bearings makes them particularly suitable when there is very little space and
where large end thrusts must be supported in both
axial directions. Thus in the case of a straddled
final drive pinion bearing, these double row taper
bearings are more convenient than two single row
bearings back to back. Another application for
these double tow taper roller bearings is for transversely mounted gearbox output shaft support.
Double row spherical roller bearing (Fig. 6.19)
Two rows of rollers operate between a double
Fig. 6.17 Single row taper roller bearing
Fig. 6.16 Single row cylindrical roller bearing
Fig. 6.18 Double row taper roller bearing
Fig. 6.19
Double row spherical roller bearing
grooved inner raceway and a common spherically
shaped outer raceway ring. With both spherical
rollers having the same radii as the outer spherical
raceway, line contact area is achieved for both
inner and outer tracks. The inner double inclined
raceway ring retains the two rows of rollers within
their tracks, whereas the outer spherical track will
accommodate the rollers even with the inner track
ring axis tilted relative to the outer track ring axis.
This feature provides the bearing with its self-alignment property so that a large amount of shaft
deflection can be tolerated together with its capacity, due to roller to track conformity, to operate
with heavy loads in both radial and axial directions.
This type of bearing finds favour where both high
radial and axial loads are to be supported within
the constraints of a degree of shaft misalignment.
Single row thrust ball bearing (Fig. 6.20) These
bearings have three load bearing members, two
grooved annular disc plates and a ring of balls
lodged between them. A no-load-carrying cage
fourth member of the bearing has two functions;
firstly to ease assembly of the bearing when being
installed and secondly to evenly space the balls
around their grooved tracks. Bearings of this type
operate with one raceway plate held stationary
while the other one is attached to the rotating shaft.
In comparison to radial ball bearings, thrust ball
bearings suffer in operation from an inherent
increase in friction due to the balls sliding between
the grooved tracks. To minimize the friction, the
groove radii are made 6±8% larger than the radii
of the balls so that there is a reduction in ball contact
area. Another limitation of these bearings is that
they do not work very satisfactorily at high rotative
speeds since with increased speed the centrifugal
force pushes the balls radially outwards, so causing
the line of contact, which was originally in the middle of the grooves, to shift further out. This in effect
increases ball to track sliding and subsequently the
rise in friction generates heat. These bearings must
deal purely with thrust loads acting in one direction
and they can only tolerate very small shaft misalignment. This type of bearing is used for injection pump
governor linkage axial thrust loads, steering boxes
and auxiliary vehicle equipment.
Fig. 6.20
Single row thrust bearing
Fig. 6.21
Needle roller bearing
raceway ring. The outer raceway is shouldered
either side to retain the needles and has a circular
groove machined on the outside with two or four
radial holes to provide a passageway for needle
lubrication. The length to diameter ratio for the
needles usually lies between 3 to 8 and the needle
roller diameter normally ranges from 1.5 to 4.5 mm.
Sometimes there is no inner raceway ring and the
Needle roller bearings (Fig. 6.21) Needle roller
bearings are similar to the cylindrical roller bearing
but the needle rollers are slender and long and there
is no cage (container) to space out the needles
around the tracks. The bearing has an inner plain
There are three stages of friction clutch loading:
1 Belleville spring action,
2 Bevel gear separating force action,
3 Vee slot wedging action.
Belleville spring action (Fig. 7.11) This is achieved
by having one of the clutch plates dished to form a
Belleville spring so that there is always some spring
axial loading in the clutch plates. This then produces
a small amount of friction which tends to lock the
half shaft to the differential cage when the torque
transmitted is very low. The spring thus ensures that
when adhesion is so low that hardly any torque can
be transmitted, some drive will still be applied to the
wheel which is not spinning.
Bevel gear separating force action (Fig. 7.11) This
arises from the tendency of the bevel planet pinions
in the differential cage to force the bevel sun gears
outwards. Each bevel sun gear forms part of a hub
which is internally splined to the half shaft so that it
is free to move outwards. The sun gear hub is also
splined externally to align with one set of clutch
plates, the other set being attached by splines to the
differential cage. Thus the extra outward force
exerted by the bevel pinions when one wheel tends
to spin is transmitted via cup thrust plates to the
clutches, causing both sets of plates to be camped
together and thereby preventing relative movement
between the half shaft and cage.
Fig. 7.12 Comparison of tractive effort and tyre to road
adhesion for both conventional and limited slip differential
continuously, provided the friction of the multiplate clutches can be overcome. When one wheel
spins the traction of the other wheel is increased by
an amount equal to the friction torque generated
by the clutch plates until wheel traction is restored.
A comparison of a conventional differential and
a limited slip differential tractive effort response
against varying tyre to road adhesion is shown in
Fig. 7.12.
7.3.2 Torsen worm and wheel differential
Vee slot wedging action (Fig. 7.11(a and b)) When
the torque is increased still further, a third stage of
friction clutch loading comes into being. The bevel
pinions are not mounted directly in the differential
cage but rotate on two separate arms which cross at
right angles and are cranked to avoid each other.
The ends of these arms are machined to the shape of
a vee wedge and are located in vee-shaped slots in
the differential cage. With engine torque applied, the
drag reaction of the bevel planet pinion cross-pin
arms relative to the cage will force them to slide
inwards along the ramps framed by the vee-shaped
slots in the direction of the wedge (Fig. 7.11(a and b)).
The abutment shoulder of the bevel planet pinions
press against the cup thrust plates and each set of
clutch plates are therefore squeezed further together,
increasing the multiclutch locking effect.
Differential construction (Figs 7.13 and 7.14) The
Torsen differential has a pair of worm gears, the
left hand half shaft is splined to one of these worm
gears while the right hand half shaft is splined to
the other hand (Fig. 7.13). Meshing with each
worm gear on each side is a pair of worm wheels
(for large units triple worm wheels on each side). At
both ends of each worm wheel are spur gears which
mesh with adjacent spur gears so that both worm
gear and half shafts are indirectly coupled together.
Normally with a worm gear and worm wheel
combination the worm wheel is larger than the
worm gear, but with the Torsen system the worm
gear is made larger than the worm wheel. The
important feature of any worm gear and worm
wheel is that the teeth are cut at a helix angle such
that the worm gear can turn the worm wheel but the
worm wheel cannot rotate the worm gear. This is
achieved with the Torsen differential by giving the
Speed differential and traction control (Fig. 7.12)
Normal differential speed adjustment takes place
Fig. 7.13
Pictorial view of Torsen worm and spur gear differential
worm gear teeth a fine pitch while the worm wheel
has a coarse pitch.
Note that with the conventional meshing spur
gear, be it straight or helical teeth, the input and
output drivers can be applied to either gear. The
reversibility and irreversibility of the conventional
bevel gear differential and the worm and worm
wheel differential is illustrated in Fig. 7.14 by the
high and low mechanical efficiencies of the two
types of differential.
Differential action when moving straight ahead
(Fig. 7.15) When the vehicle is moving straight
ahead power is transferred from the propellor shaft
to the bevel pinion and crownwheel. The crownwheel and differential cage therefore revolve as one
unit (Fig. 7.15). Power is divided between the left
and right hand worm wheel by way of the spur gear
pins which are attached to the differential cage. It
then flows to the pair of meshing worm gears, where
it finally passes to each splined half shaft. Under
these conditions, the drive in terms of speed and
torque is proportioned equally to both half shafts
and road wheels. Note that there is no relative
rotary motion between the half shafts and the differential cage so that they all revolve as a single unit.
Fig. 7.14 Comparison of internal friction expressed in
terms of mechanical efficiency of both bevel pinion type
and worm and spur type differentials
tend to rotate faster than the inside wheel due to its
turning circle being larger than that of the inside
wheel. It follows that the outside wheel will have to
rotate relatively faster than the differential cage, say
by ‡20 rev/min, and conversely the inside wheel has
to reduce its speed in the same proportion, of say
20 rev/min.
Differential action when cornering (Fig. 7.15) When
cornering, the outside wheel of the driven axle will
Fig. 7.15 Sectioned views of Torsen worm and spur gear differential
When there is a difference in speed between the
two half shafts, the faster turning half shaft via the
splined worm gears drives its worm wheels about
their axes (pins) in one direction of rotation. The
corresponding slower turning half shaft on the
other side drives its worm wheels about their axes
(pins) in the opposite direction but at the same
speed (Fig. 7.15).
Since the worm wheels on opposite sides will be
revolving at the same speed but in the opposite sense
while the vehicle is cornering they can be simply
interlinked by pairs of meshing spur gears without
interfering with the independent road speed requirements for both inner and outer driving road wheels.
transmits drive from its set of worm gears to the
worm wheels. The drive is then transferred from
the worm wheels on the spinning side to the
opposite (good traction wheel) side worm wheels
by way of the bridging spur gears (Fig. 7.15). At
this point the engaging teeth of the worm wheel
with the corresponding worm gear teeth jam.
Thus the wheel which has lost its traction locks
up the gear mechanism on the other side every
time there is a tendency for it to spin. As a result
of the low traction wheel being prevented from
spinning, the transmission of torque from the
engine will be concentrated on the wheel which
has traction.
Another feature of this mechanism is that speed
differentiation between both road wheels is maintained even when the wheel traction differs considerably between wheels.
Differential torque distribution (Fig. 7.15) When
one wheel loses traction and attempts to spin, it
Fig. 7.16
Viscous coupling differential
7.3.3 Viscous coupling differential
ing torque will be proportional to the fluid viscosity
and the relative speed difference between the sets of
plates. The dilatent silicon compound fluid which
has been developed for this type of application has
the ability to maintain a constant level of viscosity
throughout the operating temperature range and
life expectancy of the coupling (Fig. 7.17).
Description of differential and viscous coupling
(Figs 7.16 and 7.17) The crownwheel is bolted to
the differential bevel gearing and multiplate housing. Speed differentiation is achieved in the normal
manner by a pair of bevel sun (side) gears, each
splined to a half shaft. Bridging these two bevel sun
gears are a pair of bevel planet pinions supported
on a cross-pin mounted on the housing cage.
A multiplate back assembly is situated around
the left hand half shaft slightly outboard from the
corresponding sun gear (Fig. 7.16).
The viscous coupling consists of a series of
spaced interleaved multiplates which are alternatively splined to a half shaft hub and the outer
differential cage. The cage plates have pierced
holes but the hub plates have radial slots. Both
sets of plates are separated from each other by a
0.25 mm gap. Thus the free gap between adjacent
plates and the interruption of their surface areas
with slots and holes ensures there is an adequate
storage of fluid between plates after the sealed plate
unit has been filled and that the necessary progressive viscous fluid torque characteristics will be
obtained when relative movement of the plates
takes place.
When one set of plates rotate relative to the
other, the fluid will be sheared between each pair
of adjacent plate faces and in so doing will generate
an opposing torque. The magnitude of this resist-
Fig. 7.17 Comparison of torque transmitted to wheel
having the greater adhesion with respect to speed
difference between half shafts for both limited slip and
viscous coupling
Speed differential action (Fig. 7.16) In the straight
ahead driving mode the crownwheel and differential cage driven by the bevel pinion act as the input
to the differential gearing and in so doing the
power path transfers to the cross-pin and bevel
planet gears. One of the functions of these planet
gears is to link (bridge) the two sun (side) gears so
that the power flow is divided equally between the
sun gears and correspondently both half shafts
(Fig. 7.16).
When rounding a bend or turning a corner, the
outer wheel will have a greater turning circle than
the inner one. Therefore the outer wheel tends to
increase its speed and the inner wheel decrease its
speed relative to the differential cage rotational
speed. This speed differential is made possible by
the different torque reactions each sun gear conveys back from the road wheel to the bevel planet
pinions. The planet gears `float' between the sun
gears by rotating on their cross-pin, thus the speed
lost relative to the cage speed by the inner road
wheel and sun gear due to the speed retarding
ground reaction will be that gained by the outer
road wheel and sun gear.
the torque transmitted to the wheel supplying tractive effort rises with increased relative speed
between the half shaft and differential cage.
7.4 Double reduction axles
7.4.1 The need for double reduction final drives
The gearbox provides the means to adjust and
match the engine's speed and torque so that the
vehicle's performance responds to the driver's
expectations under the varying operating conditions. The gearbox gear reduction ratios are inadequate to supply the drive axle with sufficient
torque multiplication and therefore a further permanent gear reduction stage is required at the drive
axle to produce the necessary road wheel tractive
effect. For light vehicles of 0.5±2.0 tonne, a final
drive gear reduction between 3.5:1 and 4.5:1 is
generally sufficient to meet all normal driving conditions, but with commercial vehicles carrying
considerably heavier payloads a demand for a
much larger final drive gear reduction of 4.5±9.0:1
is essential. This cannot be provided by a single
stage final drive crownwheel and pinion without
the crownwheel being abnormally large. Double
reduction axles partially fulfil the needs for heavy
goods vehicles operating under normal conditions
by providing two stages of gear reduction at the
In all double reduction final drive arrangements
the crownwheel and pinion are used to provide one
stage of speed step down. At the same time the
bevel gearing redirects the drive perpendicular to
the input propellor shaft so that the drive then
aligns with the axle half shafts.
Viscous coupling action (Figs 7.16 and 7.17) In the
situation when one wheel loses traction caused by
possibly loose soil, mud, ice or snow, the tyre±road
tractive effort reaction is lost. Because of this lost
traction there is nothing to prevent the planet
pinions revolving on their axes, rolling around the
opposite sun gear, which is connected to the road
wheel sustaining its traction, with the result that the
wheel which has lost its grip will just spin (race)
with no power being able to drive the good wheel
(Fig. 7.16). Subsequently, a speed difference
between the cage plates and half shaft hub plates
will be established and in proportion to this relative
speed, the two sets of coupling plates will shear the
silicon fluid and thereby generate a viscous drag
torque between adjacent plate faces (Fig. 7.17). As
a result of this viscous drag torque the half shaft
hub plates will proportionally resist the rate of fluid
shear and so partially lock the differential gear
mechanism. A degree of driving torque will be
transmitted to the good traction wheel. Fig. 7.17
also compares the viscous coupling differential
transmitted torque to the limited slip differential.
Here it can be seen that the limited slip differential
approximately provides a constant torque to the
good traction wheel at all relative speeds, whereas
the viscous coupling differential is dependent on
speed differences between both half shafts so that
7.4.2 Double reduction axles with first stage
reduction before the crownwheel and pinion
Double reduction with spur gears ahead of bevel gears
(Fig. 7.18) With a pair of helical gears providing
the first gear reduction before the crownwheel and
pinion, a high mounted and compact final drive
arrangement is obtained. This layout has the disadvantage of the final gear reduction and thus torque
multiplication is transmitted through the crownwheel and pinion bevel gears which therefore
absorbs more end thrust and is generally considered
to be less efficient in operation compared to helical
spur type gears. The first stage of a double reduction
axle is normally no more than 2:1 leaving the much
larger reduction for the output stage.
Fig. 7.18
Final drive spur double reduction ahead of bevel pinion
Fig. 7.19
Final drive spur double reduction between crownwheel and differential
Double reduction with bevel gears ahead of spur
gears (Fig. 7.19) A popular double reduction
arrangement has the input from the propellor
shaft going directly to the bevel pinion and crownwheel. The drive is redirected at right angles to that
of the input so making it flow parallel to the half
shafts, the first stage gear reduction being deter-
mined by the relative sizes (number of teeth) of the
bevel gears. A helical pinion gear mounted on the
same shaft as the crownwheel meshes with a helical
gear wheel bolted to the differential cage. The combination of these two gear sizes provides the second
stage gear reduction. Having the bevel gears ahead
of the helical gears ensures that only a proportion of
torque multiplication will be constrained by them,
while the helical gears will absorb the full torque
reaction of the final gear reduction.
shaft, as opposed to a single double reduction drive
if the reduction takes place before the differential.
7.4.3 Inboard and outboard double reduction
Where very heavy loads are to be carried by on-off
highway vehicles, the load imposed on the crownwheel and pinion and differential unit can be
reduced by locating a further gear reduction on
either side of the differential exit. If the second
gear reduction is arranged on both sides close to
the differential cage, it is referred to as an inboard
reduction. They can be situated at the wheel ends of
the half shafts, where they are known as outboard
second stage gear reduction. By having the reduction directly after the differential, the increased
torque multiplication will only be transmitted to
the half shafts leaving the crownwheel, pinion and
differential with a torque load capacity proportional to their gear ratio. The torque at this point
may be smaller than with the normal final drive
gear ratio since less gear reduction will be needed at
the crownwheel and pinion if a second reduction is
to be provided. Alternatively, if the second reduction is in the axle hub, less torque will be transmitted by the half shafts and final drive differential
and the dimensions of these components can be
kept to a minimum. Having either an inboard or
outboard second stage gear reduction enables
lighter crownwheel and pinion combinations and
differential assembly to be employed, but it does
mean there are two gear reductions for each half
Inboard epicyclic double reduction final drive axle
(Scammell) (Fig. 7.20) With this type of double
reduction axle, the first stage conforms to the conventional crownwheel and pinion whereas the second stage reduction occurs after passing through
the differential. The divided drive has a step down
gear reduction via twin epicyclic gear trains on
either side of the differential cage (Fig. 7.20).
Short shafts connect the differential bevel sun
gears to the pinion sun gear of the epicyclic gear
train. When drive is being transmitted, the rotation
of the sun gears rotates the planet pinions so that
they are forced to roll `walk' around the inside of
the reaction annulus gear attached firmly to the
axle casing. Support to the planet pinions and
their pins is given by the planet carrier which is
itself mounted on a ball race. Thus when the planet
pinions are made to rotate on their own axes they
also bodily rotate about the same axis of rotation
as the sun gear, but at a reduced speed, and in turn
convey power to the half shafts splined to the central hub portion of the planet carriers.
Inboard epicyclic differential and double reduction
axle (Kirkstall) (Fig. 7.21) This unique double
reduction axle has a worm and worm wheel first
stage gear reduction. The drive is transferred to an
epicyclic gear train which has the dual function of
providing the second stage gear reduction while at
Fig. 7.20 Inboard epicyclic double reduction final drive axle
Fig. 7.21
Inboard epicyclic double reduction axle
the same time performing as the final drive differential (Fig. 7.21).
Differential action of the epicyclic gears The
operation of the differential is quite straight forward if one imagines either the left or right hand
half shaft to slow down as in the case when they are
attached to the inner wheel of a cornering vehicle.
If when cornering the left hand half shaft slows
down, the planet carrier will correspondingly
reduce speed and force the planet pinions revolving
on their pins to spin at an increased speed. This
raises the speed of the sun gear which indirectly
drives, in this case, the outer right hand half shaft
at a slightly higher speed. Conversely, when cornering if the right hand half shaft should slow down, it
indirectly reduces the speed of the central pinion
and sun gear. Hence the planet pinions will not
revolve on their pins, but will increase their speed
at which they also roll round the outside of the sun
gear. Subsequently the planet pins will drive the
planet carrier and the left hand half shaft at an
increased speed.
Principle of operation Power is transmitted from
the propellor shaft to the worm and worm wheel
which produces a gear reduction and redirects the
drive at right angles and below the worm axis of
rotation (Fig. 7.21). The worm wheel is mounted
on the annulus carrier so that they both rotate as
one. Therefore the three evenly spaced
planet pinions meshing with both the annulus and
the sun gear are forced to revolve and move bodily
on their pins in a forward direction. Since the
sun gear is free to rotate (not held stationary) it
will revolve in a backward direction so that the
planet carrier and the attached left hand half
shaft will turn at a reduced speed relative to the
annulus gear.
Simultaneously, as the sun gear and shaft transfers motion to the right hand concentric gear train
central pinion, it passes to the three idler pinions,
compelling them to rotate on their fixed axes, and in
so doing drives round the annulus ring gear and with
it the right hand half shaft which is splined to it.
The right hand gear train with an outer internal
ring gear (annulus) does not form an epicyclic gear
train since the planet pins are fixed to the casing
and do not bodily revolve with their pins (attached
to a carrier) about some common centre of rotation. It is the purpose of the right hand gear train to
produce an additional gear reduction to equalize
the gear reduction caused by the planet carrier output on the left hand epicyclic gearing with the sun
gear output on the right hand side.
7.4.4 Outboard double reduction axles
Outboard epicyclic spur gear double reduction axle
(Fig. 7.22)
Description of construction (Fig. 7.22) A gear
reduction between the half shaft and road wheel
hub may be obtained through an epicyclic gear
train. A typical step down gear ratio would be 4:1.
The sun gear may be formed integrally with or it
may be splined to the half shaft (Fig. 7.22). It is
made to engage with three planet gears carried on
pins fixed to and rotating with the hub, thus driving
Fig. 7.22 Outboard epicyclic spur double reduction axle
the latter against the reaction of an outer annulus
splined to the stationary axle tube. The sun wheel
floats freely in a radial direction in mesh with the
planet pinions so that driving forces are distributed
equally on the three planet pinions and on their axes
of rotation. A half shaft and sun gear end float is
controlled and absorbed by a thrust pad mounted
on the outside end cover which can be initially
adjusted by altering the thickness of a shim pack.
axes of rotation, and the planet pins, are forced
to revolve about the sun gear axis. Since the
planet pins are mounted on the axle hub, which
is itself mounted via a fully floating taper bearing
arrangement on the axle tube, the whole hub
assembly will rotate at a much reduced speed
relative to the half shaft's input speed.
Outboard epicyclic bevel gear single and two speed
double reduction axle (Fig. 7.23) This type of outboard double reduction road wheel hub employs
bevel epicyclic gearing to provide an axle hub
reduction. To achieve this gear reduction there
are two bevel sun (side) gears. One is splined to
and mounted on the axle tube and is therefore
fixed. The other one is splined via the sliding sleeve
Description of operation (Fig. 7.22) In operation, power flows from the differential and half
shaft to the sun gear where its rotary motion is
distributed between the three planet pinions. The
forced rotation of these planet pinions compels
them to roll around the inside of the reaction
annulus ring gear (held stationary) so that their
dog clutch to the half shaft and so is permitted to
revolve (Fig. 7.23). Bridging both of these bevel sun
gears are two planetary bevel gears which are supported on a cross-pin mounted on the axle hub.
The planetary bevel gear double reduction axle
hub may be either two speed, as explained in the
following text, or a single speed arrangement in
which the half shaft is splined permanently and
directly to the outer sun gear.
Low ratio (Fig. 7.23) When low ratio is engaged,
the sleeve dog clutch is pushed inwards (to the
right) until the external teeth of the dog clutch are
moved out of engagement from the internal teeth of
the hub plate and into engagement with the internal
teeth of the outer bevel sun gear. The input drive is
now transmitted to the half shaft where it rotates
the outer bevel sun gear so that the bevel planet
gears are compelled to revolve on the cross-pin. In
doing so they are forced to roll around the fixed
inner bevel sun gear. Consequently, the cross-pin
which is attached to the axle hub is made to revolve
about the half shaft but at half its speed.
High ratio (Fig. 7.23) High ratio is selected and
engaged by twisting the speed selector eccentric so
that its offset peg pushes the sliding sleeve outwards (to the left) until the external teeth of the
dog clutch move out of engagement from the sun
gear and into engagement with the internal teeth
formed inside the axle hub end plate. Power is
transferred from the differential and half shaft
via the sleeve dog clutch directly to the axle hub
without producing any gear reduction.
Fig. 7.23
7.5 Two speed axles
The demands for a truck to operate under a varying
range of operating conditions means that the overall
transmission ratio spread needs to be extensive, which
is not possible with a single or double reduction final
Outboard epicyclic bevel gear two speed double reduction axle
Fig. 7.24 Two speed double reduction helical gear axle
drive. For example, with a single reduction final
drive the gear reduction can be so chosen as to
provide a high cruising speed on good roads with
a five speed gearbox. Conversely, if the truck is to
be used on hilly country or for off-road use then a
double reduction axle may provide the necessary
gear reduction.
Therefore, to enable the vehicle to operate effectively under both motorway cruising and town
stopping and accelerating conditions without overloading or overspeeding and without having to
have an eight, ten or twelve speed gearbox, a dual
purpose two speed gear reduction may be built into
the final drive axle.
Combining a high and low ratio in the same axle
doubles the number of gears available from the
standard gearbox. The low range of gears will then
provide the maximum pulling power for heavy duty
operations on rough roads, whereas the high range
of gears allows maximum speed when conditions are
favourable. From the wide range of gear ratios the
driver can choose the exact combination to suit any
conditions of load and road so that the engine will
always operate at peak efficiency and near to its
maximum torque speed band.
from the differential cage so that the dog teeth of
both the sleeve and the fixed ring teeth disengage. At
the same time the sun gear partially slides out of
mesh with the planet pinions and into engagement
with the outside pinion carrier internal dog teeth.
Subsequently, the sun gear is free to rotate. In addition the planet pinions and carrier are locked to the
sun gear, so that there can be no further relative
motion within the epicyclic gear train (i.e. annulus,
planet pinions, carrier, differential cage and sun
gear). In other words, the crownwheel and differential cage are compelled to revolve as one so that the
final drive second stage gear reduction is removed.
7.5.1 Two speed double reduction helical gear
axle (Rockwell-Standard) (Fig. 7.24)
This two speed double reduction helical gear axle
has a conventional crownwheel and bevel pinion
single speed first reduction with a second stage speed
reduction consisting of two pairs of adjacent pinion
and wheel helical cut gears. These pinions mounted
on the crownwheel support shaft act as intermediate
gears linking the crownwheel to the differential cage
final reduction wheel gears (Fig. 7.24).
Low ratio (Fig. 7.25) When the engagement sleeve
is moved inwards its dog clutch teeth engage with
the stationary ring teeth and the sun gear is pushed
fully in to mesh with the planet pinion low ratio that
has been selected. Under these conditions, the input
drive from the propellor shaft to the bevel pinion
still rotates the crownwheel but now the sun gear is
prevented from turning. Therefore the rotating
crownwheel with its internal annulus ring gear revolving about the fixed sun gear makes the planet
pinions rotate on their own axes (pins) and roll
around the outside of the held sun gear. As a result
of the planet pinions meshing with both the annulus
and sun gear, and the crownwheel and annulus
rotating while the sun gear is held stationary, the
planetary pinions are forced to revolve on their pins
which are mounted on one side of the differential
cage. Thus the cage acts as a planet pinion carrier
and in so doing is compelled to rotate at a slower
rate relative to the annulus gear speed. Subsequently, the slower rotation of the differential cage
relative to that of the crownwheel produces the
second stage gear reduction of the final drive.
Low ratio (Fig. 7.24) Low ratio is engaged when
the central sliding dog clutch splined to the crownwheel shaft slides over the selected (left hand) low
speed smaller pinion dog teeth. Power from the
propellor shaft now flows to the bevel pinion
where it is redirected at right angles to the crownwheel and shaft. From here it passes from the
locked pinion gear and crownwheel to the final
reduction wheel gear bolted to the differential
cage. The drive is then divided via the differential
cross-pin and planet pinions between both sun
gears where it is transmitted finally to the half
shafts and road wheels.
High ratio (Fig. 7.24) High ratio is engaged in a
similar way as for low ratio but the central sliding
dog clutch slides in the opposite direction (right
hand) over the larger pinion dog teeth. The slightly
larger pinion meshing with a correspondently
smaller differential wheel gear produces a more
direct second stage reduction and hence a higher
overall final drive axle gear ratio.
7.6 The third (central) differential
7.6.1 The necessity for a third differential
When four wheel drive cars or tandem drive axle
bogie trucks are to be utilized, provision must be
provided between drive axles to compensate for
any difference in the mean speeds of each drive
axle as opposed to speed differentiation between
pairs of axle road wheels.
Speed difference between driving axles are influenced by the following factors:
7.5.2 Two speed epicyclic gear train axle
(Eaton) (Fig. 7.25)
With this arrangement an epicyclic gear train is
incorporated between the crownwheel and differential cage (Fig. 7.25).
1 Speed variation between axles when a vehicle
moves on a curved track due to the slight differ-
High ratio (Fig. 7.25) When a high ratio is
required, the engagement sleeve is moved outwards
Fig. 7.25 Two speed epicyclic gear train axle
ence in rolling radius of both axles about some
instantaneous centre of rotation.
2 Small road surface irregularities, causing pairs of
driving wheels to locally roll into and over small
dips and humps so that each pair of wheels are
actually travelling at different speeds at any one
3 Tyres which have different amounts of wear or
different tread patterns and construction such as
cross-ply and radials, high and low profile etc. and
are mixed between axles so that their effective rolling radius of the wheel and tyre combination varies.
4 Uneven payload distribution will alter the effective rolling radius of a wheel and tyre so that
heavily laden axles will have smaller rolling radii
wheels and therefore complete more revolutions
over a given distance than lightly laden axles.
5 Unequal load distribution between axles when
accelerating and braking will produce a variation
of wheel effective rolling radius.
6 Loss of grip between pairs of road wheels produces momentary wheel spin and hence speed
differences between axles.
7.6.2 Benefits of a third differential (Fig. 7.26)
Operating a third differential between front and
rear wheel drive axles or rear tandem axles has
certain advantages:
1 The third differential equally divides driving torque and provides speed differentiation between
both final drive axles so that the relative torque
and speed per axle are better able to meet the
individual road wheel requirements, thereby
minimizing tyre distortion and scrub.
2 Transmission torsional wind-up between axles is
minimized (Fig. 7.26) since driving and reaction
torques within each axle are not opposing but
are permitted to equalize themselves through the
third differential.
3 Odd tyres with different diameters are interchangeable without transmission wind-up.
4 Tractive effect and tyre grip is shared between
four wheels so that wheel traction will be more
evenly distributed. Therefore the amount of tractive effect per wheel necessary to propel a vehicle
can be reduced.
5 Under slippery, snow or ice conditions, the third
differential can generally be locked-out so that if
one pair of wheels should lose traction, the other
pair of wheels are still able to transmit traction.
Power from the gearbox propellor shaft drives
the axle input shaft. Support for this shaft is
provided by a ball race mounted in the casing at
the flanged end and by a spigot bearing built into
the integral sun gear and output shaft at the other
end. Bevel planet pinions supported on the crosspin spider splined to the input shaft divide the drive
between both of the bevel sun gears. The left hand
sun gear is integral with the input helical gear and is
free to rotate relative to the input shaft which it is
mounted on, whereas the right hand bevel sun gear
is integral with the output shaft. This output shaft
is supported at the differential end by a large taper
roller bearing and by a much smaller parallel roller
bearing at the opposite flanged output end.
A tandem axle transmission arrangement is
shown in Fig. 7.28(a) where D1, D2 and D3 represent the first axle, second axle and inter axle differential respectively.
When power is supplied to the inter axle (forward rear axle) through the input shaft and to the
bevel planet pinion via the cross-pin spider, the
power flow is then divided between both sun
gears. The drive from the left hand sun gear then
passes to the input helical gear to the final drive
bevel pinion helical gear where it is redirected at
right angles by the crownwheel and pinion to the
axle differential and half shafts.
At the same time the power flowing to the right
hand sun gear goes directly to the output shaft flange
where it is then transmitted to the rear axle via a pair
of universal joints and a short propellor shaft.
7.6.3 Inter axle with third differential
Description of forward rear drive axle (Fig. 7.27)
A third differential is generally incorporated in
the forward rear axle of a tandem bogie axle drive
layout because in this position it can be conveniently arranged to extend the drive to the rear axle
(Fig. 7.27).
Third differential action (Fig. 7.27) When both
drive axles rotate at the same speed, the bevel planet
pinions bridging the opposing sun gears bodily
move around with the spider but do not revolve
on their own axes. If one axle should reduce its
speed relative to the other one, the planet pinions
will start to revolve on their cross-pins so that the
speed lost by one sun gear relative to the spider's
input speed will be gained by the other sun gear.
Therefore the third differential connecting the
two axles permits each axle mean speed to automatically adjust itself to suit the road operating
conditions without causing any torsional wind-up
between axle drives.
Third differential lock-out (Fig. 7.27) For providing maximum traction when road conditions are
unfavourable such as driving over soft, slippery or
steep ground, a differential lock-out clutch is
incorporated. When engaged this device couples
Fig. 7.26 Relationship of relative speeds of double drive
axles and the amount of transmission twist
Fig. 7.27 Final drive with third differential and lock and optional transfer gearing for front
the input shaft directly with the input helical gear
and left hand bevel sun gear so that the differential
planet pinions are prevented from equally dividing
the input torque between the two axles at the
expense of axle speed differentiation. Consequently, when the third differential is locked out
each axle is able to deliver independently to the
other axle tractive effect which is only limited by
the grip between the road wheels and the quality of
surface it is being driven over. It should be
observed that when the third differential lock-out
is engaged the vehicle should only be operated at
slow road speeds, otherwise excessive transmission
wind-up and tyre wear will result.
Front wheel drive transfer gear take-up (Fig. 7.28)
An additional optional feature is the transfer
gear take-up which is desirable for on-off highway applications where the ground can be rough
and uneven. With the front wheel drive lock
clutch engaged, 25% of the total input torque
from the gearbox will be transmitted to the
front steer drive axle, while the remainder of
the input torque 75% will be converted into
tractive effect by the tandem axles. Again it
should be pointed out that this mode of torque
delivery and distribution with the third differential locked-out must only be used at relatively
low speeds.
Fig. 7.28 (a and b)
Tandem drive axle layout
7.6.4 Worm and worm wheel inter axle with third
differential (Fig. 7.29)
Where large final drive gear reductions are required
which may range from 5:1 to 9:1, either a double
reduction axle must be used or alternatively a
worm and worm wheel can provide a similar step
down reduction. When compared with the conventional crownwheel and pinion final drive gear
reduction the worm and worm wheel mechanical
efficiency is lower but with the double reduction
axle the worm and worm wheel efficiency is very
similar to the latter.
Worm and worm wheel axles usually have the
worm underslung when used on cars so that a very
low floor pan can be used. For heavy trucks the
worm is arranged to be overslung, enabling a large
ground to axle clearance to be achieved.
Fig. 7.29
When tandem axles are used, an inter axle third
differential is necessary to prevent transmission
wind-up. This unit is normally built onto the axle
casing as an extension of the forward axle's worm
(Fig. 7.29).
The worm is manufactured with a hollow axis
and is mounted between a double taper bearing to
absorb end thrust in both directions at one end and
a parallel roller bearing at the other end which just
sustains radial loads. The left hand sun gear is
attached on splines to the worm but the right
hand sun gear and output shaft are mounted on a
pair of roller and ball bearings.
Power flow from the gearbox and propellor shaft
is provided by the input spigot shaft passing through
the hollow worm and coming out in the centre of the
bevel gear cluster where it supports the internally
Worm and worm wheel inter axle differential
splined cross-pin spider and their corresponding
planet pinions. Power is then split between the
front axle (left hand) sun gear and worm and the
rear axle (right hand) sun gear and output shaft,
thus transmitting drive to the second axle.
Consequently if the two axle speeds should vary, as
for example when cornering, the planet pinions will
revolve on their axes so that the sun gears are able to
rotate at speeds slightly above and below that of the
input shaft and spider, but at the same time still
equally divide the torque between both axles.
Fig. 7.28(b) shows the general layout of a tandem axle worm and worm wheel drive where D1,
D2 and D3 represent the first axle, second axle and
inter axle differentials respectively.
four wheel drive, the engine's power will be divided
by four so that each wheel will only have to cope
with a quarter of the power available, so that each
individual wheel will be far below the point of
transmitting its limiting traction force before
breakaway (skid) is likely to occur.
During cornering, body roll will cause a certain
amount of weight transfer from the inner wheels to
the outer ones. Instead of most of the tractive effort
being concentrated on just one driving wheel, both
front and rear outer wheels will share the vertical
load and driving thrust in proportion to the weight
distribution between front and rear axles. Thus a
four wheel drive (4WD) when compared to a two
wheel drive (2WD) vehicle has a much greater margin of safety before tyre to ground traction is lost.
Transmission losses overall for front wheel drive
(FWD) are in the order of 10%, whereas rear wheel
drive (RWD) will vary from 10% in direct fourth
gear to 13% in 1st, 2nd, 3rd, and 5th indirect gears.
In general, overall transmission losses with four
wheel drive (4WD) will depend upon the transmission configuration and may range from 13% to 15%.
7.7 Four wheel drive arrangements
7.7.1 Comparison of two and four wheel drives
The total force that a tyre can transmit to the road
surface resulting from tractive force and cornering
for straight and curved track driving is limited by
the adhesive grip available per wheel.
When employing two wheel drive, the power
thrust at the wheels will be shared between two
wheels only and so may exceed the limiting traction
for the tyre and condition of the road surface. With
Fig. 7.30 (a and b)
7.7.2 Understeer and oversteer characteristics
(Figs 7.30 and 7.31)
In general, tractive or braking effort will reduce the
cornering force (lateral force) that can be generated
The influence of front and rear tyre slip angles on steering characteristics
for a given slip angle by the tyre. In other words
the presence of tractive or braking effort requires
larger slip angles to be produced for the same cornering force; it reduces the cornering stiffness of the
tyres. The ratio of the slip angle generated at the
front and rear wheels largely determines the vehicle's tendency to oversteer or understeer (Fig. 7.30).
The ratio of the front to rear slip angles when
greater than unity produces understeer,
Experimental results (Fig. 7.31) have shown that
rear wheel drive (RWD) inherently tends to give
oversteering by a small slightly increasing amount,
but front and four wheel drives tend to understeer
by amounts which increase progressively with
speed, this tendency being slightly greater for the
front wheel drive (FWD) than for the four wheel
drive (4WD).
If the slip angle of the rear tyres is greater than the
front tyres the vehicle will tend to oversteer, but if
the front tyres generate a greater slip angle than the
rear tyres the vehicle will have a bias to understeer.
Armed with the previous knowledge of tyre
behaviour when tractive effort is present during
cornering, it can readily be seen that with a rear
wheel drive (RWD) vehicle the tractive effort
applied to propel the vehicle round a bend
increases the slip angle of the rear tyres, thus introducing an oversteer effect. Conversely with a front
wheel drive (FWD) vehicle, the tractive effort input
during a turn increases the slip angle of the front
tyres so producing an understeering effect.
7.7.3 Power loss (Figs 7.32 and 7.33)
Tyre losses become greater with increasing tractive
force caused partially by tyre to surface slippage.
This means that if the total propulsion power is
shared out with more driving wheels less tractive
force will be generated per wheel and therefore less
overall power will be consumed. The tractive force
per wheel generated for a four wheel drive compared to a two wheel drive vehicle will only be half
as great for each wheel, so that the overall tyre to
road slippage will be far less. It has been found that
the power consumed (Fig. 7.32) is least for the front
wheel drive and greatest for the rear wheel drive,
while the four wheel drive loss is somewhere in
between the other two extremes.
The general relationship between the limiting tractive power delivered per wheel with either propulsion
or retardation and the power loss at the wheels is
shown to be a rapidly increasing loss as the power
delivered to each wheel approaches the limiting
adhesion condition of the road surface. Thus with
a dry road the power loss is relatively small with
Fig. 7.31 Comparison of the over- and understeer
tendency of RWD, FWD and 4WD cars on a curved track
Fig. 7.32 Comparison of the power required to drive
RWD, FWD and 4WD cars on a curved track at various
i:e: Ratio
< 1:
When the ratios of the front to rear slip angles
are less than unity oversteer is produced,
i:e: Ratio
> 1:
7.7.4 Maximum speed (Fig. 7.34)
If friction between the tyre and road sets the limit
to the maximum stable speed of a car on a bend,
then the increasing centrifugal force will raise the
cornering force (lateral force) and reduce the effective tractive effort which can be applied with rising
speed (Fig. 7.34). The maximum stable speed a
vehicle is capable of on a curved track is highest
with four wheel drive followed in order by the front
wheel drive and rear wheel drive.
7.7.5 Permanent four wheel drive transfer box
(Land and Range Rover) (Fig. 7.35)
Transfer gearboxes are used to transmit power
from the gearbox via a step down gear train to a
central differential, where it is equally divided
between the front and rear output shafts (Fig.
7.35). Power then passes through the front and
rear propellor shafts to their respective axles and
road wheels. Both front and rear coaxial output
shafts are offset from the gearbox input to output
shafts centres by 230 mm.
The transfer box has a low ratio of 3.32:1 which
has been found to suit all vehicle applications. The
high ratio uses alternative 1.003:1 and 1.667:1
ratios to match the Range Rover and Land Rover
requirements respectively. This two stage reduction
unit incorporates a three shaft six gear layout inside
an aluminium housing. The first stage reduction
from the input shaft to the central intermediate
gear provides a 1.577:1 step down. The two outer
intermediate cluster gears mesh with low and high
range output gears mounted on an extension of the
differential cage.
Drive is engaged by sliding an internally splined
sleeve to the left or right over dog teeth formed on
both low and high range output gears respectively.
Power is transferred from either the low or high
range gears to the differential cage and the bevel
planet pinions then divide the torque between the
front and rear bevel sun gears and their respective
output shafts. Any variation in relative speeds
between front and rear axles is automatically compensated by permitting the planet pinions to revolve
on their pins so that speed lost by one output shaft
will be equal to that gained by the other output shaft
relative to the differential cage input speed.
A differential lock-out dog clutch is provided
which, when engaged, locks the differential cage
directly to the front output shaft so that the bevel
gears are unable to revolve within the differential
cage. Consequently the front and rear output shafts
are compelled to revolve under these conditions at
the same speed.
Fig. 7.33 Relationship of tractive power and power loss
for different road conditions
Fig. 7.34 Comparison of the adhesive traction available
to Drive, RWD, FWD and 4WD cars on a curved track at
various speeds
increasing tractive power because the tyre grip on the
road is nowhere near its limiting value. With semiwet or wet road surface conditions the tyre's ability
to maintain full grip deteriorates and therefore the
power loss increases at a very fast rate (Fig. 7.33).
Fig. 7.35
Permanent 4WD Land and Range Rover type of transfer box
A power take-off coupling point can be taken
from the rear of the integral input gear and shaft.
There is also a central drum parking brake which
locks both front and rear axles when applied.
It is interesting that the low range provides an
overall ratio down to 40:1, which means that the
gearbox, transfer box and crownwheel and pinion
combined produce a gear reduction for gradient
ability up to 45 .
referred to as the central differential). This shaft is
splined to the planet pinion carrier (Fig. 7.36). The
four planet pinions are supported on the carrier
mesh on the outside with the internal teeth of the
annulus ring gear, while on the inside the teeth
of the planet pinions mesh with the sun gear teeth.
A hollow shaft supports the sun gear. This gear
transfers power to the front wheels via the offset
input and output sprocket wheel chain drive.
The power path is then completed by way of a propellor shaft and two universal joints to the front
crownwheel and pinion. Mounted on a partially
tubular shaped carrier is the annulus ring gear
which transfers power from the planet pinions
directly to the output shaft of the transfer box
unit. Here the power is conveyed to the rear axle
7.7.6 Third (central) differential with viscous
Description of third differential and viscous coupling
(Fig. 7.36) The gearbox mainshaft provides the
input of power to the third differential (sometimes
Fig. 7.36 Third differential with viscous coupling
by a conventional propellor shaft and coupled at
either end by a pair of universal joints.
cular epicyclic gear train. This torque distribution
is achieved by the ratio of the radii of the meshing
teeth pitch point of both planet to annulus gear and
planet to sun gear from the centre of shaft rotation.
Since the distance from the planet to annulus teeth
pitch point is twice that of the planet to sun teeth
pitch point, the leverage applied to the rear wheel
drive will be double that going to the front wheel
Speed balance of third differential assembly with
common front and rear wheel speed (Fig. 7.36)
Power from the gearbox is split between the sun
gear, taking the drive to the front final drive. The
annulus gear conveys power to the rear axle. When
the vehicle is moving in the straight ahead direction
and all wheels are rotating at the same speed, the
whole third differential assembly (the gearbox
mainshaft attached to the planet carrier), planet
pinions, sun gear and annulus ring gear will all
revolve at the same speed.
Viscous coupling action (Fig. 7.36) Built in with
the epicyclic differential is a viscous coupling resembling a multiplate clutch. It comprises two sets of
mild steel disc plates; one set of plates are splined to
the hollow sun gear shafts while the other plates are
splined to a drum which forms an extension to the
annulus ring gear. The sun gear plates are disfigured
by circular holes and the annulus drum plates have
radial slots. The space between adjacent plates is
filled with a silicon fluid. When the front and rear
road wheels are moving at slightly different
Torque distribution with common front and rear
wheel speed (Fig. 7.36) While rear and front propellor shafts turn at the same speed, the torque split
will be 66% to the rear and 34% to the front,
determined by the 2:1 leverage ratio of this parti258
speeds, the sun and annulus gears are permitted to
revolve at speeds relative to the input planet carrier
speed and yet still transmit power without causing
any transmission wind-up.
Conversely, if the front or rear road wheels
should lose traction and spin, a relatively large
speed difference will be established between the sets
of plates attached to the front drive (sun gear) and
those fixed to the rear drive (annulus gear). Immediately the fluid film between pairs of adjacent
plate faces shears, a viscous resisting torque is generated which increases with the relative plate speed.
This opposing torque between plates produces a
semi-lock-up reaction effect so that tractive effort
will still be maintained by the good traction road
wheel tyres. A speed difference will always exist
between both sets of plates when slip occurs
between the road wheels either at the front or
rear. It is this speed variation that is essential to
establish the fluid reaction torque between plates,
and thus prevent the two sets of plates and gears
(sun and annulus) from racing around relative to
each other. Therefore power will be delivered to the
axle and road wheels retaining traction even when
the other axle wheels lose their road adhesion.
pinions between the left and right hand bevel sun
gears. Half the power flows to the front crownwheel
via the long pinion shaft passing through the centre
of the secondary hollow output shaft while the
other half flows from the right hand sun gear to
the rear axle via the universal joints and propellor
When the vehicle is moving forward in a straight
line, both the front and rear axles rotate at one
common speed so that the axle pinions will revolve
at the same speed as the central differential cage.
Therefore the bevel gears will rotate bodily with the
cage but cannot revolve relative to each other.
Steering the vehicle or moving onto a bend or
curved track will immediately produce unequal
turning radii for both front and rear axles which
meet at some common centre (instantaneous
centre). Both axles will be compelled to rotate at
slightly different speeds. Due to this speed variation between front and rear axles, one of the central differential sun gears will tend to rotate faster
than its cage while the other one will move
correspondently slower than its cage. As a result,
the sun gears will force the planet pinions to
revolve on their pins and at the same time revolve
bodily with the cage. This speed difference on both
sides of the differential is automatically absorbed
by the revolving planet pinions now being permitted to move relative to the sun gears by rolling
on their toothed faces. By these means, the bevel
gears enable both axles to rotate at speeds
demanded by their instantaneous rolling radii at
any one moment without causing torsional windup. If travelling over very rough, soft, wet or steep
terrain, better traction may be achieved with the
central differential locked-out.
7.7.7 Longitudinal mounted engine with integral
front final drive four wheel drive layout (Fig. 7.37)
The power flow is transmitted via the engine to the
five speed gearbox input primary shaft. It then
transfers to the output secondary hollow shaft by
way of pairs of gears, each pair combination having
different number of teeth to provide the necessary
range of gear ratios (Fig. 7.37). The hollow secondary shaft extends rearwards to the central differential cage. Power is then divided by the planet
Fig. 7.37
Longitudinally mounted engine with integral front final drive four wheel drive system
Fig. 7.38 Longitudinally mounted engine with independent front final drive four wheel drive system
7.7.8 Longitudinal mounted engine with
independent front axle four wheel drive layout
(Fig. 7.38)
then passes to the spider cross-pins which support
the bevel planet pinions. Here the torque is distributed equally between the front and rear bevel sun
gears, these being connected indirectly through
universal joints and propellor shafts to their respective axles. When the vehicle is moving along a
straight path, the planet pinions do not rotate but
just revolve bodily with the cage assembly.
Immediately the vehicle is manoeuvred or is negotiating a bend, the planet pinions commence rotating on their own pins and thereby absorb speed
differences between the two axles by permitting
them not only to turn with the cage but also to roll
round the bevel sun gear teeth at the same differential. However, they are linked together by bevel gearing which permits them independently to vary their
speeds without torsional wind-up and tyre scuffing.
Epicyclic gear central differential (Fig. 7.38) A
popular four wheel drive arrangement for a front
longitudinally mounted engine has a transfer box
behind its five speed gearbox. This incorporates a
viscous coupling and an epicyclic gear train to split
the drive torque, 34% to the front and 66% to the
rear (Fig. 7.38). A chain drives a forward facing
drive shaft which provides power to the front differential mounted beside the engine sump. The
input drive from the gearbox mainshaft directly
drives the planet carrier and pinions. Power is
diverted to the front axle through the sun gear
and then flows to the hollow output shaft to the
chain sprockets. Output to the rear wheels is taken
from the annulus ring gear and carrier which transmits power directly to the rear axle. To minimize
wheel spin between the rear road wheels a
combined differential and viscous coupling is
incorporated in the rear axle housing.
7.7.9 Transversely mounted engine with four
wheel drive layout (Fig. 7.39)
One method of providing four wheel drive to a
front transversely mounted engine is shown in
Fig. 7.39. A 50/50 torque split is provided by an
epicyclic twin planet pinion gear train using the
annulus ring gear as the input. The drive to the
front axle is taken from the central sun gears
which is attached to the front differential cage,
while the rear axle is driven by the twin planet
pinions and the crownwheel, which forms the
planet carrier. Twin planet pinions are used to
make the sun gear rotate in the same direction of
rotation as that of the annulus gear. A viscous coupling is incorporated in the front axle differential
to provide a measure of wheel spin control.
Power from the gearbox is transferred to the
annulus ring gear by a pinion and wheel, the ring
Bevel gear central differential (Fig. 7.38) In some
cases vehicles may have a weight distribution or a
cross-counting application which may find 50/50
torque split between front and rear wheel drives
more suitable than the 34/66 front to rear torque
split. To meet these requirements a conventional
central (third) bevel gear differential may be preferred, see insert in Fig. 7.38. Again a transfer box
is used behind the gearbox to house the offset
central differential and transfer gears. The transfer
gear train transmits the drive from the gearbox
mainshaft to the central differential cage. Power
Fig. 7.39
Transversely mounted engine four wheel drive system
Fig. 7.40
Rear mounted engine four wheel drive system
shaft to the crownwheel and pinion through 90 to
the wheel hubs. Similarly power to the front axle is
taken from the front end of the gearbox secondary
output shaft to the front axle assembly comprised
of the crownwheel and pinion differential and
viscous coupling.
The viscous relative speed-sensitive fluid coupling
has two independent perforated and slotted sets of
steel discs. One set is attached via a splined shaft to a
stub shaft driven by the propellor shaft from the
gearbox, the other to the bevel pinion shank of the
front final drive. The construction of the multi-interleaf discs is similar to a multiplate clutch but there is
no engagement or release mechanism. Discs always
remain equidistant from each other and power
transmission is only by the silicon fluid which stiffens and produces a very positive fluid drag between
plates. The sensitivity and effectiveness of the transference of torque is dependent upon the diameter
and number of plates (in this case 59 plates), size of
gear having external teeth to mesh with the input
pinion from the gearbox and internal teeth to drive
the twin planet gears. Rotation of the annulus ring
gear drives the outer and inner planet pinions and
subsequently rotates the planet carrier (crownwheel in this case). The front crownwheel and
pinion redirect the drive at right angles to impart
motion to the propellor shaft. Simultaneously
the inner planet pinion meshes with the central
sun gear so that it also relays motion to the front
differential cage.
7.7.10 Rear mounted engine four wheel drive
layout (Fig. 7.40)
This arrangement has an integral rear engine and
axle with the horizontal opposed four cylinder
engine mounted longitudinally to the rear of the
drive shafts and with the gearbox forward of the
drive shafts (Fig. 7.40). Power to the rear axle is
taken directly from the gearbox secondary output
perforated holes and slots, surface roughness of the
plates as well as temperature and generated pressure
of fluid.
The drive to the front axle passes through the
viscous coupling so that when both front and rear
axle speeds are similar no power is transmitted to
the front axle. Inevitably, in practice small differences in wheel speeds between front and back due
to variations of effective wheel radii (caused by
uneven load distribution, different tyre profiles,
wear and cornering speeds) will provide a small
degree of continuous drive to the front axle. The
degree of speed sensitivity is such that it takes only
one eighth of a turn in speed rotational difference
between each end of the coupling for the fluid to
commence to stiffen. Only when there is a loss of
grip through the rear wheels so that they begin to
slip does the mid-viscous coupling tend to lock-up
to provide positive additional drive to the front
wheels. A mechanical differential lock can be incorporated in the front or rear axles for travelling over
really rough ground.
7.8.1 Description of multiplate clutch mechanism
(Fig. 7.41)
To overcome this deficiency a multiplate wet clutch
is incorporated to one side of the differential cage,
see Fig. 7.41. One set of the clutch plates have
internal spin teeth which mesh with splines formed
externally on an extended sun (side) gear, whereas
the other set of inter disc plates have external spline
teeth which mesh with internal splines formed
inside the differential cage. Thus, when there are
signs of any of the wheels losing their grip the
clutch plates are automatically clamped partially
or fully together. The consequence of this is to
partially or fully lock both left and right hand
side output drive shafts together so that the loss
of drive of one drive wheel will not affect the
effectiveness of the other wheel. To activate the
engagement and release of the multiplate clutch,
a servo-piston mounted on the right hand side
bearing support flange is used: the piston is stepped
and has internal seals for each step so that hydraulic fluid is trapped between the internal and external stepped piston and bearing support flange
7.8 Electro-hydraulic limited slip differential
A final drive differential allows the driving wheels
on each side of a vehicle to revolve at their true
rolling speed without wheel slip when travelling
along a straight uneven surface, a winding road or
negotiating a sharp corner. If the surface should be
soft, wet, muddy, or slippery for any other reason,
then one or the other or even both drive wheels
may lose their tyre to ground traction, the vehicle
will then rely on its momentum to ride over these
patchy slippery low traction surfaces. However if
the vehicle is travelling very slowly and the ground
surface is particularly uneven, soft or slippery, then
loss of traction of one of the wheels could easily be
sufficient to cause the wheel to spin and therefore
to transmit no drive. Unfortunately, due to the
inherent design of the bevelled gear differential
the traction delivered at the good gripping wheel
will be no more than that of the tyre that has lost its
grip. A conventional bevelled gear differential
requires that each sun (side) gear provides equal
driving torque to each wheel and at the same time
opposite sun (side) gears provide reaction torque
equal to the driving torque of the opposite wheel.
Therefore as soon as one wheel' loses ground traction its opposite wheel, even though it may have a
firm tyre to ground contact, is only able to produce
the same amount of effective traction as the wheel
with limited ground grip.
7.8.2 Operating conditions
Normal differential action (Fig. 7.41) With good
road wheel grip the multiplate clutch is disengaged
by closing the delivery solenoid valve and releasing
fluid to the reservoir tank via the open return solenoid valve. Under these conditions when there is a
difference in speed between the inner and outer
road wheels, the bevel-planet pinions are free to
revolve on their axes and hence permit each sun
(side) gear to rotate at the same speed as its adjacent road wheel thereby eliminating any final drive
transmission windup and tyre scrub.
One wheel on the threshold of spinning (Fig. 7.41)
If one wheel commences to spin due to loss of
traction the wheel speed sensor instantly detects
the wheel's acceleration and signals the ECU; the
computer then processes this information and taking into account that a small amount of slip
improves the tyre to ground traction will then energize and de-energize the delivery and return solenoid valves respectively. Fluid will now be pumped
from the power assistant steering systems pump to
the servo-piston, the pressure build up against the
piston will engage and clamp the multiplate clutch
via the thrust-pins and plate so that the differential
Left hand
pin Thrust
Right hand
drive shaft
Right hand
bearing and
piston support
Right hand
output drive
shaft / coupling
taper bearing
taper bearing
Spacer sleeve
taper bearing
Oil seal
control valves
Flanged universal
joint coupling
Fig. 7.41
Electro-hydraulic limited-slip (differential in locked position)
cage now is able to provide the reaction torque
for the other wheel still delivering traction to the
The ECU is able to take into account the speed
of the vehicle and if the vehicle is turning gently or
sharply which is monitored by the individual brake
speed sensors and the steering wheel accelerator
sensor. These two sensors therefore indirectly control the degree of lock-up which would be severe
when pulling away from a standstill but would ease
up with increased vehicle speed and when negotiating a bend.
snow covered roads the coefficient of friction
would be as low as 0.2.
Wheel slip for accelerating and decelerating is
usually measured as the slip ratio or the percentage
of slip and may be defined as follows:
7.9 Tyre grip when braking and accelerating with
good and poor road surfaces (Fig. 7.42)
The function of the tyre and tread is to transfer the
accelerating and decelerating forces from the
wheels to the road. The optimum tyre grip is
achieved when there is about 15±25% slip between
the tyre tread and road under both accelerating and
decelerating driving conditions, see Fig. 7.42.
Tyre grip is a measure of the coefficient of friction () generated between the tyre and road surface at any instant, this may be defined as ˆ F/W
where F is the frictional force and W is the perpendicular force between the tyre and road. If the
frictional and perpendicular forces are equal
( ˆ 1:0) the tyre tread is producing its maximum
grip, whereas if ˆ 0 then the grip between the
tyre and road is zero, that is, it is frictionless.
Typical tyre to road coefficient for a good tarmac
dry and wet surface would be 1.0 and 0.7 respectively, conversely for poor surfaces such as soft
accelerating slip ratio = road speed/tyre
decelerating (braking) = tyre speed/road
slip ratio
where the tyre speed is the linear periphery speed.
Note the percentage of slip may be taken as the
slip ratio 100. There is no slippage or very little
that takes place between the tyre and road surface
when a vehicle is driven at a constant speed along a
dry road, under these conditions the slip ratio is
zero (slip ratio ˆ 0). Conversely heavy acceleration
or braking may make the wheels spin or lock
respectively thus causing the slip ratio to approach
unity (slip ratio ˆ 1:0).
If the intensity of acceleration or deceleration is
increased the slip ratio tends to increase since during
acceleration the wheels tend to slip and in the
Optimum slip ratio for ABS
Optimum slip ratio for TCS
Coefficient of friction (µ)
c re
Dr y
g itu
(l on
ed (lo
ered (la
Snow cov
(l a
di n
d in
ed (lo
Slip ratio (braking)
Slip ratio (accelerating)
Fig. 7.42 Tyre grip as a function of slip ratio for various driving conditions
extreme spin, whereas during deceleration (braking)
the wheels tend to move slower than the vehicle
speed and under very heavy braking will lock, that
is stop rotating and just slide along the surface.
When considering the relationship between tyre slip
and grip it should be observed that the tyre grip
measurements are in two forms, longitudinal (lengthways) forces and lateral (sideways) forces, see Fig.
7.42. In both acceleration and deceleration in the
longitudinal direction mode a tyre tends to produce
its maximum grip (high ) with a slip ratio of about
0.2 and as the slip ratio decreases towards zero, the
tyre grip falls sharply; however, if the slip ratio
increases beyond the optimum slip ratio of 0.2 the
tyre grip will tend to decrease but at a much slower
rate. With lateral direction grip in terms of sideways
force coefficient of friction, the value of (grip) is
much lower than for the forward rolling frictional
grip and the maximum grip (high ) is now produced
with zero tyre slip. Traction control systems (TCS)
respond to wheel acceleration caused by a wheel
spinning as its tyre loses its grip with the road surface, as opposed to antilock braking systems (ABS)
which respond to wheel deceleration caused by a
wheel braking and preventing the wheel from turning and in the limit making it completely lock.
individual wheel speed sensors, and as soon as one
of the driving wheels tends to accelerate (spin due to
loss of tyre to ground traction), the sensor's generated voltage change is processed by the ECU computer, and subsequently current is directed to the
relevant traction solenoid valve unit so that hydraulic brake pressure is transmitted to the brake of the
wheel about to lose its traction. As soon as the
braked wheel's speed has been reduced to a desirable
level, then the ECU signals the traction solenoid
valve unit to release the relevant wheel brake.
7.10.1 Description of system (Fig. 7.43)
This traction control system consists of: an electric
motor driven hydraulic pump which is able to generate brake pressure independently to the foot brake
master cylinder and a pressure storage accumulator;
a traction boost unit which comprises a cylinder
housing, piston and poppet valve, the purpose of
which is to relay hydraulic pressure to the appropriate wheel brake caliper and at the same time maintain the traction boost unit circuit fluid separate
from the foot brake master cylinder fluid system; a
pair of traction solenoid valve units each having an
outlet and return valve regulates the cut-in and -out
of the traction control; an electronic control unit
(ECU) is provided and individual wheel speed sensors which monitor the acceleration of both driven
and non-driving wheels. Should the speed of either
of the driven wheels exceed the mean wheel speed of
the non-driven wheels by more than about 20%,
then the ECU will automatically apply the appropriate wheel brake via the traction solenoid valves
and traction boost device.
7.10 Traction control system
With a conventional final drive differential the
torque output from each driving wheel is always
equal. Thus if one wheel is driven over a slippery
patch, that wheel will tend to spin and its adjacent
sun (side) gear will not now be able to provide the
reaction torque for the other (opposite) sun (side)
gear and driving road wheel. Accordingly, the output torque on the other wheel which still has a good
tyre to surface grip will be no more than that of the
slipping wheel and it doesn't matter how much the
driver accelerates to attempt to regain traction,
there still will be insufficient reaction torque on
the spinning side of the differential for the good
wheel to propel the vehicle forward.
One method which may be used to overcome this
loss of traction when one wheel loses its road grip, is
to simply apply the wheel brake of the wheel showing signs of spinning so that a positive reaction
torque is provided in the differential; this counteracts the delivery to the good wheel of the half share
of the driving torque being supplied by the combined engine and transmission system in terms of
tractive effort between the tyre and ground.
To achieve this traction control an electronic control unit (ECU) is used which receives signals from
7.10.2 Operating conditions
Foot brake applied (Fig. 7.43(a)) With the foot
pedal released brake fluid is able to flow freely
between the master cylinder and both brake calipers via the open traction boost unit. The boost
piston will be in its outer-most position thereby
holding the poppet valve in its fully open position.
When the foot pedal is pushed down, brake fluid
pressure will be transmitted though the fluid via the
traction boost poppet valve to both of the wheel
calipers thus causing the brakes to be applied.
Traction control system applied (off-side slipping
wheel braked) (Fig. 7.43(b)) When one of the
driving wheels begins to spin (off-side wheel in
this example) the adjacent speed sensor voltage
change signals the ECU, immediately it computes
Outlet Poppet
valve valve
(closed) (open)
Traction control system
Electronic control unit
(a) Conventional foot brakes applied
Fig. 7.43 (a and b)
Traction control system
Traction boost
Off-side wheel
Wheel tending
to spin
Near-side wheel
(b) Off-side spinning wheel braked
Fig. 7.43
Electric motor
control valve
Plunger Eccentric
pump cam
Fig. 7.44 Combined antilock brake system/traction control system (brakes applied)
and relays current to the relevant traction solenoid
valve unit to energize both valves, this closes the
return valve and opens the outlet valve. Fluid pressure from the accumulator now flows through the
open outlet valve and passes into the traction boost
cylinder, the upward movement of the piston will
instantly snap closed the poppet valve thereby trapping fluid between the upper side of the cylinder
and piston chamber and the off-side caliper. The
fluid pressure build up underneath the piston will
pressurize the fluid above the piston so that the
pressure increase is able to clamp the caliper pads
against the brake disc.
As the wheel spin speed reduces to a predetermined
value the monitoring speed sensor signals the ECU
to release the wheel brake, immediately the solenoid outlet and return valves will be de-energized,
thus causing them to close and open respectively.
Fluid pressure previously reaching the boost piston
will now be blocked and the fluid underneath the
piston will be able to return to the reservoir tank.
The same cycle of events will take place for the
near-side wheel if it happens to move over a slippery surface.
7.10.3 Combined ABS/TCS arrangement
(Fig. 7.44)
Normally a traction control system (TCS) is incorporated with the antilock braking system (ABS) so
that it can share common components such as the
electric motor, pump, accumulator, wheel brake
sensors and high pressure piping. As can be seen
in Fig. 7.44 a conventional ABS system described
in section 11.7.2 has been added to. This illustration shows when the brakes are applied fluid pressure is transmitted indirectly through the antilock
solenoid control valve to the front brake calipers;
however, assuming a rear wheel drive, fluid pressure also is transmitted to the rear brakes via the
antilock solenoid control valve and then through
the traction boost unit to the wheel brake calipers
thus applying the brakes.
towards slip and spin the respective traction solenoid control valve closes its return valve and opens
its outlet valve; fluid pressure from the pump now
provides the corresponding boost piston with an
outward thrust thereby causing the poppet valve
to close (note Fig. 7.44 only shows the system in
the foot brake applied position). Further fluid
pressure acting on the head of the piston now
raises the pressure of the trapped fluid in the
pipe line between the boost piston and the wheel
caliper. Accordingly the relevant drive wheel is
braked to a level that transmits a reaction torque
to the opposite driving wheel which still retains
One limitation of a brake type traction control is
that a continuous application of the TCS when
driving over a prolonged slippery terrain will
cause the brake pads and disc to become excessively hot; it thus may lead to brake fade and a
very high wear rate of the pads and disc.
ABS operating (Fig. 7.44) When the wheel brake
speed sensor signals that a particular wheel is tending towards wheel lock, the appropriate antilock
solenoid control valve will be energized so that
fluid pressure to that individual wheel brake is
blocked and the entrapped fluid pressure is released
to the pressure reducing accumulator (note Fig.
7.44 only shows the system in the foot brake
applied position).
TCS operating (Fig. 7.44) If one of the wheel
speed sensors signals that a wheel is moving
8.1 Tractive and braking properties of tyres
8.1.1 Tyre grip
Tyres are made to grip the road surface when the
vehicle is being steered, accelerated, braked and/or
negotiating a corner and so the ability to control
the tyre to ground interaction is of fundamental
importance. Road grip or friction is a property
which resists the sliding of the tyre over the road
surface due to a retardant force generated at the
tyre to ground contact area. The grip of different tyres
sliding over various road surface finishes may be
compared by determining the coefficient of friction
for each pair of rubbing surfaces.
The coefficient of friction may be defined as the
ratio of the sliding force necessary to steadily move
a solid body over a horizontal surface to the normal reaction supporting the weight of the body on
the surface (Fig. 8.1).
Fig. 8.2
Frictional force
Normal reaction
Variation of friction with relative movement
between the flexible tread elements and road.
Therefore when dealing with tyres it is usual to
refer to the coefficient of adhesive friction. The maximum coefficient of adhesive friction created
between a sliding tread block and a solid surface
occurs under conditions of slow movement or creep
(Fig. 8.2). This critical stage is known as the peak
coefficient, p , and if the relative movement of the
rubber on the surface is increased beyond this point
the friction coefficient falls. It continues to fall until
bodily sliding occurs, this stage being known as the
sliding coefficient s . Sliding friction characteristics
are consistent with the behaviour of rolling tyres.
A modern compound rubber tyre will develop
a higher coefficient of friction than natural rubber.
In both cases their values decrease as the road surface changes from dry to a wet condition. The rate
of fall in the coefficient of friction is far greater with
a worn tyre tread as opposed to a new tyre as the
degree of road surface wetness increases (Fig. 8.3).
It has been found that the frictional grip of a bald
tyre tread on a rough dry road surface is as good or
even better than that achieved with a new tread (Fig.
8.3). The reason for this unexpected result is due to
the greater amount of rubber interaction with the
ground surface for a given size of contact patch.
It therefore develops a larger reaction force which
i:e: Coefficient of friction () ˆ
where ˆ coefficient of friction
F ˆ frictional force (N)
W ˆ normal reaction (weight of body) (N)
Strictly speaking, the coefficient of friction does
not take into account the surface area tread pattern which maximizes the interlocking mechanism
Fig. 8.1 Sliding block and board
the speed of the vehicle. If, however, the braking is
such that the wheels are still rotating, the actual
speed between the tyre tread and the road must be
less than that of the vehicle. Even on surfaces giving
good braking when wet, maximum coefficients occur
at around 10±20% slip. This means that the
actual speed between the tyre tread and the road
is around one eighth of the vehicle speed or less.
Under these conditions, it is possible to visualize
that the high initial peak value occurs because the
actual tyre ground relative speed relates to a
locked wheel condition at a very low vehicle speed
(Fig. 8.3).
The ability to utilize initial peak retardation
under controlled conditions is a real practical
asset to vehicle retardation and, because the tyre
is still rolling, to vehicle directional control.
Braking effectiveness can therefore be controlled
and improved if the wheels are prevented from
completely locking in contrast to the wheels actually being locked when the brakes are applied. Thus
when braking from different speeds (Fig. 8.4) it can
be seen that the unlocked wheels produce a higher
peak coefficient of adhesive friction as opposed to
the locked condition which generates only a sliding
coefficient of adhesive friction. In both situations
the coefficient of adhesive friction decreases as
the speed from which braking first commences
Fig. 8.3 Effect of surface condition on the coefficient of
adhesive friction with natural and synthetic rubber using
new and bald tyre treads
opposes the movement of the tyre. Under ideal road
conditions and the amount of deformable rubber
actually in contact with the road maximized for a
given contact path area, the retarding force which
can be generated between the tyre and ground can
equal the vertical load the wheel supports. In other
words, the coefficient of adhesive friction can reach
a value of 1.0. However, any deterioration in surface
roughness due to surface ridges being worn, or chippings becoming submerged in asphalt, or the slightest amount of wetness completely changes the
situation. A smooth bald tyre will not be able to
grip the contour of the road, whereas the tyre with a
good tread pattern will easily cope and maintain
a relatively high value of retardant force.
When transmitting tractive or braking forces, the
tyre is operating with slip or creep. It is believed
that the maximum friction is developed when a
maximum number of individual tread elements
are creeping at or near an optimum speed relative
to the ground. The distribution by each element of
the tread is not equal nor is it uniform throughout
the contact patch. The frictional forces developed
depend upon the pressure distribution within the
contact patch area and the creep effects. Once bodily slippage begins to occur in one region of the
contact area, the progression to the fully sliding
condition of the contact area as a whole is extremely
Under locked wheel conditions, the relative
sliding speed between a tyre tread and the road is
Fig. 8.4 Effect of speed on both peak and sliding
coefficient of adhesive friction
8.1.2 Grip control
Factors influencing the ability of a tyre to grip the
road when being braked are:
the vehicle speed,
the amount of tyre wear,
the nature of the road surface,
the degree of surface wetness.
Vehicle speed (Fig. 8.4) Generally as the speed of
the vehicle rises, the time permitted for tread to
ground retardation is reduced so that the grip or
coefficient of adhesive friction declines (Fig. 8.4).
Tyre wear (Fig. 8.5) As the tyre depth is reduced,
the ability for the tread to drain off water being
swept in front of the tread is reduced. Therefore
with increased vehicle speed inadequate drainage
will reduce the tyre grip when braking (Fig. 8.5).
Road surface wetness (Fig. 8.6) The reduction in
tyre grip when braking from increased vehicle speed
drops off at a much greater rate as the rainfall
changes from light rain, producing a surface water
depth of 1 mm, to a heavy rainstorm flooding the
road to a water depth of about 2.5 mm (Fig. 8.6).
Fig. 8.6 Effect of speed on relative tyre grip with various
road surface water depths
Road surface texture (Fig. 8.7) A new tyre braked
from various speeds will generate a higher peak
coefficient of adhesive friction with a smaller fall
off at the higher speeds on wet rough surfaces compared to braking on wet smooth surfaces (Fig. 8.7).
Fig. 8.7 Effect of speed on the coefficient of adhesive
friction with both wet rough and smooth surfaces
The reduction in the coefficient of adhesive friction
when braking with worn tyres on both rough and
particularly smooth wet surfaces will be considerably greater.
8.1.3 Road surface texture (Fig. 8.8)
A road surface finish may be classified by its texture
which may be broadly divided in macrotexture,
Fig. 8.5 Effect of speed on relative tyre grip with various
tread depth when braking on a wet road
trate any remaining film of water and so interact
with the tread elements. If these conditions are
fulfilled, a well designed tyre tread will provide
grip not only under dry conditions but also in wet
weather. A worn road surface may be caused by the
hard chippings becoming embedded below the soft
asphalt matrix or the microtexture of these chippings may become polished. In the case of concrete
roads, the roughness of the brushed or mechanically ridged surface may become blunted and over
smooth. To obtain high frictional grip over a wide
speed range and during dry and wet conditions, it is
essential that the microtexture is harsh so that pure
rubber to road interaction takes place.
Fig. 8.8 Terminology and road surface texture
8.1.4 Braking characteristics on wet roads
(Fig. 8.9)
Maximum friction is developed between a rubber
tyre tread and the road surface under conditions of
slow movement or creep. A tyre's braking response
on a smooth wet road with the vehicle travelling at
a speed, say 100 km/h, will show the following
characteristics (Fig. 8.9).
When the brakes are in the first instance steadily
applied, the retardation rate measured as a fraction
of the gravitational acceleration (g m=s2 ) will rise
rapidly in a short time interval up to about 0.5 g.
This phase of braking is the normal mode of braking
when driving on motorways. In traffic, it enables the
which represents the surface section peak to valley
ripple or roughness, and microtexture which is a
measure of the smoothness of the ripple contour
(Fig. 8.8). Further subdivisions may be made;
macrotexture may range from closed or fine going
onto open or coarse whereas microtexture may
range from smooth or polished extending to sharp
or harsh.
For good tyre grip under dry and wet conditions
the road must fulfil two requirements. Firstly, it
must have an open macrotexture to permit water
drainage. Secondly, it should have a microtexture
which is harsh; the asperities of the texture ripples
should consist of many sharp points that can pene-
Fig. 8.9 Possible retardation braking cycle on a wet road
driver to reduce the vehicle speed fairly rapidly with
good directional stability and no wheel lock taking
place. If an emergency braking application becomes
necessary, the driver can raise the foot brake effort
slightly to bring the vehicle retardation to its peak
value of just over 0.6 g, but then should immediately
release the brake, pause and repeat this on-off
sequence until the road situation is under control.
Failing to release the brake will lock the wheels so
that the tyre road grip changes from one of rolling to
sliding. As the wheels are prevented from rotating,
the braking grip generated between the contact
patches of the tyres drops drastically as shown in
the crash stop phase. If the wheels then remain
locked, the retardation rate will steady at a much
lower value of just over 0.2 g. The tyres will now
be in an entirely sliding mode, with no directional
stability and with a retardation at about one third
of the attainable peak value. With worn tyre treads
the braking characteristics of the tyres will be similar
but the braking retardation capacity is considerably
all returned as strain energy as the tyre takes up its
original shape. (Note that this has nothing to do
with a tractive force being applied to the wheel to
propel it forward.) Unfortunately when the carcass
is stressed, the strain produced is a function of the
stress. On releasing the stress, because the tyre
material is not perfectly elastic, the strain lags
behind so that the strain for a given value of stress
is greater when the stress is decreasing than when it
is increasing. Therefore, on removing the stress
completely, a residual strain remains. This is known
as hysteresis and it is the primary cause of the rolling
resistance of the tyre.
The secondary causes of rolling resistance are air
circulation inside the tyre, fan effect of the rotating
tyre by the air on the outside and the friction
between the tyre and road caused by tread slippage.
A typical analysis of tyre rolling resistance losses at
high speed can be taken as 90±95% due to internal
hysteresis, 2±10% due to friction between the tread
and ground, and 1.5±3.5% due to air resistance.
Rolling resistance is influenced by a number of
factors as follows:
8.1.5 Rolling resistance (Figs 8.10 and 8.11)
When a loaded wheel and tyre is compelled to roll
in a given direction, the tyre carcass at the ground
interface will be deflected due to a combination of
the vertical load and the forward rolling effect on
the tyre carcass (Fig. 8.10). The vertical load tends
to flatten the tyre's circular profile at ground level,
whereas the forward rolling movement of the wheel
will compress and spread the leading contact edge
and wall in the region of the tread. At the same
time, the trailing edge will tend to reduce its contact
pressure and expand as it is progressively freed
from the ground reaction. The consequences of
the continuous distortion and recovery of the tyre
carcass at ground level means that energy is being
used in rolling the tyre over the ground and it is not
a) cross-ply tyres have higher rolling resistance
than radial ply (Fig. 8.11),
b) the number of carcass plies and tread thickness
increase the rolling resistance due to increased
c) natural rubber tyres tend to have lower rolling
resistance than those made from synthetic rubber,
Fig. 8.10 Illustration of side wall distortion at ground
Fig. 8.11
Effect of tyre construction on rolling resistance
d) hard smooth dry surfaces have lower rolling
resistances than rough or worn out surfaces,
e) the inflation pressure decreases the rolling resistance on hard surfaces,
f) higher driving speed increases the rolling resistance due to the increase in work being done in
deforming the tyre over a given time (Fig. 8.11),
g) increasing the wheel and tyre diameter reduces the
rolling resistance only slightly on hard surfaces but
it has a pronounced effect on soft ground,
h) increasing the tractive effort also raises the rolling resistance due to the increased deformation
of the tyre carcass and the extra work needed to
be done.
in the direction of the leading edge of the tread
contact patch is continuously opposed by the tyre
contact patch reaction on the ground. Before it
enters the contact patch region a portion of the
tread and casing will be deformed and compressed.
Hence the distance that the tyre tread travels when
subjected to a driving torque will be less than that
in free rolling (Fig. 8.12).
If a braking torque is now applied to the wheel
and tyre, the inertia on the vehicle will tend to pull
the wheel forward while the interaction between the
tyre contact patch and ground will oppose this
motion. Because of this action, the casing and tread
elements on the leading side of the tyre become
stretched just before they enter the contact patch
region in contrast with the compressive effect for
driving tyres (Fig. 8.13). As a result, when braking
torque is applied the distance the tyre moves will be
greater than when the tyre is subjected to free rolling
only. The loss or gain in the distance the tread
8.1.6 Tractive and braking effort (Figs 8.12,
8.13, 8.14, 8.15, 8.16 and 8.17)
A tractive effort at the tyre to ground interface is
produced when a driving torque is transmitted to
the wheel and tyre. The twisting of the tyre carcass
Fig. 8.12
Deformation of a tyre under the action of a driving torque
Fig. 8.13
Deformation of a tyre under the action of a braking torque
Fig. 8.14 Effect of tyre slip on tractive effort
travels under tractive or braking conditions relative
to that in free rolling is known as deformation slip,
and it can be said that under steady state conditions
slip is a function of tractive or braking effort.
When a driving torque is applied to a wheel and
tyre there will be a steep initial rise in tractive force
matched proportionally with a degree of tyre slip,
due to the elastic deformation of the tyre tread. Eventually, when the tread elements have reached their
distortion limit, parts of the tread elements will begin
to slip so that a further rise in tractive force will
produce a much larger increase in tyre slip until the
peak or limiting tractive effort is developed. This
normally corresponds to on a hard road surface to
roughly 15±20% slip (Fig. 8.14). Beyond the peak
tractive effort a further increase in slip produces an
unstable condition with a considerable reduction in
tractive effort until pure wheel spin results (the tyre
just slides over the road surface). A tyre subjected to
a braking torque produces a very similar braking
effort response with respect to wheel slip, which is
now referred to as skid. It will be seen that the maximum braking effort developed is largely dependent
upon the nature of the road surface (Fig. 8.15) and
the normal wheel loads (Fig. 8.16), whereas wheel
speed has more influences on the unstable skid region
of a braking sequence (Fig. 8.17).
Fig. 8.15
Effect of ground surface on braking effort
Fig. 8.16
Effect of vertical load on braking effort
resist the tyre slipping over the surface when the
tyre is subjected to longitudinal (tractive or braking)
forces and lateral (side) (cornering or crosswind)
forces simultaneously. Therefore the resultant components of the longitudinal and lateral forces must
not exceed the tread to ground resultant reaction
force generated by all of the tread elements within
the contact area biting into the ground.
The relative relationship of the longitudinal and
lateral forces acting on the tyre can be shown by
8.1.7 Tyre reaction due to concurrent
longitudinal and lateral forces (Fig. 8.18)
A loaded wheel and tyre rolling can generate only
a limited amount of tread to ground reaction to
force. If the side force caused either by cornering or
a crosswind is large, the traction or braking effort
must be much reduced.
8.2 Tyre materials
Fig. 8.17
8.2.1 The structure and properties of rubber
(Figs 8.19, 8.20 and 8.21)
The outside carcass and tread of a tyre is made from
a rubber compound that is a mix of several substances to produce a combination of properties
necessary for the tyre to function effectively. Most
metallic materials are derived from simple molecules held together by electrostatic bonds which
sustain only a limited amount of stretch when subjected to tension (Fig. 8.19). Because of this, the
material's elasticity may be restricted to something
like 2% of its original length. Rubber itself may be
either natural or synthetic in origin. In both cases
the material consists of many thousands of long
chain molecules all entangled together. When
stretched, the giant rubber molecules begin to
untangle themselves from their normal coiled state
and in the process of straightening out, provide a
considerable amount of extension which may be of
the order of 300% of the material's original length.
Thus it is not the electrostatic bonds being stretched
Effect of vehicle speed on braking effort
resolving both forces perpendicularly to each other
within the boundary of limiting reaction force
circle (Fig. 8.18(a and b)). This circle with its vector
forces shows that when longitudinal forces due to
traction or braking forces is large (Fig. 8.18(c and
d)), the tyre can only sustain a much smaller side
Fig. 8.18 (a±d)
Limiting reaction force circle
Fig. 8.19 Metal atomic lattice network
Fig. 8.20 Raw rubber network of long chain molecules
Fig. 8.21 Vulcanized rubber cross-linked network of long chain molecules
but the uncoiling and aligning of the molecules in
the direction of the forces pulling the material apart
(Fig. 8.20). Consequently, when the tensile force is
removed the molecules revert to their free state
and thereby draw themselves into an entangled
network again. Hence it is not the bonds being
stretched but the uncoiling and aligning of the molecules in the direction of the force pulling the material
pressure. The chemical reaction produced is known
either as curing or more commonly as vulcanization
(named after Vulcan, the Roman god of fire). As a
result, the sulphur molecules form a network of
cross-links between some of the giant rubber molecules (Fig. 8.21). The outcome of the cross-linking
between the entangled long chain molecules is that it
makes it more difficult for these molecules to slip
over each other so that the rubber becomes stronger
with a considerable reduction in flexibility.
Vulcanization To reduce the elasticity and to
increase the strength of the rubber, that is to restrict
the molecules sliding past each other when the substance is stretched, the rubber is mixed with a small
amount of sulphur and then heated, usually under
Initiators and accelerator To start off and speed
up the vulcanization process, activators such as a
metallic zinc oxide are used to initiate the reaction
and an organic accelerator reduces the reaction
time and temperature needed for the sulphur to
produce a cross-link network.
Material hysteresis This is the sluggish response of
a distorted material taking up its original form so
that some of the energy put into deforming the carcass, side walls and tread of a tyre at the contact
patch region will still not be released when the tyre
has completed one revolution and the next distortion
period commences. As the cycle of events continues,
more and more energy will be absorbed by the tyre,
causing its temperature to rise. If this heat is not
dissipated by the surrounding air, the inner tyre
fabric will eventually become fatigued and therefore
break away from the rubber encasing it, thus
destroying the tyre. For effective tyre grip a high
hysteresis material is necessary so that the distorted
rubber in contact with the ground does not immediately spring away from the surface but is inclined to
mould and cling to the contour of the road surface.
Carbon black Vulcanized rubber does not have
sufficient abrasive resistance and therefore its rate
of wear as a tyre tread material would be very high.
To improve the rubber's resistance against wear
and tear about a quarter of a rubber compound
content is made up of a very fine carbon powder
known as carbon black. When it is heated to a
molten state the carbon combines chemically with
the rubber to produce a much harder and tougher
wear resistant material.
Oil extension To assist in producing an even
dispersion of the rubber compound ingredients
and to make processing of the tyre shape easier,
an emulsion of hydrocarbon oil is added (up to 8%)
to the rubber latex to dilute or extend the rubber.
This makes the rubber more plastic as opposed to
elastic with the result that it becomes tougher,
offers greater wear resistance and increases the
rubber's hysteresis characteristics thereby improving its wet grip properties.
Material fatigue This is the ability of the tyre
structure to resist the effects of repeated flexing
without fracture, particularly with operating temperatures which may reach something of the order
of 100 C for a heavy duty tyre although temperatures of 80±85 C are more common.
8.2.3 Natural and synthetic rubbers
Synthetic materials which have been developed
as substitutes for natural rubber and have been
utilized for tyre construction are listed with natural
rubber as follows:
Anti-oxidants and -ozonates Other ingredients such
as an anti-oxidant and anti-ozonate are added to
preserve the desirable properties of the rubber compound over its service life. The addition of antioxidants and -ozonates (1 or 2 parts per 100 parts
of rubber) prevents heat, light and particularly oxygen ageing the rubber and making it hard and brittle.
8.2.2 Mechanical properties
To help the reader understand some of the terms
used to define the mechanical properties of rubber
the following brief definitions are given:
Natural rubber (NR)
Chloroprene (Neoprene) rubber (CR)
Styrene±butadiene rubber (SBR)
Polyisoprene rubber (IR)
Ethylene propylene rubber (EPR)
Polybutadiene rubber (BR)
Isobutene±isoprene (Butyl) rubber (IIR)
Natural rubber (NR) Natural rubber has good
wear resistance and excellent tear resistance. It
offers good road holding on dry roads but retains
only a moderately good grip on wet surfaces. One
further merit is its low heat build-up, but this is
contrasted by high gas permeability and its resistance to ageing and ozone deterioration is only fair.
The side walls and treads have been made from
natural rubber but nowadays it is usually blended
with other synthetic rubbers to exploit their desirable properties and to minimize their shortcomings.
Material resilience This is the ability for a solid
substance to rebound or spring back to its original
dimensions after being distorted by a force. A
material which has a high resilience generally has
poor road grip as it tends to spring away from the
ground contact area as the wheel rolls forward.
Material plasticity This is the ability for a solid
material to deform without returning to its original
shape when the applied force is removed. A material which has a large amount of plasticity promotes
good road grip as each layer of material tends to
cling to the road surface as the wheel rolls.
Chloroprene (Neoprene) rubber (CR) This synthetic rubber is made from acetylene and hydro279
chloric acid. Wear and tear resistance for this rubber
compound, which was one of the earliest to compete with natural rubber, is good with a reasonable
road surface grip. A major limitation is its inability
to bond with the carcass fabric so a natural rubber
film has to be interposed between the cords and the
Neoprene covering. Neoprene rubber has a moderately low gas permeability and does not show signs
of weathering or ageing throughout a tyre's working life. When blended with natural rubber it is
particularly suitable for side wall covering.
to be mixed with large amounts of cheap carbon
black and oil without destroying its rubbery properties. It has excellent abrasive ageing and ozone
resistance with varying road holding qualities in
wet weather depending upon the compound composition. Skid resistance on ice has also been varied
from good to poor. A great disadvantage, however,
is that the rubber compound bonds poorly to cord
fabric. Generally, the higher the ethylene content
the higher the abrasive resistance, but at the
expense of a reduction in skid resistance on ice.
Rubber compounds containing EPR have not
proved to be successful up to the present time.
Styrene±butadiene rubber (SBR) Compounds of
this material are made from styrene (a liquid) and
butadiene (a gas). It is probably the most widely
used synthetic rubber within the tyre industry.
Styrene±butadiene rubber (SBR) forms a very
strong bond to fabrics and it has a very good
resistance to wear, but suffers from poor tear resistance compared to natural rubber. One outstanding
feature of this rubber is its high degree of energy
absorption or high hysteresis and low resilience. It
is these properties which give it exceptional grip,
especially on wet surfaces. Due to the high heat
build up, SBR is restricted to the tyre tread while
the side walls are normally made from low hysteresis compounds which provide greater rebound
response and run cooler. Blending SBR with NR
enables the best properties of both synthetic and
natural rubber to be utilized so that only one rubber compound is necessary for some types of car
tyres. The high hysteresis obtained with SBR is
partially achieved by using an extra high styrene
content and by adding a large proportion of oil to
extend the compound, the effects being to increase
the rubber plastic properties and to lower its resilience (i.e. reduce its rebound response).
Polybutadiene rubber (BR) This rubbery material
has outstanding wear resistance properties and is
exceptionally stable with temperature changes. It
has a high resilience that is a low hysteresis level.
When blended with SBR in the correct proportions,
it reduces the wet road holding slightly and considerably improves its ability to resist wear. Because of its
high resilience (large rebound response), if mixed in
large proportions, the road holding in wet weather
can be relatively poor. It is expensive to produce.
When it is used for tyres it is normally mixed with
SBR in the proportion of 15 to 50%.
Isobutene±isoprene (Butyl) rubber (IIR) Rubber of
this kind has exceptionally low permeability to gas.
In fact it retains air ten times longer than tubes
made from natural rubber, with the result that it
has been used extensively for tyre inner tubes and
for linings of tubeless tyres. Unfortunately it will
not blend with SBR and NR unless it is chlorinated,
but in this way it can be utilized as an inner tube
lining material for tubeless tyres. The resistance
to wear is good and it has a high hysteresis so
that it responds more like plastic than rubber to
distortion at ground level. Road grip is good for
both dry and wet conditions. When mixed with
carbon black its desirable properties are generally
improved. Due to its high hysteresis tyre treads
made from this material do not generate noise in
the form of squeal since it does not readily give out
energy to the surroundings.
Polyisoprene rubber (IR) This compound has very
similar characteristics to natural rubber but has
improved wear and particularly tear resistance with
a further advantage of an extremely low heat build up
with normal tyre flexing. These properties make this
material attractive when blended with natural rubber
and styrene±butadiene rubber to produce tyre treads
with very high abrasion resistance. For heavy duty
application such as track tyres where high temperatures and driving on rough terrains are a problem, this material has proved to be successful.
8.2.4 Summary of the merits and limitations of
natural and synthetic rubber compounds
Some cross-ply tyres are made from one compound
from bead to bead, but the severity of the carcass
flexure with radial ply tyres encourages the
manufacturers of tyres to use different rubber
Ethylene propylene rubber (EPR) The major
advantage of this rubber compound is its ability
composition for various parts of the tyre structure
so that their properties match the duty requirements of each functional part of the tyre
(i.e. tread, side wall, inner lining, bead etc.).
Side walls are usually made from natural rubber
blended with polybutadiene rubber (BR) or styrene±butadiene rubber (SBR) or to a lesser extent
Neoprene or Butyl rubber or even natural rubber
alone. The properties needed for side wall material
are a resistance against ozone and oxygen attack,
a high fatigue resistance to prevent flex cracking
and good compatibility with fabrics and other
rubber compounds when moulded together.
Tread wear fatigue life and road grip depends to
a great extent upon the surrounding temperatures,
weather conditions, be they dry, wet, snow or ice
bound, and the type of rubber compound being
used. A comparison will now be made with natural
rubber and possibly the most important synthetic
rubber, styrene±butadiene (SBR). At low temperatures styrene±butadiene (SBR) tends to wear more
than natural rubber but at higher temperatures the
situation reverses and styrene±butadiene rubber
(SBR) shows less wear than natural rubber. As
the severity of the operating condition of the tyre
increases SBR tends to wear less relative to NR.
The fatigue life of all rubber compounds is reduced
as the degree of cyclic distortion increases. For
small tyre deflection SBR has a better fatigue life
but when deflections are large NR provides a
longer service life. Experience on ice and snow
shows that NR offers better skid resistance, but as
temperatures rise above freezing, SBR provides an
improved resistance to skidding. This cannot be
clearly defined since it depends to some extent on
the amount of oil extension (plasticizer) provided
in the blending in both NR and SBR compounds.
Oil extension when included in SBR and NR provides similarly improved skid resistance and in
both cases becomes inferior to compounds which
do not have oil extension.
Two examples of typical rubber compositions
suitable for tyre treads are:
a) High styrene butadiene rubber
Oil extended butadiene rubber
Carbon black
b) Styrene butadiene rubber
Natural rubber
Carbon black
8.3 Tyre tread design
8.3.1 Tyre construction
The construction of the tyre consists basically of
a carcass, inner beads, side walls, crown belt
(radials) and tread.
Carcass The carcass is made from layers of textile
core plies. Cross-ply tyres tend to still use nylon
whereas radial-ply tyres use either raylon or polyester.
Beads The inside diameter of both tyre walls support the carcass and seat on the wheel rim. The
edges of the tyre contacting the wheel are known
as beads and moulded inside each bead is a
strengthening endless steel wire cord.
Side walls The outside of the tyre carcass, known
as the side walls, is covered with rubber compound.
Side walls need to be very flexible and capable of
protecting the carcass from external damage such
as cuts which can occur when the tyre is made to
climb up a kerb.
Bracing belt Between the carcass and tyre tread is
a crown reinforcement belt made from either synthetic fabric cord such as raylon or for greater
strength steel cores. This circumferential endless
cord belt provides the rigidity to the tread rubber.
Tread The outside circumferential crown portion
of the tyre is known as the tread. It is made from
a hard wearing rubber compound whose function
is to grip the contour of the road.
8.3.2 Tyre tread considerations
The purpose of a pneumatic tyre is to support the
wheel load by a cushion of air trapped between the
well of the wheel rim and the toroid-shaped casing
known as the carcass. Wrapped around the outside
of the tyre carcass is a thick layer of rubber compound known as the tread whose purpose is to protect the carcass from road damage due to tyre
impact with the irregular contour of the ground and
the abrasive wear which occurs as the tyre rolls along
the road. While the wheel is rotating the tread provides driving, braking, cornering and steering grip
between the tyre and ground. Tyre grip must be
available under a variety of road conditions such as
smooth or rough hard roads, dry or wet surfaces,
muddly tracks, fresh snow or hard packed snow and
ice and sandy or soft soil terrain. Tread grip may be
defined as the ability of a rolling tyre to continuously
develop an interaction between the individual tread
elements and the ground so that any longitudinal
(driving) or lateral (side) forces imposed on the
wheel will not be able to make the tread in contact
with the ground slide.
A tyre tread pattern has two main functions:
enclosed channels made by the road sealing the
underside of the grooves. Water therefore emerges
from the trailing side of the contact patch in the
form of jets. If these grooves are to be effective,
their total cross-sectional area should be adequate
to channel all the water immediately ahead of the
leading edge of the contact patch away. If it cannot
cope the water will become trapped between the
tread ribs or blocks so that these elements lift and
become separated from the ground, thus reducing
the effective area of the contact patch and the tyre's
ability to grip the ground.
To speed up the water removal process under the
contact patch, lateral grooves may be used to join
together the individual circumferential grooves and
to provide a direct side exit for the outer circumferential grooves. Normally many grooves are preferred to a few large ones as this provides a better
drainage distribution across the tread.
1 to provide a path for drainage of water which
might become trapped between the tyre contact
patch and the road,
2 to provide tread to ground bite when the wheel is
subjected to both longitudinal and lateral forces
under driving conditions.
8.3.3 Tread bite
Bite is obtained by selecting a pattern which divides
the tread into many separate elements and providing each element with a reasonably sharp well
defined edge. Thus as the wheel rotates these
tread edges engage with the ground to provide a
degree of mechanical tyre to ground interlock in
addition to the frictional forces generated when
transmitting tractive or braking forces.
The major features controlling the effectiveness
of the tread pattern in wet weather are:
Tread ribs (Fig. 8.22(a and b)) Circumferential
ribs not only provide a supportive wearing surface
for the tyre but also become the walls for the drainage
grooves (Fig. 8.22(a and b)). Lateral (transverse)
ribs or bars provide the optimum bite for tractive
and braking forces but circumferential ribs are most
effective in controlling cornering and steering stability. To satisfy both longitudinal and lateral directional requirements which may be acting concurrently
on the tyre, ribs may be arranged diagonally or in
the form of zig-zag circumferential ribs to improve
the wiping effect across the tread surface under wet
conditions. It is generally better to break the tread
pattern into many narrow ribs than a few wide ones,
as this prevents the formation of hydrodynamic
water wedges which may otherwise tend to develop
drainage grooves or channels,
load carrying ribs,
load bearing blocks,
multiple microslits or sipes.
Tread drainage grooves (Fig. 8.22(a, b, c and d))
The removal of water films from the tyre to ground
interface is greatly facilitated by having a number of
circumferential grooves spaced out across the tread
width (Fig. 8.22(a)). These grooves enable the leading elements of the tread to push water through the
Fig. 8.22 (a±d)
Basic tyre tread patterns
with the consequent separation of the tread elements
from the road.
wedge action may result, causing a mild form of
aquaplaning to take place.
For tread block elements to maintain their wiping
action on wet surfaces, wear should be from toe to
heel (Fig. 8.23(a)). If, however, wear occurs in the
reverse order, that is from heel to toe (Fig. 8.23(b)),
the effectiveness of the tread pattern will be severely
reduced since the tread blocks then become the platform for a hydrodynamic water wedge which at
speed tries to lift the tread blocks off the ground.
Tread blocks (Figs 8.22(c and d) and 8.23(a and b))
If longitudinal circumferential grooves in the middle of the tread are complemented by lateral (transverse) grooves channelled to the tread shoulders,
then with some tread designs the drainage of water
can be more effective at speed. The consequences of
both longitudinal and lateral drainage channels is
that the grooves encircle portions of the tread so
that they become isolated island blocks (Fig. 8.22
(c and d)). These blocks can be put to good use as
they provide a sharp wiping and biting edge where
the interface of the tread and ground meet. To
improve their biting effectiveness for tractive and
braking forces as well as steering and cornering
forces, these forces may be resolved into diagonal
resultants so that the blocks are sometimes
arranged in an oblique formation. A limitation to
the block pattern concept is caused by inadequate
support around the blocks so that under severe
operating conditions, the bulky rubber blocks
tend to bend and distort. This can be partially
overcome by incorporating miniature buttresses
between the drainage grooves which lean between
blocks so that adjacent blocks support each other.
At the same time, drainage channels which burrow
below the high mounted buttresses are prevented
from closing. Tread blocks in the form of bars, if
arranged in a herringbone fashion, have proved to
be effective on rugged ground. Square or rhombusshaped blocks provide a tank track unrolling
action greatly reducing movement in the tread contact area. This pattern helps to avoid the break-up
on the top layer of sand or soil and thus prevents
the tyre from digging into the ground. Because of
the inherent tendency of the individual blocks to
bend somewhat when they are subjected to ground
reaction forces, they suffer from toe to heel rolling
action which causes blunting of the leading edge
and trailing edge feathering. Generally tyres which
develop this type of wear provide a very good water
sweeping action when new, which permits the tread
elements to bite effectively into the ground, but
after the tyre has been on the road for a while, the
blunted leading edge allows water to enter underneath the tread elements. Consequently the slightest amount of water interaction between the block
elements and ground reduces the ability for the
tread to bite and in the extreme cases under locked
wheel braking conditions a hydrodynamic water
Tread slits or sipes (Figs 8.22(a, b, c and d) and
8.24(a, b and c)) Microslits, or sipes as they are
commonly called, are incisions made at the surface
of the tyre tread, going down to the full depth of
the tread grooves. They resemble a knife cut, except
that instead of being straight they are mostly of
a zig-zag nature (Fig. 8.22(a, b, c and d)). Normally
these sipes terminate within the tread elements, but
sometimes one end is permitted to intersect the side
wall of a drainage groove. In some tread patterns
the sipes are all set at a similar angle to each other,
the zig-zag shape providing a large number of edges
which point in various directions. Other designs
have sets of sipes formed at different angles to
each other so that these sipes are effective whichever way the wheel points and whatever the direction the ground reaction forces operate.
Sipes or slits in their free state are almost closed,
but as they move into the contact patch zone the
ribs or blocks distort and open up (Fig. 8.24(a)).
Because of this, the sipe lips scoop up small quantities of water which still exist underneath the tread.
This wiping action enables some biting edge reaction with the ground. Generally, the smaller the
sipes are and more numerous they are the greater
will be their effective contribution to road grip. The
Fig. 8.23 (a and b)
Effect of irregular tread block wear
Fig. 8.24 (a±c) Effectiveness of microslits on wet road surfaces
normal spacing of sipes (microslits) on a tyre tread
makes them ineffective on a pebbled road surface
because there will be several pebbles between the
pitch of the sipes (Fig. 8.24(b)), and water will lie
between these rounded stones, therefore only a few
of the stones will be subjected to the wiping edge
action of the opened lips. An alternative method to
improve the wiping process would be to have many
more wiping slits (Fig. 8.24(c)), but this is very
difficult to implement with the present manufacturing techniques. The advantages to be gained by
multislits are greatest under conditions of low friction associated with thin water films on smooth
and polished road surfaces. This is because the
road surface asperities are not large and sharp
enough to penetrate the thin water film trapped
under plain ribs and blocks.
ively at all speeds tend to have tread blocks situated in an oblique fashion with a network of
surrounding drainage grooves which provide both
circumferential and lateral water release.
Winter car tyres (Fig. 8.25(d, e and f)) Winter car
tyres are normally very similar to the general duty
car tyre but the tread grooves are usually wider to
permit easier water dispersion and to provide better
exposure of the tread blocks to snow and soft ice
without sacrificing too much tread as this would
severely reduce the tyre's life.
Truck tyres (Fig. 8.25(g and h)) Truck tyres
designed for steered axles usually have circumferential zig-zag ribs and grooves since they provide very
good lateral reaction when being steered on curved
tracks. Drive axle tyres, on the other hand, are
designed with tread blocks with adequate grooving
so that optimum traction grip is obtained under
both dry and wet conditions. Some of these tyres
also have provision for metal studs to be inserted for
severe winter hard packed snow and ice conditions.
Selection of tread patterns (Fig. 8.25(a±1))
Normal car tyres (Fig. 8.25(a, b and c)) General
duty car tyres which are capable of operating effect284
Fig. 8.25 (a±l)
Survey of tyre tread patterns
Fig. 8.25 contd
Off/on road vehicles (Fig. 8.25(i)) Off/on road
vehicle tyres usually have a much simpler bold
block tread with a relatively large surrounding
groove. This enables each individual block to react
independently with the ground and in this manner
bite and exert traction on soil which may be hard on
the surface but soft underneath without break-up of
the top layer, thus preventing the tyre digging in.
The tread pattern blocks are also designed to be
small enough to operate on hard surfaced roads at
moderate speeds without excessive ride harshness.
arranged with separate overlapping diagonal bars,
as this configuration tends to provide exceptionally
good traction on muddy soil, snow and soft ice.
8.3.4 The three zone concept of tyre to ground
contact on a wet surface (Fig. 8.26)
The interaction of a tyre with the ground when
rolling on a wet surface may be considered in
three phases (Fig. 8.26):
Leading zone of unbroken water film (1) The leading zone of the tread contacts the stagnant water
film covering the road surface and displaces the
majority of the water into the grooves between
the ribs and blocks of the tread pattern.
Truck and tractor off road and cross-country tyres
(Fig. 8.25(j, k and l)) Truck or tractor tyres
designed for building sites or quarries generally
have slightly curved rectangular blocks separated
with wide grooves to provide a strong flexible casing and at the same time present a deliberately
penetrating grip. Cross-country tyres which tend
to operate on soft soil tend to prefer diagonal
bars either merging into a common central rib or
Intermediate region of partial breakdown of water
film (2) The middle zone of the tread traps and
reduces the thickness of the remaining water
Fig. 8.26 Tyre to ground zones of interaction
between the faces of the ribs or blocks and ground
so that some of the road surface asperities now
penetrate through the film of water and may actually touch the tread. It is this region which is responsible for the final removal of water and is greatly
assisted by multiple sipes and grooved drainage
channels. If the ribs and blocks are insufficiently
relieved with sipes and grooves it is possible that
under certain conditions aquaplaning may occur in
this region.
The effectiveness of this phase is determined to
some extent by the texture of the road surface, as
this considerably influences the dryness and
potency of the third road grip phase.
created. This hydrodynamic pressure acts between
the tyre and ground, its magnitude being proportional to the square of the wheel speed. With the
wheel in motion, the water will form a converging
wedge between the tread face and ground and so
exert an upthrust on the underside of the tread. As
a result of the pressure generated, the tyre tread will
tend to separate itself from the ground. This condition is known as aquaplaning or hydroplaning. If
the wheel speed is low only the front region of the
tread rides on the wedge of water, but if the speed is
rising the water wedge will progressively extend
backward well into the contact patch area (Fig.
8.27). Eventually the upthrust created by the product of the hydrodynamic pressure and contact area
equals the vertical wheel load. At this point the tyre
is completely supported by a cushion of water and
therefore provides no traction or directional control.
If the tread has circumferential (longitudinal) and
transverse (lateral) grooves of adequate depth then
the water will drain through these passages at
ground level so that aquaplaning is minimized even
at high speeds. As the tyre tread wears the critical
speed at which aquaplaning occurs becomes much
lower. On very wet roads a bald tyre is certain to be
subjected to aquaplaning at speeds above 60 km/h
and therefore the vehicle when driven has no
directional stability. Low aspect ratio tyres may
find it difficult to channel the water away from
the centre of the tread at a sufficiently high
Trailing zone of dry tyre to road contact (3) The
water film has more or less been completely
squeezed out at the beginning of this region so that
the faces of the ribs and blocks bearing down on the
ground are able to generate the bite which produces
the tractive, braking and cornering reaction forces.
8.3.5 Aquaplaning (hydroplaning) (Fig. 8.27)
The performance of a tyre rolling on wet or semiflooded surface will depend to some degree upon
the tyre profile tread pattern and wear. If a smooth
tread is braked over a very wet surface, the forward
rotation of the tyre will drag in the water immediately in front between the tread face and ground
and squeeze it so that a hydrodynamic pressure is
Fig. 8.27
Tyre aquaplaning
1 The tyre side wall height is reduced which
increases the vertical and lateral stiffness of the
2 A shorter and wider contact patch is established.
The overall effect is to raise the load carrying
capacity of the tyre.
3 The wider contact patch enables larger cornering
forces to be generated so that vehicles are able to
travel faster on bends.
4 The shorter and wider contact patch decreases
the pneumatic trail which correspondingly
reduces and makes more consistent the selfaligning torque.
5 The shorter and broader contact patch will,
under certain driving conditions, reduce the slip
angles generated by the tyre when subjected to
side forces. Accordingly this reduces the tread
distortion and as a result scuffing and wear will
6 With an increase in vertical stiffness and a reduction in tyre deflection with lower aspect ratio
tyres, less energy will be dissipated by the tyre
casing so that rolling resistance will be reduced.
This also results in the tyre being able to run
continuously at high speeds at lower temperatures which tends to prolong the tyre's life.
7 The increased lateral stiffness of a low profile tyre
will increase the sensitivity to camber variations
and quicken the response to steering changes.
8 Wider tyre contact patches make it more difficult
for water drainage at speed particularly in the
mid tread region. Hence the tread pattern design
with low profile tyres becomes more critical on
wet roads, if their holding is to match that of
higher aspect ratio tyres.
9 The increased vertical stiffness of the tyre
reduces static deflection of the tyre under load,
so that more road vibrations are transmitted
through the tyre. This makes it a harsher ride
so that ride comfort is reduced unless the suspension design has been able to provide more isolation for the body.
Fig. 8.28 Tyre profiles with different aspect ratios
rate and therefore must rely more on the circumferential grooves than on transverse grooving.
8.3.6 Tyre profile and aspect ratio (Fig. 8.28)
The profile of a tyre carcass considerably influences
its rolling and handling behaviour. Because of the
importance of the tyre's cross-sectional configuration in predicting its suitability and performance for
various applications, the aspect ratio was introduced. This constant for a particular tyre may be
defined as the ratio of the tyre cross-sectional height
(the distance between the tip of the tread to the bead
seat) to that of the section width (the outermost
distance between the tyre walls) (Fig. 8.28).
Section height
Section width
A tyre with a large aspect ratio is referred to as
a high aspect ratio profile tyre and a tyre with a small
aspect is known as a low aspect ratio profile. Until
about 1934 aspect ratios of 100% were used, but
with the better understanding of pneumatic tyre
properties and improvement in tyre construction
lower aspect ratio tyres became available. The availability of lower aspect ratio tyres over the years was
as follows; 1950s±95%, 1962±88% (this was the
standard for many years), 1965±80% and about
1968±70%. Since then for special applications even
lower aspect ratios of 65%, 60%, 55% and even 50%
have become available.
Lowering the aspect ratio has the following
i:e: Aspect ratio ˆ
8.4 Cornering properties of tyres
8.4.1 Static load and standard wheel height
(Figs 8.29 and 8.30)
A vertical load acting on a wheel will radially
distort the tyre casing from a circular profile to
a short flat one in the region of the tread to ground
interface (Fig. 8.29). The area of the tyre contact
with the ground is known as the tyre contact patch
area; its plan view shape is roughly elliptical. The
consequence of this tyre deflection is to reduce the
Fig. 8.29
Illustration of static tyre deflection
standard height of the wheel, that is the distance
between the wheel axis and the ground. Generally
tyre deflection will be proportional to the radial
load imposed on the wheel; increasing the tyre
inflation pressure reduces the tyre deflection for
a given vertical load (Fig. 8.30). Note that there is
an initial deflection (Fig. 8.30) due to the weight of
the wheel and tyre alone. The steepness of the load
deflection curve is useful in estimating the static
stiffness of the tyres which can be interpreted as
a measure of its vibration and ride qualities.
8.4.2 Tyre contact patch (Figs 8.29 and 8.31)
The downward radial load imposed on a road
wheel causes the circular profile of the tyre in contact with the ground to flatten and spread towards
the front and rear of its natural rolling plane. When
the wheel is stationary, the interface area between
the tyre and ground known as the contact patch
will take up an elliptical shape (Fig. 8.29), but if the
wheel is now subjected to a side thrust the grip
between the tread and ground will distort the
patch into a semibanana configuration (Fig. 8.31
(a)). It is the ability of the tyre contact patch casing,
and elements of the tread to comply and change
shape due to the imposed reaction forces, which
gives tyres their steering properties. Generally,
radial ply tyres form longer and broader contact
patches than their counterpart cross-ply tyres,
hence their superior road holding.
Fig. 8.30
Effect of tyre vertical load on static deflection
motion due to road camber, side winds, weight
transfer and centrifugal force caused by travelling
round bends and steering the vehicle on turns. When
a side force, sometimes referred to as lateral force, is
imposed on a road wheel and tyre, a reaction
between the tyre tread contact patch and road surface will oppose any sideway motion. This resisting
force generated at the tyre to road interface is
known as the cornering force (Fig. 8.31(a and b)),
its magnitude being equal to the lateral force
but it acts in the opposite sense. The amount of
8.4.3 Cornering force (Fig. 8.31(a and b))
Tyres are subjected not only to vertical forces but
also to side (lateral) forces when the wheels are in
Fig. 8.31 Tyre tread contact patch distortion when subjected to a side force
cornering force developed increases roughly in proportion with the rise in lateral force until the grip
between the tyre tread and ground diminishes.
Beyond this point the cornering force cannot
match further increases in lateral force with the
result that tyre breakaway is likely to occur. Note
that the greater the cornering force generated
between tyre and ground, the greater the tyre's grip
on the road.
The influencing factors which determine the
amount of cornering force developed between the
tyre and road are as follows:
8.4.4 Slip angle (Fig. 8.31(a))
Any lateral force applied to a road wheel will tend to
push the supple tyre walls sideways, but the opposing tyre to ground reaction causes the tyre contact
patch to take up a curved distorted shape. As a
result, the rigid wheel will point and roll in the
direction it is steered, whereas the tyre region in
contact with the ground will follow the track continuously laid down by the deformed tread path of
the contact patch (Fig. 8.31(a)). The angle made
between the direction of the wheel plane and that
which it travels is known as the slip angle. Provided
the slip angle is small, the compliance of the tyre
construction will allow each element of tread to
remain in contact with the ground without slippage.
Slip angle Initially the cornering force increases
linearly with increased slip angle, but beyond about
four degrees slip angle the rise in cornering force is
non-linear and increases at a much reduced rate (Fig.
8.32), depending to a greater extent on tyre design.
8.4.5 Cornering stiffness (cornering power)
(Fig. 8.32)
When a vehicle travels on a curved path, the centrifugal force (lateral force) tends to push sideways
each wheel against the opposing tyre contact patch
to ground reaction. As a result, the tyre casing and
tread in the region of the contact patch very slightly
deform into a semicircle so that the path followed by
the tyre at ground level will not be quite the same as
the direction of the wheel points. The resistance
offered by the tyre crown or belted tread region by
the casing preventing it from deforming and generating a slip angle is a measure of the tyre's cornering power. The cornering power, nowadays more
Vertical tyre load As the vertical or radial load on
the tyre is increased for a given slip angle, the
cornering force rises very modestly for small slip
angles but at a far greater rate with larger slip angles
(Fig. 8.33).
Tyre inflation pressure Raising the tyre inflation
pressure linearly increases the cornering force for a
given slip angle (Fig. 8.34). These graphs also show
that increasing the tyre slip angle considerably
raises the cornering forces generated.
usually termed cornering stiffness, may be defined as
the cornering force required to be developed for
every degree of slip angle generated.
Cornering stiffness ˆ
Cornering force
Slip angle
In other words, the cornering stiffness of a tyre is
the steepness of the cornering force to slip angle
curve normally along its linear region (Fig. 8.32).
The larger the cornering force needed to be generated for one degree of slip angle the greater the
cornering stiffness of the tyre will be and the
smaller the steering angle correction will be to sustain the intended path the vehicle is to follow. Note
that the supple flexing of a radial ply side wall
should not be confused with the actual stiffness of
the tread portion of the tyre casing.
Fig. 8.33
8.4.6 Centre of pressure (Fig. 8.35)
When a wheel standing stationary is loaded, the
contact patch will be distributed about the geometric centre of the tyre at ground level, but as the
wheel rolls forward the casing supporting the tread
is deformed and pushed slightly to the rear (Fig.
8.35). Thus in effect the majority of the cornering
force generated between the ground and each
element of tread moves from the static centre of
pressure to some dynamic centre of pressure behind
the vertical centre of the tyre, the amount of
displacement corresponding to the wheel construction, load, speed and traction. The larger area of
Effect of tyre vertical load on cornering force
tread to ground reaction will be concentrated behind
the static centre of the wheel and the actual distribution of cornering force from front to rear of the
contact patch is shown by the shaded area between
the centre line of the tyre and the cornering force
plotted line. The total cornering force is therefore
roughly proportional to this shaded area and its
resultant dynamic position is known as the centre
of pressure (Fig. 8.35).
8.4.7 Pneumatic trail (Fig. 8.35)
The cornering force generated at any one time will
be approximately proportional to the shaded area
between the tyre centre line and the cornering force
plotted line so that the resultant cornering force
(centre of pressure) will act behind the static centre
of contact. The distance between the static and
dynamic centres of pressure is known as the pneumatic trail (Fig. 8.35), its magnitude being dependent upon the degree of creep between tyre and
ground, the vertical wheel load, inflation pressure,
speed and tyre constriction. Generally with the
longer contact patch, radial ply tyres have a greater
pneumatic trail than those of the cross-ply
Fig. 8.32
8.4.8 Self-aligning torque (Fig. 8.35)
When a moving vehicle has its steering wheels
turned to negotiate a bend in the road, the lateral
(side) force generates an equal and opposite
Effect of slip angle on cornering force
wheel centre which endeavours to turn both steering wheels towards the straight ahead position.
This self-generating torque attempts to restore the
plane of the wheels with the direction of motion
and it is known as the self-aligning torque (Fig.
8.35). It is this inherent tyre property which helps
steered tyres to return to the original position after
negotiating a turn in the road. The self-aligning
torque (SAT) may be defined as the product of
the cornering force and the pneumatic trail.
TSAT ˆ Fc tp (Nm)
Higher tyre loads increase deflection and accordingly enlarge the contact patch so that the pneumatic trail is extended. Correspondingly this causes
a rise in self-aligning torque. On the other hand
increasing the inflation pressure for a given tyre
load will shorten the pneumatic trail and reduce
the self-aligning torque. Other factors which influence self-aligning torque are load transfer during
braking, accelerating and cornering which alter the
contact patch area. As a general rule, anything
which increases or decreases the contact patch
length raises or reduces the self-aligning torque
respectively. The self-aligning torque is little
affected with small slip angles when braking or
accelerating, but with larger slip angles braking
decreases the aligning torque and acceleration
increases it (Fig. 8.36).
Fig. 8.34 Effect of tyre inflation pressure on cornering
reaction force at ground level known as the cornering force. As the cornering force centre of pressure
is to the rear of the geometric centre of the wheel
and the side force acts perpendicularly through the
centre of the wheel hub, the offset between the
these two forces, known as the pneumatic trail,
causes a moment (couple) about the geometric
Fig. 8.35 Illustration of self-aligning torque
Fig. 8.36
8.4.9 Camber thrust (Figs 8.37 and 8.38)
The tilt of the wheel from the vertical is known as
the camber. When it leans inwards towards the
turning centre it is considered to be negative and
when the top of the wheel leans away from the
turning centre it is positive (Fig. 8.37). A positive
camber reduces the cornering force for a given slip
angle relative to that achieved with zero camber but
negative camber raises it.
Constructing a vector triangle of forces with the
known vertical reaction force and the camber inclination angle, and projecting a horizontal component
perpendicular to the reaction vector so that it intersects the camber inclination vector, enables the
magnitude of the horizontal component, known
as camber thrust, to be determined (Fig. 8.37).
The camber thrust can also be calculated as the
product of the reaction force and the tangent of
the camber angle.
Variation of self-aligning torque with cornering
The total lateral force reaction acting on the tyre
is equal to the sum of the cornering force and
camber thrust.
Static steering torque, that is the torque needed
to rotate the steering when the wheels are not rolling, has nothing to do with the generated selfaligning torque when the vehicle is moving. The
heavy static steering torque experienced when the
vehicle is stationary is due to the distortion of
the tyre casing and the friction created between
the tyre tread elements being dragged around
the wheels' point of pivot at ground level. With
radial ply tyres the more evenly distributed tyre
to ground pressure over the contact patch
makes manoeuvring the steering harder than
with cross-ply tyres when the wheels are virtually
Fig. 8.37
Camber thrust ˆ Wheel reaction tan i:e:
F ˆ Fc Ft
F ˆ total lateral force
Fc ˆ cornering force
Ft ˆ camber thrust
When both forces are acting in the same direction, that is with the wheel tilting towards the
centre of the turn, the positive sign should be
used, if the wheel tilts outwards the negative sign
applies (Fig. 8.38).
Thus negative camber increases the lateral reaction to side forces and positive camber reduces it.
Illustrating positive and negative camber and camber thrust
edge will try to speed up while the larger inner edge
will tend to slow down relative to the speed in the
middle of the tread. As a result, the tread portion in
the outer tread region will slip forward, the portion
of tread near the inner edge will slip backwards and
only in the centre of tread will true rolling be
To minimize tyre wear due to camber scrub modern suspensions usually keep the wheel camber
below 11¤2 degrees. Running wheels with a slight
negative camber on bends reduces scrub and
improves tyre grip whereas positive camber increases
tread scrub and reduces tyre to road grip.
8.4.10 Camber scrub (Fig. 8.39)
When a wheel is inclined to the vertical it becomes
cambered and a projection line drawn through the
wheel axis will intersect the ground at some point.
Thus if the wheel completes one revolution a cone
will be generated about its axis with the wheel and
tyre forming its base.
If a vehicle with cambered wheels is held on a
straight course each wheel tread will advance along
a straight path. The distance moved along the road
will correspond to the effective rolling radius at the
mid-point of tyre contact with the road (Fig. 8.39).
The outer edge of the tread (near the apex) will have
a smaller turning circumference than the inner edge
(away from apex). Accordingly, the smaller outer
8.4.11 Camber steer (Fig. 8.40)
When a vehicle's wheels are inclined (cambered) to
the vertical, the rolling radius is shorter on one side
of the tread than on the other. The tyre then forms
part of a cone and tries to rotate about its apex
(Fig. 8.40(a and b)). Over a certain angular motion
of the wheel, a point on the larger side of the tyre
will move further than a point on the smaller side of
the tyre and this causes the wheel to deviate from
Intake passage
(c) Turning right
Return passage
Return passage
Intake passage
(One of two)
Fig. 9.19 contd
When the steering wheel is turned, the tyre to
ground reaction on the front road wheels causes the
torsion bar to twist according to the torque applied
on the steering column shaft. Therefore the relative
angular movement of the worm shaft to that of the
input shaft increases in proportion to the input
torque at the steering wheel, so that the shuttle
valves will both be displaced an equal amount
from the mid-neutral position. As soon as the steering wheel effort is released, the elastic torsion bar
ensures that the two shuttle valves return to the
neutral or mid position. The function of these shuttle valves is to transfer fluid under pressure, in
accordance to the steering input torque, from
the pump delivery port to one or other end of the
integral power cylinder whilst fluid from the
opposite end of the cylinder is released and
returned to the reservoir.
Operation of control valve and power piston
Neutral position (Fig. 9.19(a)) Fluid from the
pump flows into and around an annular chamber
surrounding the worm head in a plane similar to
that of the shuttle valves where it acts on the
exposed end faces of the shuttle valve pistons.
With the shuttle valves in the neutral position,
fluid moves through the intake passages on the
right hand end of the shuttle valve pistons, to the
two annular grooves on the periphery of the worm
head. Fluid then passes from the worm head annular grooves to the left hand side of the power piston
via the horizontal long passage and sector chamber, and to the right hand piston face directly by
way of the short passage. From the worm head
grooves fluid will also flow into the shuttle valve
return grooves, over each return groove land which
is aligned with the exit groove, to the middle
waisted region of the shuttle valve and into the
torsion bar and input shaft chamber. Finally fluid
moves out from the return pipe back to the pump
displacement. Therefore the pronged input will
rotate clockwise to the worm head.
With a clockwise movement of the input shaft
relative to the worm head, the upper shuttle valve
piston moves to the right and the lower shuttle
valve piston moves to the left. Consequently, the
upper shuttle valve opens both the intake and
return passages but the lower shuttle valve closes
both the intake and return passages.
Under these conditions fluid flows from the
pump to the annular space around the worm head
in the plane of the shuttle valves. It then enters the
upper valve intake, fills the annular valve space and
passes around the left hand worm head groove.
Finally, fluid flows through the short horizontal
passage into the right hand side of the power cylinder where, in proportion to the pressure build-up, it
forces the piston to the left. Accordingly the meshing rack and sector teeth compel the sector shaft to
rotate anticlockwise.
At the same time as the fluid expands the right
hand side of the power cylinder, the left hand side
of the power cylinder will contract so that fluid will
be displaced through the long horizontal passage to
the worm head right hand annular groove. Fluid
then flows back to the reservoir via the upper shuttle valve return groove and land, through to the
torsion bar and input shaft chamber and finally
back to the reservoir.
Turning left (anticlockwise rotation) (Fig. 9.19(b))
An anticlockwise rotation of the steering wheel
against the front wheel to ground opposing resistance distorts the torsion bar as input torque is
transferred to the worm shaft via the torsion bar.
The twisting of the torsion bar means that the
worm shaft also rotates anticlockwise, but its angular movement will be less than the input shaft displacement. As a result, the prongs of the input shaft
shift the upper and lower shuttle valves to the left
and right respectively. Accordingly this movement
closes both the intake and return passages of the
upper shuttle valve and at the same time opens
both the intake and return passages of the lower
shuttle valve.
Fluid can now flow from the pump into the worm
head annular space made in the outer housing. It
then passes from the lower shuttle valve intake to the
right hand worm head annular groove. The transfer
of fluid is complete when it enters the left hand
power cylinder via the sector shaft. The amount of
power assistance is a function of the pressure buildup against the left side of the piston, which corresponds to the extent of the shuttle valve intake passage opening caused by the relative angular
movements of both the input shaft and worm shaft.
Movement of the power piston to the right displaces fluid from the right hand side of the power
cylinder, where it flows via the worm head annular
groove to the lower shuttle valve return passage to
the central torsion bar and input shaft chamber.
It then flows back to the reservoir via the flexible
return pipe.
9.2.4 Power assisted steering lock limiters
(Fig. 9.20(a and b))
Steering lock limiters are provided on power
assisted steering employed on heavy duty vehicles
to prevent excessive strain being imposed on the
steering linkage, the front axle beam and stub axles
and the supporting springs when steering full lock
is approached. It also protects the hydraulic components such as the pump and the power cylinder
assembly from very high peak pressures which
could cause damage to piston and valve seals.
Power assisted steering long stem conical valve lock
limiter The lock limiters consist of a pair of conical valves with extended probe stems located in the
sector shaft end cover (Fig. 9.20(a and b)). Each
valve is made to operate when the angular movement of the sector shaft approaches either steering
lock, at which point a cam profile machined on the
end of the sector shaft pushes open one or other of
the limiting valves. Opening one of the limiter
valves releases the hydraulic pressure in the power
cylinder end which is supplying the assistance; the
Turning right (clockwise rotation) (Fig. 9.19(c))
Rotating the steering wheel in a clockwise direction
applies a torque via the torsion bar to the worm
in proportion to the tyre to ground reaction and
the input effort. Due to the applied torque, the
torsion bar twists so that the angular movement
of the worm shaft lags behind the input shaft
Right hand
valve closed
Left hand
valve open
Piston and nut
(a) Turning left
Left hand
Right hand
valve open
(b) Turning right
Fig. 9.20 (a and b)
Power assisted steering long stem conical valve lock limiter
excess fluid is then permitted to flow back to the
reservoir via the control housing.
the steering lock movement is increased, the piston
approaches the end of its stroke until the right hand
ball valve contacts the worm shaft stop pin, thereby
forcing the ball off its seat. The hydraulic pressure
existing on the left side of the piston, which has
already opened the left hand side ball valve, is
immediately permitted to escape through the clearance formed between the internal bore of the nut
and the worm shaft. Fluid will now flow along the
return passage leading to the control reaction valve
and from there it will be returned to the reservoir.
The release of the fluid pressure on the right side of
the piston therefore prevents any further hydraulic
power assistance and any further steering wheel
rotation will be entirely manual.
Turning left (anticlockwise steering rotation)
(Fig. 9.20(a)) Rotation of the input shaft anticlockwise applies both manual and hydraulic effort
onto the combined power piston and nut of the
steering box so that it moves to the right within
the cylinder. Just before the steering reaches full
lock, one of the sector cam faces contacts the corresponding valve stem and pushes the conical valve
off its seat. Pressurized fluid will immediately
escape past the open valve through to the return
chamber in the control valve housing, where it
flows back to the reservoir. Therefore, any further
rotation of the sector shaft will be entirely achieved
by a considerable rise in manual effort at the steering wheel, this being a warning to the driver that
the steering has reached maximum lock.
Turning right (clockwise steering rotation)
(Fig. 9.21(b)) Rotation of the steering box input
shaft clockwise screws the worm out from the piston and nut. This shifts the shuttle valve pistons so
that the hydraulic pressure rises on the right hand
end of the piston. Towards the end of the left hand
stroke of the piston, the ball valve facing the blind
end of the cylinder contacts the adjustable stop pin.
Hydraulic pressure will now force the fluid from
the high pressure end chamber to pass between the
worm and the bore of the nut to open the right
hand ball valve and to escape through the left hand
ball valve into the sector gear chamber. The fluid
then continues to flow along the return passage
going to the control reaction valve and from there
it is returned to the reservoir. The circulation of
fluid from the pump through the piston and back
to the reservoir prevents further pressure build-up so
that the steering gearbox will only operate in the
manual mode. Hence the driver is made aware that
the road wheels have been turned to their safe full
lock limit.
Turning right (clockwise steering rotation)
(Fig. 9.20(b)) Rotation of the steering box input
shaft clockwise screws the worm out from the piston
and nut and actuates the control valve so that
hydraulic pressure builds up on the right hand end
of the piston. As the sector shaft rotation approaches
maximum lock, the sector cam meets the valve stem,
presses open the valve against the valve return spring
tension and causes the hydraulic pressure in the right
hand cylinder chamber to drop. The excess fluid will
now flow back to the reservoir via the right hand end
annular chamber in the control valve housing. The
driver will immediately experience a considerable
increase in manual effort at the steering wheel, indicating that the road wheels have been rotated to near
enough maximum lock.
Power assisted steering double ball valve lock
limiter This lock limiter consists of a simple double ball valve located in the blank end of the integral piston and nut. To control the stroke of the
piston an adjustable stop pin is mounted in the
enclosed end of the power cylinder housing, while
the right hand piston movement is limited by the
stop pin mounted in the end of the worm shaft.
9.2.5 Roller type hydraulic pump
(Fig. 9.22(a and b))
The components of this pump (Fig. 9.22 (a and b))
consist of the stationary casing, cam ring and the
flow and pressure control valve. The moving parts
comprise of a rotor carrier mounted on the drive
shaft and six rollers which lodge between taper
slots machined around the rotor blank. The drive
shaft itself is supported in two lead-bronze bushes,
one of which is held in the body and the other in the
end cover. A ball bearing at the drive end of the
shaft takes the load if it is belt and pulley driven.
The rotor carrier is made from silicon manganese
steel which is heat treated to a moderate hardness.
Turning left (anticlockwise steering rotation)
(Fig. 9.21(a)) Rotation of the steering input
shaft anticlockwise causes both manual and
hydraulic effort to act on the combined power
piston and nut, moving it towards the right. As
The rotor slots which guide the rollers taper in width
towards their base, but their axes instead of being
radial have an appreciable trailing angle so as to
provide better control over the radial movement of
the rollers. The hollow rollers made of case- hardened steel are roughly 10 mm in diameter and there
are three standard roller lengths of 13, 18 and 23 mm
to accommodate three different capacity pumps.
The cam ring is subjected to a combined rolling
and sliding action of the rollers under the generated
pressure. To minimize wear it is made from heat
treated nickel-chromium cast iron. The internal
profile of the cam ring is not truly cylindrical, but
is made up from a number of arcs which are shaped
to maximize the induction of delivery of the fluid as
it circulates through the pump.
To improve the fluid intake and discharge flow
there are two elongated intake ports and two similar discharge ports at different radii from the shaft
axes. The inner ports fill or discharge the space
between the rollers and the bottoms of their slots
and the outer ports feed or deliver fluid in the space
Fig. 9.21 (a and b)
formed between the internal cam ring face and the
lobes of the rotor carrier. The inner elongated
intake port has a narrow parallel trailing (transition) groove at one end and a tapered leading
(timing) groove at the other end. The inner discharge port has only a tapered trailing (timing)
groove at one end. These secondary circumferential
groove extensions to the main inner ports provide
a progressive fluid intake and discharge action as
they are either sealed or exposed by the rotor
carrier lobes and thereby reduce shock and noise
which would result if these ports were suddenly
opened or closed, particularly if air has become
trapped in the rotor carrier slots.
Operating cycle of roller pump (Fig. 9.22(a and b))
Rotation of the drive shaft immediately causes the
centrifugal force acting on the rollers to move them
outwards into contact with the internal face of the
cam ring. The functioning of the pump can be
considered by the various phases of operation as
Power assisted steering double ball valve lock limit
Fig. 9.22 (a and b)
Power assisted roller type pump and control valve unit
an individual roller moves around the internal cam
face through positions A, B, C, D, E and F.
the pressure in these chambers will drop and thus
induce fluid from the intake passages to enter by
way of the outer chamber formed by the rotor lobe
and the cam face and by the inner port into the
tapered roller slot region. Filling the two regions of
the chamber separately considerably speeds up the
fluid intake process.
Filling phase (Fig. 9.22(a)) As the roller in position A moves to position B and then to position C,
the space between the eccentric mounted rotor
carrier lobe and cam face increases. Therefore the
volume created between adjacent rollers will also
become greater. The maximum chamber volume
occurs between positions C and D. As a result,
Pressurization phase (Fig. 9.22(a)) With further
rotation of the rotor carrier, the leading edge of the
rotor slot just beyond position C is just on the point
of closing the intake ports, and the space formed
between adjacent rollers at positions C and D starts
to decrease. The squeezing action pressurizes the
the plunger valve are supplied with pressurized
fluid from the pump. Situated in the passage
which joins the two end chambers of the plunger
is a calibrated flow orifice. The end chamber which
houses the plunger return spring is downstream of
the flow orifice.
Fluid from the pump discharge ports moves
along a passage leading into the reduced diameter
portion of the flow control plunger (Fig. 9.22(a)).
This fluid circulates the annular space surrounding
the lower part of the plunger and then passes along
a right angled passage through a calibrated flow
orifice. Here some of the fluid is diverted to the
flow control plunger spring chamber, but the
majority of the fluid continues to flow to the outlet
port of the pump unit, where it then goes through a
flexible pipe to the control valve built into the
steering box (pinion) assembly. When the engine
is running, fluid will be pumped from the discharge
ports to the flow control valve through the calibrated flow orifice to the steering box control
valve. It is returned to the reservoir and then finally
passed on again to the pump's intake ports.
Discharge phase (Fig. 9.22(a)) Just beyond
roller position D the inner discharge port is uncovered by the trailing edge of the rotor carrier slot.
This immediately enables fluid to be pushed out
through the inner discharge port. As the rotor continues to rotate, the roller moves from position D
to E with a further decrease in radial chamber
space so that there is a further rise in fluid pressure.
Eventually the roller moves from position E to F.
This uncovers the outer discharge port so that an
increased amount of fluid is discharged into the
outlet passage.
Transition phase (Fig. 9.22(a)) The roller will
have completed one revolution as it moves from
position F to the starting position at A. During the
early part of this movement the leading edge of the
rotor slot position F closes both of the discharge
ports and at about the same time the trailing edge
of the rotor slot position A uncovers the transition
groove in readiness for the next filling phase. The
radial space between the rotor lobe and internal
cam face in this phase will be at a minimum.
Principle of the flow orifice (Fig. 9.22(a and b))
With low engine speed (Fig. 9.22(a)), the calibrated
orifice does not cause any restriction or apparent
resistance to the flow of fluid. Therefore the fluid
pressure on both sides of the orifice will be similar,
that is P1.
As the pump speed is raised (Fig. 9.22(b)), the
quantity of fluid discharged from the pump in a
given time also rises, this being sensed by the flow
orifice which cannot now cope with the increased
amount of fluid passing through. Thus the orifice
becomes a restriction to fluid flow, with the result
that a slight rise in pressure occurs on the intake
side of the orifice and a corresponding reduction in
pressure takes place on the outlet side. The net
outcome will be a pressure drop of P1±P2, which
will now exist across the orifice. This pressure differential will become greater as the rate of fluid
circulation increases and is therefore a measure of
the quantity of fluid moving through the system in
unit time.
Flow and pressure control valves
Description of the flow and pressure control valve
unit (Fig. 9.22(a and b)) The quantity of fluid
discharged from the roller type pump and the
build-up in fluid pressure both increase almost
directly with rising pump rotor speed. These characteristics do not meet the power assisted steering
requirements when manoeuvring at low speed since
under these conditions the fluid circulation is
restricted and a rise in fluid pressure is demanded
to operate the power cylinders double acting piston. At high engine and vehicle speed when driving
straight ahead, very little power assistance is
needed and it would be wasteful for the pump to
generate high fluid pressures and to circulate large
amounts of fluid throughout the hydraulic system.
To overcome the power assisted steering mismatch
of fluid flow rate and pressure build-up, a combined flow control and pressure relief valve unit is
incorporated within the cast iron pump housing.
The flow control valve consists of a spring loaded
plunger type valve and within the plunger body is
a ball and spring pressure relief valve. Both ends of
Operation of the flow control valve (Fig. 9.22
(a and b)) When the pump is running slowly the
pressure drop across the flow orifice is very small so
that the plunger control spring stiffness is sufficient
to fully push the plunger down onto the valve cap
stop (Fig. 9.22(a)). However, with rising pump
speed the flow rate (velocity) of the fluid increases
and so does the pressure difference between both
sides of the orifice. The lower pressure P2 on the
output side of the orifice will be applied against the
plunger crown in the control spring chamber,
whereas the higher fluid pressure P1 will act underneath the plunger against the annular shoulder area
and on the blanked off stem area of the plunger.
Eventually, as the flow rate rises and the pressure
difference becomes more pronounced, the hydraulic pressure acting on the lower part of the plunger
P1 will produce an upthrust which equals the
downthrust of the control spring and the fluid
pressure P2. Consequently any further increase in
both fluid velocity and pressure difference will
cause the flow control plunger to move back progressively against the control spring until the shouldered edge of the plunger uncovers the bypass port
(Fig. 9.22(b)). Fluid will now easily return to the
intake side of the pump instead of having to work
its tortuous way around the complete hydraulic
system. Thus the greater the potential output of
the pump due to its speed of operation the further
back the plunger will move and more fluid will be
bypassed and returned to the intake side of the
pump. This means in effect that the flow output
of the pump will be controlled and limited irrespective of the pump speed (Fig. 9.23). The maximum
output characteristics of the pump are therefore
controlled by two factors; the control spring stiffness and the flow orifice size.
is machined on the large diameter portion of the
plunger just above the shoulder. A radial relief hole
connects this groove to the central spring housing.
With this arrangement the ball relief valve is
subjected to the pump output pressure on the
downstream (output) side of the flow orifice.
If the fluid output pressure exceeds some predetermined maximum, the ball will be dislodged
from its seat, permitting fluid to escape from the
control spring chamber, through the centre of the
plunger and then out by way of the radial hole and
annular groove in the plunger body. This fluid is
then returned to the intake side of the pump via the
bypass port.
Immediately this happens, the pressure P2 in the
control spring chamber drops, so that the increased
pressure difference between both ends of the flow
control plunger pushes back the plunger. As a
result the bypass port will be uncovered, irrespective of the existing flow control conditions, so that a
rapid pressure relief by way of the flow control
plunger shoulder edge is obtained. It is the ball
valve which senses any peak pressure fluctuation
but it is the flow control valve which actually provides the relief passage for the excess of fluid. Once
the ball valve closes, the pressure difference across
the flow orifice for a given flow rate is again established so that the flow control valve will revert back
to its normal flow limiting function.
9.2.6 Fault diagnosis procedure
Pump output check (Figs 9.12, 9.13, 9.15 and 9.18)
Operation of the pressure relief valve (Fig. 9.22
(a and b)) The pressure relief valve is a small
ball and spring valve housed at one end and inside
the plunger type flow control valve at the control
spring chamber end (Fig. 9.22(a)). An annular groove
1 Disconnect the inlet hose which supplies fluid
pressure from the pump to the control (reaction)
valve, preferably at the control valve end.
2 Connect the inlet hose to the pressure gauge end
of the combined pressure gauge and shut-off
valve tester and then complete the hydraulic circuit by joining the shut-off valve hose to the
control valve.
3 Top up the reservoir if necessary.
4 Read the maximum pressure indicated on type
rating plate of pump or manufacturer's data.
5 Start the engine and allow it to idle with the shutoff valve in the open position.
6 Close the shut-off valve and observe the maximum pressure reached within a maximum time
span of 10 seconds. Do not exceed 10 seconds,
otherwise the internal components of the
pump will be overworked and will heat up
excessively with the result that the pump will
be damaged.
Fig. 9.23 Typical roller pump flow output and power
consumption characteristics
7 The permissible deviation from the rated pressure may be 10%. If the pump output is low,
the pump is at fault whereas if the difference is
higher, check the functioning of the flow and
pressure control valves.
Excessive free-play in the steering If when turning
the driving steering wheel, the play before the steering road wheels taking up the response is excessive
check the following;
1 worn steering track rod and drag link ball joints
if fitted,
2 worn reaction control valve ball pin and cups,
3 loose reaction control valve location sleeve.
An average maximum pressure figure cannot be
given as this will depend upon the type and application of the power assistant steering. A typical
value for maximum pressure may range from
45 bar for a ram type power unit to anything up
to 120 bar or even more with an integral power
unit and steering box used on a heavy commercial
Heavy steering Heavy steering is experienced over
the whole steering from lock to lock, whereas binding is normally only experienced over a portion of
the front wheel steering movement. If the steering is
heavy, inspect the following items:
1 External inspection Ð Check reservoir level and
hose connections for leakage. Check for fan belt
slippage or sheared pulley woodruff key and
adjust or renew if necessary.
2 Pump output Ð Check pump output for low
pressure. If pressure is below recommended maximum inspect pressure and flow control valves
and their respective springs. If valve's assembly
appears to be in good condition dismantle pump,
examine and renew parts as necessary.
3 Control valve Ð If pump output is up to the
manufacturer's specification dismantle the control valve. Examine the control valve spool or
rotor and their respective bore. Deep scoring or
scratches will allow internal leaks and cause
heavy steering. Worn or damaged seals will also
cause internal leakage.
4 Power cylinder Ð If the control valve assembly
appears to be in good condition, the trouble is
possibly due to excessive leakage in the power
cylinder. If there is excessive internal power
cylinder leakage, the inner tube and power piston
ring may have to be renewed.
Power cylinder performance check (Figs 9.12, 9.13,
9.15 and 9.18)
1 Connect the combined pressure gauge and shutoff valve tester between the pump and control
valve as under pump output check.
2 Open shut-off valve, start and idle the engine and
turn the steering from lock to lock to bleed out
any trapped air.
3 Turn the steering onto left hand full lock. Hold
the steering on full lock and check pressure reading which should be within 10% of the pump
output pressure.
4 Turn the steering onto the opposite lock and
again check the pump output pressure.
5 If the pressure difference between the pump output and the power cylinder on both locks is
greater than 10% then the power cylinder is at
fault and should be removed for inspection.
6 If the pressure is low on one lock only, this
indicates that the reaction control valve is not
fully closing in one direction.
A possible cause of uneven pressure is that the
control valve is not centralizing or that there is an
internal fault in the valve assembly.
Noisy operation
the following:
To identify source of noise, check
1 Reservoir fluid level Ð Check the fluid level as a
low level will permit air to be drawn into the
system which then will cause the control valve
and power cylinder to become noisy while operating.
2 Power unit Ð Worn pump components will
cause noisy operation. Therefore dismantle and
examine internal parts for wear or damage.
3 If the reservoir and pump are separately located,
check the hose supply from the reservoir to
pump for a blockage as this condition will
cause air to be drawn into the system.
Binding check A sticking or binding steering
action when the steering is moved through a portion of a lock could be due to the following:
a) Binding of steering joint ball joints or control
valve ball joint due to lack of lubrication.
Inspect all steering joints for seizure and replace
where necessary.
b) Binding of spool or rotary type control valve.
Remove and inspect for burrs wear and
Steering chatter If the steering vibrates or chatters check the following:
socket formed around the neck of the ball pin. The
other half socket which bears against the ball end
of the ball pin is generally made from oil impregnated sintered iron (Fig. 9.24(c)); another type
designed for automatic chassis lubrication, an
induction hardened pressed steel half socket, is
employed (Fig. 9.24(d)). Both cases are spring loaded
to ensure positive contact with the ball at all times.
A helical (slot) groove machined across the shoulder
of the ball ensures that the housing half socket
and ball top face is always adequately lubricated
and at the same time provides a bypass passage to
prevent pressurization within the joint.
1 power piston rod anchorage may be worn or
requires adjustment,
2 power cylinder mounting may be loose or incorrectly attached.
9.3 Steering linkage ball and socket joints
All steering linkage layouts are comprised of rods
and arms joined together by ball joints. The ball
joints enable track rods, drag-link rods and relay
rods to swivel in both the horizontal and vertical
planes relative to the steering arms to which they
are attached. Most ball joints are designed to tilt
from the perpendicular through an inclined angle
of up to 20 for the axle beam type front suspension, and as much as 30 in certain independent
front suspension steering systems.
Ball and socket joints for light and medium
duty To reduce the risk of binding or seizure
and to improve the smooth movement of the ball
when it swivels, particularly if the dust cover is
damaged and the joint becomes dry, non-metallic
sockets are preferable. These may be made from
moulded nylon and for some applications the
nylon may be impregnated with molybdenum disulphide. Polyurethane and Teflon have also been
utilized as a socket material to some extent. With
the nylon sockets (Fig. 9.24(e)) the ball pin throat
half socket and the retainer cap is a press fit in the
bore of the housing end float. The coil spring
accommodates initial settling of the nylon and subsequent wear and the retainer cap is held in position by spinning over a lip on the housing. To
prevent the spring loaded half socket from rotating
with the ball, two shallow tongues on the insert half
socket engage with slots in the floating half socket.
These ball joints are suitable for light and medium
duty and for normal road working conditions have
an exceptionally longer service life.
For a more precise adjustment of the ball and
socket joint, the end half socket may be positioned
by a threaded retainer cap (Fig. 9.24(f)) which is
screwed against the ball until all the play has been
taken up. The cap is then locked in position by
crimping the entrance of the ball bore. A Belleville
spring is positioned between the half socket and
the screw retainer cap to preload the joint and
compress the nylon.
9.3.1 Description of ball joint (Fig. 9.24(a±f))
The basic ball joint is comprised of a ball mounted
in a socket housing. The ball pin profile can be
divided into three sections; at one end the pin is
parallel and threaded, the middle section is tapered
and the opposite end section is spherically shaped.
The tapered middle section of the pin fits into a
similarly shaped hole made at one end of the steering arm so that when the pin is drawn into the hole
by the threaded nut the pin becomes wedged.
The spherical end of the ball is sandwiched
between two half hemispherical socket sets which
may be positioned at right angles to the pin's axis
(Fig. 9.24(a and b)). Alternatively, a more popular
arrangement is to have the two half sockets located
axially to the ball pin's axis, that is, one above the
other (Fig. 9.24(c±f)).
The ball pins are made from steel which when
heat treated provide an exceptionally strong tough
core with a glass hard surface finish. These properties are achieved for normal manual steering applications from forged case-hardened carbon (0.15%)
manganese (0.8%) steel, or for heavy duty power
steering durability from forged induction hardened
3% nickel 1% chromium steel. For the socket housing which might also form one of the half socket
seats, forged induction hardened steels such as a
0.35% carbon manganese 1.5% steel can be used. A
1.2% nickel 0.5% chromium steel can be used for
medium and heavy heavy duty applications.
9.3.3 Ball joint dust cover (Fig. 9.24(c±f))
An important feature for a ball type joint is its dust
cover, often referred to as the boot or rubber gaiter,
but usually made from either polyurethane or
nitrile rubber mouldings, since both these materials
have a high resistance to attack by ozone and do
not tend to crack or to become hard and brittle at
low temperature. The purpose of the dust cover is
9.3.2 Ball joint sockets (Fig. 9.24(c±f))
Modern medium and heavy duty ball and socket
joints may use the ball housing itself as the half
Fig. 9.24 (a±f)
Steering ball unit
to exclude road dirt moisture and water, which if
permitted to enter the joint would embed itself
between the ball and socket rubbing surfaces. The
consequence of moisture entering the working section of the joint is that when the air temperature
drops the moisture condenses and floods the upper
part of the joint. If salt products and grit are
sprayed up from the road, corrosion and a mild
grinding action might result which could quickly
erode the glass finish of the ball and socket surfaces. This is then followed by the pitting of the
spherical surfaces and a wear rate which will
rapidly increase as the clearance between the rubbing faces becomes larger.
Slackness within the ball joint will cause wheel
oscillation (shimmy), lack of steering response,
excessive tyre wear and harsh or notchy steering feel.
Alternatively, the combination of grease, grit,
water and salts may produce a solid compound
which is liable to seize or at least stiffen the relative
angular movement of the ball and socket joint,
resulting in steering wander.
The dust boot must give complete protection
against exposure from the road but not so good
that air and the old grease cannot be expelled when
the joint is recharged, particularly if the grease is
pumped into the joint at high pressure, otherwise
the boot will burst or it may be forced off its seat so
that the ball and socket will become exposed to the
The angular rotation of the ball joint, which
might amount to 40 or even more, must be accommodated. Therefore, to permit relative rotation to
take place between the ball pin and the dust cover,
the boot makes a loose fit over the ball pin and is
restrained from moving axially by the steering arm
and ball pin shoulder while a steel ring is moulded
into the dust cover to prevent the mouth of the boot
around the pin spreading out (Fig. 9.24(c±f)). In
contrast, the dust cover makes a tight fit over the
large diameter socket housing by a steel band which
tightly grips the boot.
demand for more positive and reliable steering,
joint lubrication and the inconvenience of periodic
off the road time, automatic chassis lubrication
systems via plastic pipes have become very popular
for heavy commercial vehicles so that a slow but
steady displacement of grease through the ball joint
system takes place. The introduction to split socket
mouldings made from non-metallic materials has
enabled a range of light and medium duty ball and
socket joints to be developed so that they are grease
packed for life. They therefore require no further
lubrication provided that the boot cover is a good
fit over the socket housing and it does not become
damaged in any way.
9.4 Steering geometry and wheel alignment
9.4.1 Wheel track alignment using Dunlop
optical measurement equipment Ð calibration of
alignment gauges
1 Fit contact prods onto vertical arms at approximately centre hub height.
2 Place each gauge against the wheel and adjust
prods to contact the wheel rim on either side of
the centre hub.
3 Place both mirror and view box gauges on a level
floor (Fig. 9.25(b)) opposite each other so that
corresponding contact prods align and touch
each other. If necessary adjust the horizontal
distance between prods so that opposing prods
are in alignment.
4 Adjust both the mirror and target plate on the
viewbox to the vertical position until the reflection of the target plate in the mirror is visible
through the periscope tube.
5 Look into the periscope and swing the indicator
pointer until the view box hairline is positioned
in the centre of the triangle between the two thick
vertical lines on the target plate.
6 If the toe-in or -out scale hairline does not align
with the zero reading on the scale, slacken off the
two holding down screws and adjust indicator
pointer until the hairline has been centred.
Finally retighten screws.
9.3.4 Ball joint lubrication
Before dust covers were fitted, ball joints needed to
be greased at least every 1600 kilometres (1000
miles). The advent of dust covers to protect the
joint against dirt and water enabled the grease
recharging intervals to be extended to 160 000 kilometres (10 000 miles). With further improvements
in socket materials, ball joint design and the choice
of lubricant the intervals between greasing can be
extended up to 50 000 kilometres (30 000 miles)
under normal road working conditions. With the
Toe-in or -out check (Fig. 9.25(a, b and c))
1 Ensure that tyre pressures are correct and that
wheel bearings and track rod ends are in good
2 Drive or push the vehicle in the forward
direction on a level surface and stop. Only take
readings with the vehicle rolled forward and
never backwards as the latter will give a false
toe angle reading.
3 With a piece of chalk mark one tyre at ground
4 Place the mirror gauge against the left hand
wheel and the view box gauge against the right
hand wheel (Fig. 9.25(b)).
5 Push each gauge firmly against the wheels so that
the prods contact the wheel on the smooth surface of the rim behind the flanged turnover since
the edge of the latter may be slightly distorted
due to the wheel scraping the kerb when the
vehicle has been parked. Sometimes gauges
may be held against the wheel rim with the aid
of rubber bands which are hooked over the tyres.
6 Observe through the periscope tube the target
image. Swing the indicator pointer to and fro
over the scale until the hairline in the view box
coincides with the centre triangle located
between the thick vertical lines on the target
plate which is reflected in the mirror.
7 Read off the toe-in or -out angle scale in degrees
and minutes where the hairline aligns with the
8 Check the toe-in or -out in two more positions by
pushing the vehicle forward in stages of a third of
a wheel revolution observed by the chalk mark on
the wheel. Repeat steps 4 to 7 in each case and
record the average of the three toe angle readings.
9 Set the pointer on the dial calculator to the
wheel rim diameter and read off the toe-in
Fig. 9.25 (a±c) Wheel track alignment using the Dunlop equipment
or -out in millimetres opposite the toe angle
reading obtained on the toe-in or -out scale.
Alternatively, use Table 9.1 to convert the toe-in
or -out angle to millimetres.
10 If the track alignment is outside the manufacturer's recommendation, slacken the track
rod locking bolts or nuts and screw the track
rods in or out until the correct wheel alignment
is achieved. Recheck the track toe angle
when the track rod locking devices have been
4 Push a measuring gauge over each wheel clamp
stub shaft and tighten thumbscrews. This should
not prevent the measuring gauge rotating
independently to the wheel clamp.
5 Attach the elastic cord between the uncoloured
hole in the rotor of each measuring gauge.
6 Wheel lateral run-out is compensated by the following procedure of steps 7±10.
7 Lift the front of the vehicle until the wheels clear
the ground and place a block underneath one of
the wheels (in the case of front wheel drive vehicles) to prevent it from rotating.
8 Position both measuring gauges horizontally
and hold the measuring gauge opposite the
blocked wheel. Slowly rotate the wheel one complete revolution and observe the measuring
gauge reading which will move to and fro and
record the extreme of the pointer movement on
the scale. Make sure that the elastic cord does
not touch any part of the vehicle or jack.
9 Further rotate wheel in the same direction
until the mid-position of the wheel rim lateral
run-out is obtained, then chalk the tyre at the
12 o'clock position.
9.4.2 Wheel track alignment using Churchill line
cord measurement equipment
Calibration of alignment gauges
(Fig. 9.26(a))
1 Clamp the centre of the calibration bar in a vice.
2 Attach an alignment gauge onto each end of the
calibration bar.
3 Using the spirit bubble gauge, level both of the
measuring gauges and tighten the clamping
4 Attach the elastic (rubber) calibration cord
between adjacent uncoloured holes formed in
each rotor.
5 Adjust measuring scale by slackening the two
wing nuts positioned beneath each measuring
scale, then move the scale until the zero line
aligns exactly with the red hairline on the pointer
lens. Carefully retighten the wing nuts so as not
to move the scale.
6 Detach the calibration cord from the rotors and
remove the measuring gauges from calibration
Table 9.1 Conversion of degrees to millimetres
Rim size
Toe-in or -out check (front or rear wheels)
(Fig. 9.26(a))
1 Position a wheel clamp against one of the front
wheels so that two of the threaded contact studs
mounted on the lower clamp arm rest inside the
rim flange in the lower half of the wheel. For
aluminium wheels change screw studs for claw
studs provided in the kit.
2 Rotate the tee handle on the centre adjustment
screw until the top screw studs mounted on the
upper clamp arm contact the inside rim flange in
the upper half of the wheel. Fully tighten centre
adjustment screw tee handle to secure clamp to
3 Repeat steps 1 and 2 for opposite side front
Fig. 9.26 (a±c) Wheel track alignment using the Churchill equipment
10 Repeat steps 7 to 9 for the opposite side front
11 Position each front wheel with the chalk mark
at 12 o'clock.
12 Utilize the brake pedal depressor to prevent the
wheels from rotating.
13 Slide a turntable underneath each front wheel,
remove the locking pins and then lower the
front wheels onto both turntables.
14 Bounce the front of the vehicle so that the
suspension quickly settles down to its normal
Tilt each measuring gauge to the horizontal
position by observing when the spirit level
bubble is in the mid-position. Lock the measuring gauge in the horizontal position with
the second thumbscrew.
16(a) Observe the left and right toe angle reading
and add them together to give the combined
toe angle of the front wheels.
(b) Alternatively, turn one road wheel until its
measuring gauge pointer reads zero, then
read the combined toe angle on the opposite
side measuring gauge (front wheels only).
17 Using Table 9.1 provided, convert the toe angle
into track toe-in or -out in millimetres and compare with the manufacturer's recommendations.
misaligned, the vehicle will move forward in a
skewed line relative to the axis of symmetry. This
second directional line is known as the thrust axis
or driving axis. The angle between the axis of symmetry and the thrust axis is referred to as the thrust
axis deviation which will cause the front and rear
wheels to be laterally offset to each other when the
vehicle moves in the straight ahead direction.
If the vehicle has been constructed and
assembled correctly the thrust axis will coincide
with the axis of symmetry, but variation in the
rear wheel toe angles relative to the axis of symmetry will cause the vehicle to be steered by the rear
wheels. As a result, the vehicle will tend to move in
a forward direction and partially in a sideway
direction. The vehicle will therefore tend to pull
or steer to one side and when driving round a
bend the steering will oversteer on one lock and
understeer on the opposite lock. In the case of
Fig. 9.27(a), with a right hand lateral offset the
vehicle will understeer on left hand bends and oversteer on right hand turns.
The self-steer effect of the rear wheels due to track
or axle misalignment will conflict with the suspension geometry such as the kingpin inclination and
castor which will therefore attempt to direct the
vehicle along the axis of symmetry. Consequently,
the tyres will be subjected to excessive scrub.
Thrust axis deviation may be produced by body
damage displacing the rear suspension mounting
points, rear suspension worn bushes, poorly
located leaf springs and distorted or incorrectly
assembled suspension members.
Toe-out on turns check (Fig. 9.26(b and c))
1 After completing the toe-in or -out check, keep
the wheel clamp and measuring gauge assembly
attached to each front wheel. Maintain the midwheel lateral run-out position with the tyre at the
12 o'clock position and ensure the brake pedal
depressor is still applied.
2 Transfer the position of the elastic cord hook
ends attached to the measuring gauge rotors
from the uncoloured holes to the red holes which
are pitched 15 relative to the uncoloured holes.
3 Rotate the right hand (offside) wheel in the direction the arrow points on the measuring gauge
facing the red hole in which the cord is hooked
until the scale reads zero. At this point the right
hand wheel (which becomes the outer wheel on
the vehicle turning circle) will have been pivoted
15 . Make sure that the cord does not touch any
4 Observe the reading on the opposite left hand
(near side) measuring gauge scale, which is the
toe-out turns angle for the left hand (near side)
wheel (the inner wheel on the vehicle's turning
5 Change the cord to the blue holes in each measuring gauge rotor.
6 Rotate the left hand (near side) wheel in the
direction the arrow points on the measuring
gauge facing the blue hole until the hairline
pointer on the left hand measuring gauge reads
7 Read the opposite right hand (offside) wheel
measuring gauge scale which gives the toe-out
on turns for the right hand (offside) wheel (the
inner wheel on the vehicle's turning circle).
8 Compare the left and right hand toe-out turns
readings which should be within one degree of
one another.
Front to rear alignment check using Churchill line
cord measurement equipment (Fig. 9.28(a and b))
1 Check rear wheel toe angle by using the procedure
adopted for front wheel toe angle measurement
(Fig. 9.26(a)). Use the convention that toe-in is
positive and toe-out is negative.
2 Keep the wheel clamp and measuring gauge
assembly on both rear wheels.
3 Attach a second pair of wheel clamps to both
front wheels.
4 Remove the rear wheel toe elastic cord from the
two measuring gauges.
5 Hook a front to rear alignment elastic cord
between the stub shaft deep outer groove of the
front wheel clamps and the single hole in the
measuring gauge rotors set at 90 from the
middle hole of the three closely spaced holes
(Fig. 9.28(a and b)).
9.4.3 Front to rear wheel misalignment
(Fig. 9.27(a))
An imaginary line projected longitudinally between
the centre of the front and rear wheel tracks is
known as the vehicle's centre line or the axis of
symmetry (Fig. 9.27(a)). If the vehicle's body and
suspension alignment is correct, the vehicle will
travel in the same direction as the axis of symmetry.
When the wheel axles at the front and rear are
Fig. 9.27 (a and b)
Front to rear wheel alignment procedure
6 Apply a slight tension to the front to rear
alignment cord using the metal plate adjusters.
7 With all four wheels pointing in the straight
ahead direction, read and record the left and
right hand measuring gauge scales (Fig. 9.27
(a and b)). To determine the thrust axis deviation
(TAD) angle subtract the left reading from the
right reading and divide the difference of the
reading by two.
Thrust axis deviation (TAD) angle ˆ
where R ˆ Right hand measuring gauge reading
From lateral offset Table 9.2, a thrust axis deviation of 300 for a wheel base of 3000 mm is equivalent to a lateral offset to the right of 22 mm when
the vehicle moves in a forward direction.
Example 2 (Fig. 9.27(b)) Determine the rear
wheel toe-in or -out and the front to rear lateral
offset of a vehicle having independent rear suspension from the following data:
= 3400 mm
Wheel diameter
= 13 inches
Left rear wheel toe angle
reading Ð out (out=negative)
= 400
Right rear wheel toe angle
reading Ð in (in=positive)
= ‡150
Left side front to rear measuring
gauge reading Ð out
= 550
Right side front to rear measuring
gauge reading Ð in
= ‡250
L ˆ Left hand measuring gauge reading
8 The lateral offset can be approximately determined from the formula
Lateral offset=Wheel base tan
however, the makers of the equipment supply
Table 9.2 to simplify the conversion from thrust
axis deviation angle to lateral offset.
a) Toe-in or -out:
Rear wheel combined toe angle ˆ ( 400 ) ‡
(‡150 ) ˆ 250
From toe conversion table a toe angle of 250
for a 13 inch diameter wheel is equivalent to
a toe-out of 2.65 mm.
Example 1 (Fig. 9.27(b)) Determine the rear
wheel toe-in or -out and the front to rear lateral
offset for a 3000 mm wheelbase vehicle having a
rigid rear axle and 13 inch diameter wheels from
the following information:
Left rear wheel toe angle reading
Right rear wheel toe angle reading
Left side front to rear measurement
reading Ð out (out=negative)
Right side front to rear measurement
reading Ð in (in=positive)
b) Lateral offset:
Thrust axis deviation R
(TAD) angle
‡250 ( 550 )
‡25 ‡ 550
ˆ 400
From lateral offset Table 9.2, a thrust axis deviation of 400 for a wheelbase of 3400 mm is equivalent
to a lateral offset to the right of 33.5 mm when the
vehicle is moving in the forward direction.
= 00
= 00
= ‡300
a) Toe-in or -out:
Rear wheel combined toe angle ˆ 00 ‡ 00 ˆ 00
Thus wheels are parallel.
b) Lateral offset:
Thrust axis deviation ˆ
300 )
‡300 ‡ 300
ˆ 300
Note A minus minus makes a plus;
9.4.4 Six wheel vehicle with tandem rear axle
steering geometry (Fig. 9.28)
For any number of road wheels on a vehicle to
achieve true rolling when cornering, all projected
lines drawn through each wheel axis must intersect
at one common point on the inside track, this being
the instantaneous centre about which the vehicle
travels. In the case of a tandem rear axle arrangement in which the axles are situated parallel to each
( )ˆ‡
Table 9.2 Lateral offset tables
Lateral offset of front wheels in relation to rear wheels
(Measurements in millimeters)
(Measurements in inches)
ft in
Fig. 9.28 Six wheel tandem rear axle vehicle steering
so that the imaginary extended lines drawn through
both rear axles would eventually meet. Unfortunately lines drawn through the front steered stub
axles and the rear skewed axles may not all meet at
one point. Nevertheless, they may almost merge so
that very near true rolling can occur for a large
proportion of the steering angle when the vehicle is
in motion. The remainder of the rear axle wheel
misalignment is absorbed by suspension spring
distortion, shackle joints or torque arm rubber
joints, and tyre compliance or as undesirable tyre
A negative ( ) value indicates front wheels offset to left
A positive ( ‡ ) value indicates front wheels offset to right
other, lines projected through the axles would
never meet and in theory true rolling cannot exist.
However, an approximate instantaneous centre for
the steered vehicle can be found by projecting a line
mid-way and parallel between both rear axles, this
being assumed to be the common axis of rotation
(Fig. 9.30). Extended lines passing through both
front wheel stub axles, if made to intersect at one
point somewhere along the common projected single rear axle line, will then produce very near true
rolling condition as predicted by the Ackermann
Improvements in rear axle suspension design
have introduced some degree of roll steer which
minimizes tyre scrub on the tandem axle wheels.
This is achieved by the camber of the leaf springs
supporting the rear axles changing as the body rolls
so that both rear axles tend to skew in the plan view
9.4.5 Dual front axle steering
Operating large rigid trucks with heavy payloads
makes it necessary in addition to utilizing tandem
axles at the rear to have two axles in the front of the
vehicle which share out and support the load.
Both of the front axles are compelled to be steer
axles and therefore need to incorporate steering
linkage such as will produce true or near true
rolling when the vehicle is driven on a curved
The advantages gained by using dual front steering axles as opposed to a single steer axle are as
1 The static payload is reduced per axle so that
static and dynamic stresses imposed on each
axle assembly are considerably lessened.
2 Road wheel holding is improved with four
steered wheels as opposed to two, particularly
over rough ground.
3 Road wheel impact shocks and the subsequent
vibrations produced will be considerably
reduced as the suspension for both sets of wheels
share out the vertical movement of each axle.
4 Damage to one axle assembly or a puncture to
one of the tyres will not prevent the vehicle being
safely steered to a standstill.
5 Tyre wear rate is considerably reduced for dual
axle wheels compared to single axle arrangements for similar payloads. Because the second
axle wheels have a smaller turning angle relative
to the foremost axle wheels, the tyre wear is
normally less with the second axle road wheels.
Fig. 9.30 Dual front axle steering linkage layout with
power assistance
A major disadvantage with dual front axles is
that it is unlikely in practice that both instantaneous centres of the first and second stub axle turning circles will actually intersect at one point for all
angles of turn. Therefore tyre scrub may be excessive
for certain angles of steering wheel rotation.
if the vehicle is to be able to negotiate a turning
For a dual front axle vehicle to be steered, the
Ackermann principle must apply to each of the
front axles. This means that each axle has two
wheels pivoted at each end of its beam. To enable
true rolling of the wheels to take place when the
vehicle is travelling along a curved track, lines
drawn through each of these four stub axles must
intersect at a common centre of rotation, somewhere along the extended line drawn between the
tandem rear axles (Fig. 9.29).
Because the wheelbase between the first front
axle is longer than the second front axle, relative
to the mid-tandem axle position, the turning angles
of both first front wheels will be greater than those
of the second front axle wheels. The correct angle
difference between the inner and outer wheels of
each axle is obtained with identical Ackermann
linkage settings, whereas the angle differential
between the first and second axles is formed by
the connecting rod ball joint coupling location on
both relay drop arms being at different distances
from their respective pivot point.
The dual steering linkage with power assistance
ram usually utilizes a pair of swing relay drop arms
bolted onto the chassis side member with their free
ends attached to each axle drag link (Fig. 9.30).
The input work done to operate the steering is
mainly supplied by the power cylinder which is
coupled by a ball joint to the steering gearbox
drop arm at the front and the power piston rod is
anchored through a ball joint and bracket to the
Dual front axle steering geometry (Fig. 9.29)
When a pair of axles are used to support the front
half of a vehicle each of these axles must be steered
Fig. 9.29
Dual front axle steering Ackermann geometry
chassis at the rear end. To transfer the driver's
input effort and power assistance effort to both
steer axles, a forward connecting rod links the
front end of the power cylinder to the first relay
drop arm. A second relay connecting rod then joins
both relay arms together.
A greater swivel movement of the first pair of
stub axles compared to the second is achieved (Fig.
9.31) by having the swing drop arm effective length
of the first relay AB shorter than the second relay
arm A0 B0 . Therefore the second relay arm push or
pull movement will be less than the input swing of
the first relay arm. As a result, the angular swing of
the first relay, ˆ 20 , will be less than for the
second relay arm angular displacement, 0 ˆ 14 .
2 Assemble mirror gauge stand with the mirror
positioned at right angles to the tubular stand.
Position the mirror gauge against a rear axle
wheel (preferably the nearest axle to the front)
with the mirror facing towards the front of the
vehicle (Fig. 9.32(b)).
3 Place the view box gauge stand on the floor in a
transverse position at least one metre in front of
the vehicle so that the view box faces the mirror
(Fig. 9.32(c)). Move the view box stand across until
the reflected image is centred in the view box with a
zero reading on the scale. Chalk mark the position
of the view box tripod legs on the ground.
4 Bring the mirror gauge stand forward to the first
steer axle wheel and place gauge prods against
wheel rim (Fig. 9.34(c)).
5 If both pairs of steer axle wheels are set parallel
(without toe-in or -out), set the pointer on the toe
angle scale to zero. Conversely, if both steer axles
have toe-in or -out settings, move the pointer on
the toe angle scale to read half the toe-in or -out
Dual front axle alignment checks using Dunlop
optical measurement equipment (Fig. 9.32(a±d))
1 Roll or drive forward. Check the toe-in or -out of
both pairs of front steering wheels and adjust
track rods if necessary (Fig. 9.32(a)).
Fig. 9.31 Dual front axle steering interconnecting relay linkage principle
Fig. 9.32 Dual front axle wheel alignment procedure
figure, i.e. with a track toe angle of 300 set the
pointer to read 150 .
6 Look through the periscope tube and with an
assistant move the driver's steering wheel in the
cab until the hairline is central in the view box
(Fig. 9.32(c)). At the same time make sure that
the mirror gauge prods are still in contact with
the front wheel rim. The first front steer axle is
now aligned to the first rigid rear axle.
7 Move the mirror gauge from the first steer axle
wheel back to the second steer axle wheel and
position the prods firmly against the wheel rim
(Fig. 9.32(d)).
8 Look into the periscope. The hairline in the
view box should be centrally positioned with
the toe angle pointer still in the same position
as used when checking the first steer axle (Fig.
If the hairline is off-centre, the relay connecting
rod between the two relay idler arms should be
adjusted until the second steer axle alignment relative to the rear rigid axle and the first steer axle has
been corrected. Whilst carrying out any adjustment
to the track rods or relay connecting rod, the overall wheel alignment may have been disturbed.
Therefore a final check should be made by repeating all steps from 2 to 8.
9.4.6 Steer angle dependent four wheel steering
system (Honda)
This steer angle dependent four wheel steering system provides dual steering characteristics enabling
same direction steer to take place for small steering
wheel angles. This then changes to opposite direction steer with increased steering wheel deviation
from the straight ahead position. Both of these
steer characteristics are explained as follows:
Fig. 9.33 Front and rear wheel steer relationship to
driver's steering wheel angular movement
Opposite direction steer (Fig. 9.33) At low speed
and large steering wheel angles the rear wheels are
turned by a small amount in the opposite direction
to the front wheels to improve manoeuvrability
when parking (Fig. 9.32). In effect opposite direction steer reduces the car's turning circle but it does
have one drawback; the rear wheels tend to bear
against the side of the kerb. Generally there is
sufficient tyre sidewall distortion and suspension
compliance to accommodate the wheel movement
as it comes into contact with the kerb so that only
at very large steering wheel angles can opposite
direction steer becomes a serious problem.
turned a small amount in the same direction as
the front wheels to improve both steering response
and stability at speed (Fig. 9.33). This feature is
particularly effective when changing lanes on
motorways. Incorporating a same direction steer
to the rear wheels introduces an understeer characteristic to the car because it counteracts the angular steering movement of the front wheels and
consequently produces a stabilizing influence in
the high speed handling of the car.
Front and rear road wheel response relative to the
steering wheel angular movement (Fig. 9.33) Moving the steering wheel approximately 120 from its
central position twists the front wheels 8 from the
Same direction steer (Fig. 9.33) At high speed and
small steering wheel angles the rear wheels are
straight ahead position. Correspondingly, it moves
the eccentric shaft peg to its maximum horizontal
annular gear offset, this being equivalent to a maximum 1.5 same direction steer for the rear road
wheels (Fig. 9.33).
Increasing the steering wheel rotation to 232
turns the front wheels 15.6 from the straight
ahead position which brings the planetary peg
towards the top of the annular gear and in vertical
alignment with the gear's centre. This then corresponds to moving the rear wheels back to the
straight ahead position (Fig. 9.33).
Further rotation of the steering wheel from the
straight ahead position orbits the planetary gear
over the right hand side of the annular gear.
Accordingly the rear wheels steadily move to the
opposite direction steer condition up to a maximum of 5.3 when the driver's steering wheel has
been turned roughly 450 (Fig. 9.33).
verts the angular input shaft motion to a transverse
linear movement. This is then conveyed to the rear
wheel swivels by the stroke rod and split track rods.
Rear steering box construction (Fig. 9.35) The
rear steering box is basically formed from an epicyclic gear set consisting of a fixed internally
toothed annular ring gear in which a planetary
gear driven by an eccentric shaft revolves (Fig.
9.35). Motion is transferred from the input eccentric
shaft to the planetary gear through an offset peg
attached to a disc which is mounted centrally on
the eccentric shaft. Rotation of the input eccentric
shaft imparts movement to the planetary gear which
is forced to orbit around the inside of the annular
gear. At the same time, motion is conveyed to the
guide fork via a second peg mounted eccentrically
on the face of the planetary gear and a slider plate
which fits over the peg (Fig. 9.35). Since the slider
plate is located between the fork fingers, the rotation
of the planetary gear and peg causes the slider plate
to move in both a vertical and horizontal direction.
Due to the construction of the guide fork, the slider
plate is free to move vertically up and down but is
constrained in the horizontal direction so that the
stroke rod is compelled to move transversely to and
fro according to the angular position of the planetary gear and peg.
Adopting this combined epicyclic gear set with
a slider fork mechanism enables a small same direction steer movement of the rear wheels to take
place for small deviation of the steering wheel
from the straight ahead position. The rear wheels
then progressively change from a same direction
steer movement into an opposite steer displacement with larger steering angles.
The actual steering wheel movement at which
the rear wheels change over from the same direc-
Four wheel steer (FWS) layout (Fig. 9.34) The
steering system is comprised of a rack and pinion
front steering box and a rear epicyclic steering box
coupled together by a central drive shaft and a pair
of Hooke's universal end joints (Fig. 9.35). Both
front and rear wheels swivel on ball suspension
joints which are steered by split track rods actuating steering arms. The front road wheels are interlinked by a rack and transverse input movement to
the track rods is provided by the input pinion shaft
which is connected to the driver's steering wheel via
a split steering shaft and two universal joints. Steering wheel movement is relayed to the rear steering
box by way of the front steering rack which meshes
with an output pinion shaft. This movement of the
front rack causes the output pinion and centre
drive shaft to transmit motion to the rear steering
box. The rear steering box mechanism then con-
Fig. 9.34 Four wheel steering (4WS) system
Fig. 9.35
Epicyclic rear steering box
180 eccentric shaft peg rotation Rotating the
eccentric shaft through its second quadrant
(90±180 ) causes the planetary gear to roll anticlockwise inside the annular gear so that it moves
with the eccentric peg to the highest position. At the
same time, the planetary peg orbits to the right
hand side of the annular gear centre line
(Fig. 9.36(c)) so that the rear road wheels turn to
the opposite direction steer condition.
tion steer to the opposite direction steer and the
magnitude of the rear wheel turning angles relative
to both conditions are dependent upon the epicyclic gear set gear ratio chosen.
Rear steering box operation (Fig. 9.36(a±e)) The
automatic correction of the rear road wheels from
a same direction steer to opposite direction steer
with increasing front road wheel turning angle and
vice versa is explained by Fig. 9.36(a±e).
Central position With the steering wheel in the
straight ahead position, the planetary gear sits at
the bottom of the annular gear with both eccentric
shaft and planetary pegs located at bottom dead
centre in the mid-position (Fig. 9.36(a)).
270 eccentric shaft peg rotation Rotating the
eccentric shaft through a third quadrant (180±270 )
moves the planetary gears and the eccentric shaft peg
to the 270 position, causing the planetary peg to
orbit even more to the right hand side (Fig. 9.36(d)).
Consequently further opposite direction steer will be
90 eccentric shaft peg rotation Rotating the
eccentric shaft through its first quadrant (0±90 )
in a chosen direction from the bottom dead centre
position compels the planetary gear to roll in an
anticlockwise direction up the left hand side of
the annular ring gear. This causes the planetary
peg and the stroke rod to be displaced slightly
to the left (Fig. 9.36(b)) and accordingly makes
the rear wheels move to a same direction steer
360 eccentric shaft peg rotation Rotating
the eccentric shaft through a fourth quadrant
(270±360 ) completes one revolution of the
eccentric shaft. It therefore brings the planetary
gear back to the base of the annular ring gear
with the eccentric shaft peg in its lowest position
(Fig. 9.36(e)). The planetary peg will have moved
back to the central position, but this time the peg is
in its highest position. The front to rear wheel
steering drive gearing is normally so arranged that
Fig. 9.36 (a±d)
Principle of rear steering box mechanism
the epicyclic gear set does not operate in the fourth
quadrant even under full steering lock conditions.
manoeuvring the vehicle can be made more direct
compared with a manual steering with a slightly
higher steering ratio. The use of a more direct
low steering ratio when the road wheels are being
turned on either lock is made possible by the servo
action of the hydraulic operated power cylinder
and piston which can easily overcome the extra
tyre scrub and swivel-pin inclination resisting
force. The variable-ratio rack is achieved by having
tooth profiles of different inclination along the
length of the rack, accordingly the pitch of the
teeth will also vary over the tooth span.
With racks designed for manual steering the
centre region of the rack has wide pitched teeth
with a 40 flank inclination, whereas the teeth on
either side of the centre region of the rack have
a closer pitch with a 20 flank inclination. Conversely, power assisted steering with variable-ratio
rack and pinion (see Fig. 9.37(c)) has narrow
pitch teeth with 20 flank inclination in the central region; the tooth profile then changes to a
wider pitch with 40 flank inclination away from
the central region of the rack for both steering
9.5 Variable-ratio rack and pinion
(Fig. 9.37(a±d))
Variable-ratio rack and pinion can be made to
improve both manual and power assisted steering
operating characteristics. For a manual rack and
pinion steering system it is desirable to have a
moderately high steering ratio to provide an almost
direct steering response while the steering wheel is
in the normally `central position' for straight ahead
driving and for very small steering wheel angular
correction movement. Conversely for parking
manoeuvres requiring a greater force to turn the
steering wheel on either lock, a more indirect lower
steering ratio is called for to reduce the steering
wheel turning effort. However, with power assisted
steering the situation is different; the steering wheel
response in the straight ahead driving position still
needs to be very slightly indirect with a relatively
high steering ratio, but with the power assistance
provided the off-centre steering response for
Pressure angle
Pressure angle
(b) Off-centre rack teeth
(a) Central rack teeth
pitch (p)
Wide pitch (P)
Wide pitch (P)
(c) Variable-ratio tooth rack
Movement ratio
Large p.c.d.
Large p.c.d.
more direct
180 120
Turning steering wheel to right
Turning steering wheel to left
Steering wheel and pinion rotation (deg)
(d) Rack and pinion movement ratio from lock to lock of the steering wheel
Fig. 9.37 (a±d)
Variable ratio rack and pinion steering suitable for power assisted steering
With variable-ratio rack and pinion involute
teeth the rack has straight sided teeth. The sides
of the teeth are normal to the line of action,
therefore, they are inclined to the vertical at the
pressure angle. If the rack has narrow pitch `p'
20 pressure-angle teeth, the pitch circle diameter
(2R) of the pinion will be small, that is, the point
of contact of the meshing teeth will be close to the
tip of the rack teeth (Fig. 9.37(a)), whereas with
wide pitched `P' 40 pressure-angle tooth contact
between teeth will be near the root of the rack
teeth (Fig. 9.37(b)) so its pitch circle diameter (2R)
will be larger.
The ratio of steering wheel radius to pinion pitch
circle radius (tooth contact radius) determines the
movement ratio. Thus the smaller the pitch circle
radius of the pinion for a given steering wheel size,
the greater will be the movement ratio (see Fig.
9.37(d)), that is, a smaller input effort will be
needed to steer the vehicle, but inversely, greater
will be the steering wheel movement relative to the
vehicle road wheel steer angle.
This design of rack and pinion tooth profile can
provide a movement-ratio variation of up to 35%
with the number of steering wheel turns limited to
2.8 from lock to lock.
9.6.2 Design and construction (Fig. 9.38(a±d))
The `ZF Servotronic' speed-sensitive power assisted
steering uses a conventional rotary control valve, with
the addition of a reaction-piston device which modifies the servo assistance to match the driving mode.
The piston and rotary control valve assembly
comprises a pinion shaft, valve rotor shaft with
six external longitudinal groove slots, valve sleeve
with six matching internal longitudinal groove
slots, torsion bar, reaction-piston device and an
electro-hydraulic transducer. The reaction-piston
device is supported between the rotary valve rotor
and valve sleeve, and guided internally by the valve
rotor via three axially arranged ball grooves and
externally guided by the valve sleeve through a
multi-ball helix thread.
The function of the reaction-piston device is to
modify the fluid flow gap formed between the valve
rotor and sleeve longitudinal groove control edges
for different vehicle driving conditions.
An electronic control unit microprocessor takes
in speed frequency signals from the electronic
speedometer, this information is then continuously
evaluated, computed and converted to an output
signal which is then transmitted to the hydraulic
transducer mounted on the rotary control valve
casing. The purpose of this transducer is to control
the amount of hydraulic pressure reaching the
reaction-piston device based on the information
supplied to the electronic control unit.
9.6 Speed sensitive rack and pinion power
assisted steering
9.6.1 Steering desirability
To meet all the steering requirements the rack and
pinion steering must be precise and direct under
normal driving conditions, to provide a sense of
feel at the steering wheel and for the steering wheel
to freely return to the straight ahead position after
the steering has been turned to one lock or the
other. The conventional power assisted steering
does not take into account the effort needed to
perform a steering function relative to the vehicle
speed, particularly it does not allow for the extra
effort needed to turn the road wheels when manoeuvring the vehicle for parking.
The `ZF Servotronic' power assisted steering is
designed to respond to vehicle speed requirements,
`not engine speed', thus it provides more steering
assistance when the vehicle is at a standstill or
moving very slowly than when travelling at speed;
at high speed the amount of steering assistance may
be tuned to be minimal, so that the steering
becomes almost direct as with a conventional manual steering system.
9.6.3 Operation of the rotary control valve and
power cylinder
Neutral position (Figs 9.38(a) and 9.39(a)) With
the steering wheel in its central free position, pressurized fluid from the pump enters the valve sleeve,
passes though the gaps formed between the longitudinal groove control edges of both sleeve and
rotor, then passes to both sides of the power cylinder. At the same time fluid will be expelled via
corresponding exit `sleeve/rotor groove' controledge gaps to return to the reservoir. The circulation
of the majority of fluid from the pump to the
reservoir via the control valve prevents any buildup of fluid pressure in the divided power cylinder,
and the equalization of the existing pressure on
both sides of the power piston neutralizes any
`servo' action.
Anticlockwise rotation of the steering wheel (turning left
Ð low speed) (Figs 9.38(b) and 9.39(b)) Rotating
Valve sleeve
Inner check valve
Outer check valve
piston (RP)
Outer orifice
valve (CO-V)
Inner orifice
Teflon ring
Power piston
Power cylinder
(a) Neutral position
Fig. 9.38 (a±d)
Speed sensitive rack and pinion power assisted steering with rotary reaction control valve
Inner check valve
Outer check valve
(b) Turning left
(low speed)
Fig. 9.38 contd
the steering wheel in an anticlockwise direction
twists the control valve rotor against the resistance
of the torsion bar until the corresponding leading
edges of the elongated groove in the valve rotor and
sleeve align. At this point the return path to the exit
port `4' is blocked by control edges `2' while fluid
from the pump enters port `1'; it then passes in
between the enlarged control-edge gaps to come
out of port `3', and finally it flows into the righthand power cylinder chamber.
Inner check valve
Outer check valve
(c) Turning left
(high speed)
Fig. 9.38
Inner check valve
Outer check valve
(d) Turning right
(high speed)
Fig. 9.38 contd
Conversely fluid from the left hand side power
cylinder chamber is pushed towards port `2'
where it is expelled via the enlarged trailing control-edge gap to the exit port `4', then is returned
to the reservoir. The greater the effort by the
driver to turn the steering wheel, the larger will be
the control-edge gap made between the valve
sleeve and rotor and greater will be the pressure
imposed on the right hand side of the power
Return long slot
Torsion bar
Supply short
Power cylinder
and piston
(a) Neutral position
Fig. 9.39 (a±c)
Rack and pinion power assisted steering sectional end views of rotary reaction control valve
When the vehicle is stationary or moving very
slowly and the steering wheel is turned to manoeuvre it into a parking space or to pull out from
a kerb, the electronic speedometer sends out its
minimal frequency signal to the electronic control
unit. This signal is processed and a corresponding
control current is transmitted to the electrohydraulic transducer. With very little vehicle movement, the control current will be at its maximum;
this closes the transducer valve thus preventing
fluid pressure from the pump reaching the reaction
valve piston device and for fluid flowing to and
through the cut-off valve. In effect, the speed sensitive rotary control valve under these conditions
now acts similarly to the conventional power
assisted steering; using only the basic rotary control valve, it therefore is able to exert relatively
more servo assistance.
Anticlockwise rotation of the steering wheel (turning
left Ð high speed) (Figs 9.38(c) and 9.39(b)) With
increasing vehicle speed the frequency of the electronic speedometer signal is received by the electronic control unit; it is then processed and converted
to a control current and relayed to the electrohydraulic transducer. The magnitude of this control current decreases with rising vehicle speed,
Torsion bar
Supply short slot
Return long slot
(b) Turning left – anticlockwise
rotation of the steering
Fig. 9.39 contd
correspondingly the electro-hydraulic transducer
valve progressively opens thus permitting fluid to
reach the reaction piston at a pressure determined
by the transducer-valve orifice opening. If the steering wheel is turned anticlockwise to the left (Fig
3.38(c)), the fluid from the pump enters radial
groove `5', passes along the upper longitudinal
groove to radial groove `7', where it circulates and
comes out at port `3' to supply the right hand side of
the power cylinder chamber with fluid.
Conversely, to allow the right hand side cylinder
chamber to expand, fluid will be pushed out from
the left hand side cylinder chamber; it then enters
port `2' and radial groove `6', passing through the
lower longitudinal groove and hollow core of the
rotor valve, finally returning to the reservoir via
port `4'. Fluid under pressure also flows from
radial groove `7' to the outer chamber check valve
to hold the ball valve firmly on its seat. With the
electro-hydraulic transducer open fluid under
pump pressure will now flow from radial grooves
`5' to the inner and outer reaction-piston device
orifices. Fluid passing though the inner orifice circulates around the reaction piston and then passes
to the inner reaction chamber check valve where it
pushes the ball off its seat. Fluid then escapes
through this open check valve back to the
reservoir by way of the radial groove `6' through
the centre of the valve rotor and out via port `4'.
At the same time fluid flows to the outer piston
(c) Turning right – clockwise
rotation of the steering
Fig. 9.39
reaction chamber and to the right hand side of
the outer check valve via the outer orifice, but
slightly higher fluid pressure from port `7' acting
on the opposite side of the outer check valve prevents the valve opening. However, the fluid pressure build-up in the outer piston reaction chamber
will tend to push the reaction piston to the left hand
side, consequently due to the pitch of the ballgroove helix, there will be a clockwise opposing
twist of the reaction piston which will be transmitted to the valve rotor shaft. Accordingly this
reaction counter twist will tend to reduce the fluid
gap made between the valve sleeve and rotor longitudinal control edges; it therefore brings about a
corresponding reaction in terms of fluid pressure
reaching the left hand side of the power piston and
likewise the amount of servo assistance.
In the high speed driving range the electrohydraulic transducer control current will be very
small or even nil; it therefore causes the transducer
valve to be fully open so that maximum fluid pressure will be applied to the outer reaction piston.
The resulting axial movement of the reaction piston will cause fluid to be displaced from the inner
reaction chamber through the open inner reaction
chamber check valve, to the reservoir via the radial
groove `6', lower longitudinal groove, hollow rotor
and finally the exit port `4'.
As a precaution to overloading the power steering, when the reaction piston fluid pressure reaches
its pre-determined upper limit, the cut-off valve
opens to relieve the pressure and to return surplus
fluid to the reservoir.
causing fluid to be returned to the reservoir via the
radial groove `7', lower elongated rotor groove, hollow rotor core and out via port `4'. At the same time
fluid flows to the inner chamber of the reaction
piston via its entrance orifice. Therefore, the pressure on the spring side of its respective ball check
valve remains higher thus preventing the ball valve
opening. Subsequently pressure builds up in the
inner chamber of the reaction piston, and therefore
causes the reaction piston to shift to the right hand
side; this results in an anticlockwise opposing twist
to the reaction piston due to the ball-groove helices.
Accordingly the reaction counter twist will reduce
the flow gap between corresponding longitudinal
grooves' control edges so that a reduced flow will
be imposed on the left hand side of the power cylinder. Correspondingly an equal quantity of fluid will
be displaced from the reaction piston outer chamber
which is then returned to the reservoir via the now
open outer check valve. Thus as the electro-hydraulic transducer valve progressively opens with respect
to vehicle speed, greater will be the fluid pressure
transmitted to the reaction piston inner chamber
and greater will be the tendency to reduce the flow
gap between the aligned sleeve and rotor valve control edges, hence the corresponding reduction in
hydro-servo assistance to the steering.
Clockwise rotation of the steering wheel (turning
right Ð low speed) (Fig. 9.39(c)) Rotation of the
steering wheel clockwise twists the control valve
against the resistance of the torsion bar until the
corresponding leading control edges of the elongated grooves in the valve rotor and sleeve are
aligned. When the leading groove control edges
align, the return path to the exit port `3' is blocked
while fluid from the pump enters port `1'; it then
passes inbetween the enlarged control-edge gap to
come out of port `2' and finally flows into the left
hand power cylinder chamber.
Conversely, fluid from the right hand side power
cylinder chamber is displaced towards port `3'where
it is expelled via the enlarged gap made between the
trailing control edges to the exit port `4'; the fluid
then returns to the reservoir. The greater the misalignment between the valve sleeve and rotor control
edges the greater will be the power assistance.
Clockwise rotation of the steering wheel (turning
right Ð high speed) (Figs 9.38(d) and 9.39(c))
With increased vehicle speed the electro-hydraulic
transducer valve commences to open thereby exposing
the reaction piston to fluid supply pressure.
If the steering wheel is turned clockwise to the
right (Fig. 9.38 (d)), the fluid from the pump enters
the radial groove `5', passes along the upper longitudinal grooves to radial groove `6' where it circulates and comes out at port `2' to supply the power
cylinder's left hand side chamber with fluid.
Correspondingly fluid will be displaced from
the power cylinder's right hand chamber back to
the reservoir via port `3' and groove `7', passing
through to the lower longitudinal groove and
hollow core of the rotor valve to come out at port
`4'; from here it is returned to the reservoir.
Fluid under pressure will also flow from radial
groove `6' to the reaction piston's outer chamber
check valve thereby keeping the ball valve in the
closed position. Simultaneously, with the electrohydraulic transducer open, fluid will flow from
radial groove `5' to the inner and outer reactionpiston orifices. Fluid under pressure will also pass
though the outer orifice, and circulates around the
reaction piston before passing to the reaction piston's outer chamber check valve; since the fluid
pressure on the spring side of the check valve ball
is much lower, the ball valve is forced to open thus
9.6.4 Characteristics of a speed sensitive power
steering system (Fig. 9.40)
Steering input effort characteristics relative to vehicle speed and servo pressure assistance are shown
in Fig. 9.40. These characteristics are derived from
the microprocessor electronic control unit which
receives signals from the electronic speedometer
and transmits a corresponding converted electric
current to the electro-hydraulic transducer valve
attached to the rotary control valve casing.
Accordingly, the amount the electro-hydraulic
transducer valve opens controls the degree of
fluid pressure reaction on the modified rotary control valve (Fig. 9.38(c)). As a result the amount of
power assistance given to the steering system at
different vehicle speeds can be made to match
more closely the driver's input to the vehicle's resistance to steer under varying driving conditions.
Referring to Fig. 9.40 at zero vehicle speed when
turning the steering, for as little an input steering
wheel torque of 2 Nm, the servo fluid pressure rises
to 40 bar and for only a further 1 Nm input rise
(3 Nm in total) the actuating pressure can reach 94
bar. For a vehicle speed of 20 km/h the rise in servo
pressure is less steep, thus for an input effort torque
of 2 Nm the actuating pressure has only risen to
80 Km/h
160 Km/h
0 Km/h
Fluid pressure (bar)
Steering wheel torque (Nm)
Fig. 9.40
Speed sensitive power steering steering wheel torque to servo fluid pressure characteristics for various road
about 14 bar and for an input of 3 Nm the pressure
just reaches 30 bar. With a higher vehicle speed of
80 km/h the servo pressure assistance is even less,
only reaching 10, 18 and 40 bar for an input torque
of 2, 3 and 6 Nm respectively; however, beyond an
input torque of 6 Nm the servo pressure rises very
steeply. Similarly for a vehicle speed of 160 km/h
the rise in servo pressure assistance for an input
torque rise ranging from 2 to 6 Nm only increases
from 6 to 17 bar respectively, again beyond this
input torque the servo pressure rises extremely
rapidly. These characteristics demonstrate that
there is considerable servo pressure assistance
when manoeuvring the vehicle at a standstill or
only moving slowly; conversely there is very little
assistance in the medium to upper speed range of a
vehicle, in fact the steering is almost operated without assistance unless a very high input torque is
applied to the steering wheel in an emergency.
to leak due to severe overloading of the steering
linkage when driving against and over stone kerbs
and when manoeuvring the car during parking in
confined spaces. The electric power assisted steering unit is relatively light, compact, reliable and
requires a maximum current supply of between 40
and 80 amperes when parking (depending on the
weight imposed on the front road wheels) and
does not consume engine power as is the case of
a hydraulic power assisted steering system which
does apply a relatively heavy load on the engine.
9.7.1 Description and construction (Fig. 9.41)
The essentials of a rack and pinion electric power
assisted steering comprises an input shaft attached
to the steering wheel via an intermediate shaft and
universal joint and a integral output shaft and
pinion which meshes directly with the steering
rack, see Fig. 9.41. A torsion bar mounted in the
centre of the hollow input shaft joins the input and
output shafts together and transfers the driver's
manual effort at the steering wheel to the pinion
output shaft. Electrical servo assistance is provided
by an electric motor which supplies the majority of
the steering torque to the output pinion shaft when
the car's steering is being manoeuvred. Torque is
transferred from the electric motor to the output
pinion shaft through a ball bearing supported
worm gear and a worm wheel mounted and
attached to the output pinion shaft.
9.7 Rack and pinion electric power assisted
The traditional hydraulic actuated power assisted
steering requires weighty high pressure equipment,
which incorporates an engine driven high pressure
pump, fluid reservoir and filter, reaction valve,
high pressure hoses, servo cylinder, piston, ram
and a suitable fluid. There is a tendency for fluid
(engine speed)
arm &
Peg & slot
Ball &
Commutator and
Fig. 9.41 Rack and pinion electric power assisted steering system
Relative angular misalignment between the input
and output shafts is measured by transforming this
angular movement into an axial linear movement
along the input shaft by means of a slide sleeve,
control ball, internal diagonal groove and a peg
and slot. The slide sleeve which fits over the input
shaft can move axially relative to the input shaft
and rotates with the output shaft due to the peg and
slot. Proportionate axial movement of the slide
sleeve to the misalignment of the input to the output shafts is achieved by the internal diagonally
formed groove in the slide sleeve and the control
ball held in the shoulder part of the input shaft.
Any axial slide-sleeve movement is registered by
the rotary potentiometer (variable resistor)
through the potentiometer arm and pin which is
located in the slide sleeve's external groove.
When the steering is initially turned against the
tyre to road surface grip resistance, the input torque
applied to the steering is transferred to the pinion
output shaft through the central torsion bar. The
torsional twist of the torsion bar, that is, the angular
misalignment of the input and output shafts, is proportional to the input effort at the steering wheel
before the servo electric motor responds and supplies the extra input torque to the pinion output
shaft to produce the desired amount of steering
turn by the front road wheels. Should the electric
servo assistance fail for any reason, then the steering
input effort will be entirely provided by the driver
though the torsion bar; under these conditions however the driver will experience a much heavier steering. A limit to the maximum torsion bar twist is
provided when protruding ridges formed on the
input and output shafts butt with each other.
An electronic control unit which is a microprocessor takes in information from various electrical
sensors and then translates this from a programmed map into the required steering assistance
to be delivered by the servo electric motor.
Mechanical power is supplied by a servo electric
motor which is able to change its polarity so that it
can rotate either in a clockwise or anticlockwise
direction as commanded by the direction of steering turn, the drive being transferred from the output pinion shaft via a warm gear and warm wheel.
The large gear reduction ratio provided with this
type of drive gearing enables the warm wheel to
rotate at a much reduced speed to that of the warm
gear and enables a relatively large torque to be
applied to the output pinion shaft with a moderately small electric motor.
Steering wheel torque is monitored in terms of
relative angular misalignment of the input and
output shafts by the slide-sleeve movement, this is
then converted into an electrical signal via the
interlinked rotary potentiometer sensor. Engine
and road speed sensors enable the electronic control unit to provide speed-sensitive assistance by
providing more assistance at low vehicle speed
when manoeuvring in a restricted space and to
reduce this assistance progressively with rising
speed so that the driver experiences a positive feel
to the steering wheel. Note the engine and vehicle
speeds are monitored by the tachometer and antilock brake sensors respectively.
9.7.2 Operating principle (Figs 9.42(a±c))
Neutral position (Fig. 9.42(b)) When the input
and output shafts are aligned as when the steering
wheel is in a neutral no turning effort position, the
control ball will be in the central position of the
diagonal control groove. Correspondingly the
potentiometer lever arm will be in the horizontal
position, with zero signal feed current to the electronic control unit and the power supply from the
electronic control unit to the servo electric motor
switched off. Note the potentiometer is calibrated
with the wiper arm in its mid-track position to
signal a zero feed current.
Clockwise right hand turn (Fig. 9.42(a)) When the
steering wheel is turned clockwise to give a right
hand turn, the input torque applied by the steering
wheel causes a relative angular misalignment
between the input and output shaft, this being proportional to the degree of effort the driver applies.
As a result the control ball rotates clockwise with the
input shaft relative to the output shaft, and since the
slide sleeve cannot rotate independently to the output pinion shaft due to the peg and slot, the flanks of
the diagonal groove are compelled to slide past the
stationary control ball, thus constraining the slide
sleeve to an axial upward movement only.
Accordingly the rotary potentiometer lever arm
will twist anticlockwise thereby causing the wiper
arm to brush over the wire or ceramic resistive
track. The change in resistance and current flow
signals to the electronic control unit that servo
assistance is required, being in proportion to the
amount the slide sleeve and rotary potentiometer
moves. Once the initial effort at the steering wheel
has been applied the torsional twist of the torsion
bar relaxes; this reduces the relative misalignment
of the input and output shafts so that the rotary
potentiometer lever arm moves to a reduced feed
Axial movement (up)
from neutral
Axial movement (down)
from neutral
lever arm
Wiper arm
(a) Clockwise
right hand
(b) Neutral
Fig. 9.42 (a±c) Operating principles for a rack and pinion electric power assisted steering
(c) Anticlockwise
left hand turn
current position or even to zero feed current position. At this point the electronic control unit
switched `off' the electrical supply to the servo
electric motor so that servo assistance via the
warm gear and warm wheel to output pinion shaft
comes to an abrupt end.
peg and slot only permits the slide sleeve to move
axially. The vertical downward displacement of the
sleeve is relayed to the rotary potentiometer lever
arm which will now partially rotate in a clockwise
direction; its wiper arm will therefore brush over
the resistive track, and an appropriate signal current will then be fed to the electronic control unit.
The servo electric motor is then switched on, and
thereby rotates the worm gear and in turn the worm
wheel but at much reduced speed (due to the very
large gear reduction ratio provided by a worm gear
and worm wheel) in an anticlockwise direction. As
the input torque effort by the driver on the steering
wheel is reduced almost to nil, the relative misalignment of the input and output shaft will likewise be
reduced; correspondingly the rotary potentiometer
wiper arm will move to its mid-resistance position
signalling zero current feed to the electronic control
unit; it therefore switches off and stops the servo
electric motor.
Anticlockwise left hand turn (Fig. 9.42(c)) When
the steering wheel is turned anticlockwise to negotiate a left hand turn, the input effect applied by the
driver to the steering wheel causes a relative angular misalignment between the input and output
shafts, the relative twist of the torsion bar being
proportional to the driver's input effort on the
steering wheel. Due to the rotary movement of
the input shaft, and control ball relative to the
pinion output shaft, the diagonal groove in the
sleeve will be forced to move over the stationary
control ball in a downward axial direction since the
10 Suspension
10.1 Suspension geometry
The stability and effective handling of a vehicle
depends upon the designers' selection of the
optimum steering and suspension geometry which
particularly includes the wheel camber, castor
and kingpin inclination. It is essential for the suspension members to maintain these settings
throughout their service life.
Unfortunately, the pivoting and swivelling joints
of the suspension system are subject to both wear
and damage and therefore must be checked periodically. With the understanding of the principles
of the suspension geometry and their measurements it is possible to diagnose and rectify steering
and suspension faults. Consideration will be given
to the terminology and fundamentals of suspension
construction and design.
10.1.2 Wheel camber angle (Figs 10.1 and 10.2)
Wheel camber is the lateral tilt or sideway inclination of the wheel relative to the vertical (Fig. 10.1).
When the top of the wheel leans inwards towards the
body the camber is said to be negative, conversely
an outward leaning wheel has positive camber.
Road wheels were originally positively cambered
to maintain the wheel perpendicular to the early
highly cambered roads (Fig. 10.2) and so shaped as
to facilitate the drainage of rain water. With modern underground drainage, road camber has been
greatly reduced or even eliminated and therefore
wheel camber has been reduced to something like
‰ to 1‰ degrees.
The axis of rotation of a cambered wheel if projected outwards will intersect the ground at the
apex of a cone generated if the wheel was permitted
to roll freely for one revolution. The wheel itself
then resembles the frustrum of a cone (Fig. 10.1).
The path taken by the cambered wheel (frustrum of
a cone) if free to roll would be a circle about the
apex. Consequently both front wheels will tend to
steer outwards in opposite directions as the vehicle
moves forwards. In practice, the track rods and ball
joints are therefore preloaded as they restrain the
wheels from swivelling away from each other
when the vehicle is in motion. If both wheels
have similar camber angles, their outward pull
on the track rods will be equal and therefore
balance out. If one wheel is slightly more cambered than the other, due maybe to body roll with
independent suspension or because of misalignment, the steering wheels will tend to wander or
pull to one side as the vehicle is steered in the
straight ahead position.
10.1.1 Suspension terminology
Swivel joints or king pins These are the points
about which the steering wheel stub axles pivot.
Pivot centre The point where the swivel ball joint
axis or kingpin axis projects and intersects the
Contact patch This is the flattened crown area of
a tyre which contacts the ground.
Contact centres This is the tyre contact patch
central point which is in contact with the ground.
Track This is the transverse distance between
both steering wheel contact centres.
Fig. 10.1 Wheel camber geometry
Fig. 10.2
10.1.3 Swivel or kingpin inclination
(Figs 10.3±10.7)
Swivel pin or kingpin inclination is the lateral
inward tilt (inclination) from the top between the
upper and lower swivel ball joints or the kingpin to
the vertical (Fig. 10.3). If the swivel ball or pin axis
is vertical (perpendicular) to the ground, its contact
centre on the ground would be offset to the centre
of the tyre contact patch (Fig. 10.4). The offset
between the pivot centre and contact patch centre
is equal to the radius (known as the scrub radius) of
a semicircular path followed by the rolling wheels
when being turned about their pivots. When turning the steering the offset scrub produces a torque
T created by the product of the offset radius r and
the opposing horizontal ground reaction force F
(i:e: T=Fr (Nm)). A large pivot to wheel contact
centre offset requires a big input torque to overcome the opposing ground reaction, therefore the
steering will tend to be heavy. No offset (zero offset
radius) (Fig. 10.5) prevents the tread rolling and
instead causes it to scrub as the wheel is steered so
that at low speed the steering also has a heavy
response. A compromise is usually made by offsetting the pivot and contact wheel centres to
roughly 10±25% of the tread width for a standard
sized tyre. This small offset permits the pivot axis
to remain within the contact patch, thereby
enabling a rolling movement to still take place
when the wheels are pivoted so that tyre scruff
and creep (slippage) are minimized. One other
Road camber
A negatively cambered wheel leaning towards
the radius of a curved track or bend increases its
cornering power and reduces the tyre contact patch
slip angle for a given cornering force compared to
a wheel rolling in an upright position. Conversely,
a positively cambered wheel leaning away from the
centre of rotation reduces its cornering power and
increases the tyre slip angle for a similar cornering
force compared to a wheel rolling perpendicular to
the ground.
To provide a small amount of understeer, the
front wheels are normally made to generate a
greater slip angle than the rear wheels by introducing positive wheel camber on the front wheels and
maintaining the rear wheels virtually perpendicular
to the ground.
When cornering with positive camber angles on
both front wheels, the inner and outer wheels will
lean inwards and outwards respectively relative to
the centre of rotation of the turn. At the same time,
body roll transfers weight from the inner wheel to
the outer one. As a result the inner wheel will generate less slip angle than the outer wheel because it
provides an inward leaning, more effective tyre grip
with less vertical load than that of the less effective
outward leaning tyre, which supports a greater proportion of the vehicle's weight. The front cambered
tyres will generate on average more slip angle than
the upright rear wheels and this causes the vehicle to
have an understeer cornering tendency.
Steered positive cambered wheels develop
slightly more slip angle than uncambered wheels.
When they are subjected to sudden crosswinds or
irregular road ridging, the tyres do not instantly
deviate from their steered path, with the result that
a more stable steering is achieved.
With the adoption of wider tyres as standard on
some cars, wheel camber has to be kept to a minimum to avoid excessive edge wear on the tyres
unless the suspension has been designed to cope
with the new generation of low profile wide tread
width tyre.
Fig. 10.3
Swivel (king) pin inclination
Fig. 10.4 Swivel (king) pin vertical axis offset
Fig. 10.5
effect of a large pivot to contact centre offset is
when one of the wheels hits an obstacle like a
bump or pothole in the road; a large opposing
twisting force would be created momentarily
which would be relayed back to the driver's steering
wheel in a twitching fashion.
To reduce or even eliminate pivot to wheel centre
offset, the whole stub axle, hub bearing assembly
and disc or drum would have to be positioned
within the centre region of the wheel rim and also
extend, and therefore protrude, beyond the wheel
rim flange (Fig. 10.5). A dished wheel arrangement
of this type is known as centre point steering
because both pivot centre and contact patch centres
coincide in the middle of the wheel.
The alternative and realistic way of reducing the
pivot to contact patch centre offset is to laterally
incline the axis of the swivel joints so that the whole
hub assembly and disc or drum is positioned inside
the wheel and only the upper swivel joint may
protrude outside the wheel rim.
The consequences of tilting the swivel pin axis is
the proportional lowering of the stub axle axis in
the horizontal plane as the wheel assembly swivels
about its pivot points relative to the straight ahead
position (Fig. 10.6(a and b)). Because the road
wheels are already supported at ground level, the
reverse happens, that is, both upper and lower
wishbone arms or axle beam which supports
the vehicle body are slightly raised. This unstable
state produces a downward vehicle weight component which tends to return both steered wheel
assemblies to a more stable straight ahead position.
In other words, the pivot inclination produces a
self-centring action which is independent of vehicle
speed or traction but is dependent upon the weight
concentration on the swivel joints and their inclination. A very large swivel ball or pin inclination
produces an excessively strong self-centring effect
which tends to kick back on turns so that the swivel
ball or pin inclination angle is usually set between
5 and 15 . A typical and popular value would be
something like 8 or 12 .
The combination of both camber and swivel
joint inclination is known as the included angle
and the intersection of both of these axes at one
point at ground level classifies this geometry as
centre point steering (Fig. 10.7). In practice, these
centre lines projected through the ball joints or pins
and through the centre of the wheel are made to
meet at some point below ground level. Thus an
offset exists between the projected lines at ground
level, which produces a small twisting movement
when the wheels are steered. As a result, the wheels
tend to roll about a circular path with the offset as
its radius, rather than twist about its swivel centre
with a continuous slip-grip action which occurs
when there is no offset as with the centre point
steering geometry.
Dished wheel centre point steering
10.1.4 Castor angle (Figs 10.8 and 10.9)
The inclination of the swivel ball joint axis or kingpin axis in the fore and aft direction, so that the tyre
contact centre is either behind or in front of the
imaginary pivot centre produced to the ground, is
known as the castor angle (Fig. 10.8(b and c)).
Positive castor angle is established when the wheel
contact centre trails behind the pivot point at
Fig. 10.6 (a and b)
Swivel and kingpin inclination self-straightening tendency
force of the front tyres on the road causes both
tyres to move until they are in a position where
no out of balance force exists, that is, positioned
directly to the rear of the pivot swivel balls or pin
With front wheel drive vehicles the situation is
different because the driving torque is transmitted
through the steered front wheels (Fig. 10.9(b)). By
inclining the pivot axis forwards, a negative castor
is produced and instead of the pivot axis being
pushed by the rear wheel drive thrust, traction is
now transmitted through the front wheels so that
the pivot axis is pulled forwards. The swivel balls or
pin mounting swing to the rear of the contact patch
centre, due to the vehicle rolling resistance acting
through the rear wheels, opposing any forward
The effects of castor angle can be seen in Fig.
10.9(a and b), when the steering is partially turned
on one lock. The trail or lead distance between the
pivot centre and contact patch centre rotates as the
steered wheels are turned so that the forward driving force FD and the equal but opposite ground
reaction FR are still parallel but are now offset by
a distance x. Therefore a couple (twisting movement) M is generated of magnitude M ˆ Fx, where
F ˆ FD ˆ FR . With the vehicle in motion, the couple M will continuously try to reduce itself to zero
by eliminating the offset x. In other words, the
driving and reaction forces FD and FR are at all
Fig. 10.7 Camber and swivel pin inclination centre point
ground level (Fig. 10.8(b)). Negative castor angle
exists if the wheel contact centre leads the pivot axis
intersection at ground level (Fig. 10.8(c)).
If the pivot centre and wheel contact patch centre
coincide the castor is nil (Fig. 10.8(a)). Under these
conditions the steered wheels become unstable as
they tend to twitch from side to side when the
vehicle travels along a straight path.
A rear wheel drive vehicle has the front wheel
steer pivot axis inclined backward to produce positive castor (Fig. 10.9(a)). As the vehicle is propelled
from the rear (the front wheels are pushed by the
driving thrust transmitted by the rear drive wheels),
it causes the front wheels to swing around their
pivot axis until the tyre contact centre trails directly
behind. This action takes place because the drag
Fig. 10.8 Castor angle steering geometry
Castor angle
torque (M)
Castor angle
torque (M)
(a) Rear wheel drive castor
angle self-righting torque effect
Fig. 10.9 (a and b)
(b) Front wheel drive castor
angle self-righting torque effect
Illustration of steered wheel castor self-straightening tendency
times tending to align themselves with the wheels
rolling when the steering has been turned to one
lock. As a result the trailing or leading offset x
produces a self-righting effect to the steered wheels.
The greater the angle the wheels have been steered,
the larger the pivot centre to contact patch centre
offset x and the greater the castor self-centring
action will be. The self-righting action which
tends to straighten out the steering after it has
been turned from the straight position, increases
with both wheel traction and vehicle speed.
10.1.5 Swivel joint positive and negative offset
(Figs 10.10±10.15)
When one of the front wheels slips during a brake
application, the inertia of the moving mass will
tend to swing the vehicle about the effective wheel
which is bringing about the retardation because
Fig. 10.11
Swivel pin inclination negative offset
Fig. 10.10 Swivel pin inclination positive offset
there is very little opposing resistance from the
wheel on the opposite side (Fig. 10.12).
If the offset of the swivel ball joints is on the
inside of the tyre contact patch the swivel inclination is known as positive offset (Fig. 10.10). When
the wheels are braked the positive offset distance
and the inertia force of the vehicle produce a turning movement which makes the wheels pivot about
the contact patch centre in an outward direction at
the front (Fig. 10.10). If the off side (right) wheel
moves onto a slippery patch, the vehicle will not
only veer to the left, due to the retarding effect of
the good braked wheel preventing the vehicle moving forward, but the near side (left) wheel will also
turn and steer to the left (Fig. 10.13). Therefore the
positive offset compounds the natural tendency for
the vehicle to swerve towards the left if the right
hand wheel skids instead of continuing on a stable
straight ahead path.
Arranging for the swivel ball joint inclination
centre line to intersect the ground on the outside of
the contact patch centre produces what is known as
negative offset (Fig. 10.11). With negative offset the
Fig. 10.12 Directional stability when one wheel skids
whilst being braked
Fig. 10.13 Directional stability with positive offset when
one wheel skids whilst being braked
Fig. 10.14 Directional stability with negative offset when
one wheel skids whilst being braked
momentum of the vehicle will produce a turning
moment that makes the wheels swivel inwards at
the front about the contact patch centre (Fig. 10.11)
because the swivel ball joints and stub axle assembly
are being pulled forwards and around the patch
centre caused by the negative offset distance. The
consequence of negative offset is that the effective
braked wheel twists in the opposite direction to that
to which the vehicle tends to veer (Fig. 10.14) and so
counteracts the swerving tendency, enabling the
vehicle to remain in a stable straight ahead direction.
In both positive and negative offset layouts, the
skidding wheel turns in the same direction as the
initial swerving tendency, but since it is not contributing greatly to the tyre to ground grip, its
influence on directional stability is small.
The effect of negative offset is ideal for a split
line braking system where if one brake line should
fail, the front brake on the opposite side will still
operate as normal (Fig. 10.14). The tendency for
the car to veer to the side of the braked wheel is
partially corrected by the wheel being turned due to
the negative offset in the opposite direction
(inwards), away from the direction in which the
car wants to swerve.
When cornering, the sideways distortion of the
tyre walls will misalign the wheel centre to that of
the tread centre so that the swivel ball joint inclination offset will alter. The outer front wheel
which supports the increase in weight due to
body roll reduces positive offset (Fig. 10.15(a)),
while negative offset becomes larger (Fig.
10.15(b)) and therefore makes it easier for the
car to be steered when negotiating a bend in the
10.1.6 MacPherson strut friction and spring
offset (Figs 10.16 and 10.17)
The MacPherson strut suffers from stickiness
in the sliding motion of the strut, particularly
under light load with an extended strut since
the cylinder rod bearing and the damper piston
will be closer together. Because the alignment
of the strut depends upon these two sliding
members, extending and reducing their distance will increase the side loading under these
The problem of reducing friction between the
inner and outer sliding members is largely overcome in two ways:
Fig. 10.17 Coil spring to swivel pin axis offset
counteracts bending moment
Fig. 10.15 (a and b)
when cornering
(a) By reducing the friction, particularly with any
initial movement, using a condition which is
known as stiction. This is achieved by facing
the bearing surfaces with impregnated polytetra-fluorethytene (PTFE) which gives the
rubbing pairs an exceptionally low coefficient
of friction.
(b) By eliminating the bending moment on the
strut under normal straight ahead driving
although there will be a bending moment
under cornering conditions.
Swivel pin inclination offset change
The tendency for the strut to bend arises because
the wheel is offset sideways from the strut, causing
the stub axle to act as a cantilever from the base of
the strut to the wheel it supports, with the result the
strut bends in a curve when extended or under
heavy loads (Fig. 10.16).
A simple solution which is commonly applied to
reduce the bending moment on the strut is to angle
the axis of the coil spring relative to the swivel joint
axis causing the spring to apply a bending moment
in the opposite sense to the vehicle load bending
moment (Fig. 10.17). Under normal conditions this
coil spring axis tilt is sufficient to neutralize the
bending moment caused by the inclined strut and
the stub axle offset, but the forces involved while
cornering produce much larger bending moments
which are absorbed by the rigidity of the strut alone.
10.2 Suspension roll centres
Roll centres (Fig. 10.29) The roll centre of a suspension system refers to that centre relative to the
ground about which the body will instantaneously
Fig. 10.16 Concentric coil spring and swivel pin axes
permit bending moment reaction
rotate. The actual position of the roll centre varies
with the geometry of the suspension and the angle
of roll.
10.2.2 Short swing arm suspension
(Fig. 10.18)
When cornering, an overturning moment is generated which makes the body roll outwards from the
centre of turn. The immediate response is that the
inner and outer swing arm rise and dip respectively
at their pivoted ends so that the inner and outer
wheels are compelled to tilt on their instantaneous
tyre to ground centres, IWG1 and IWG2, in the opposite direction to the body roll.
For effective body roll to take place there
must be two movements within the suspension
Roll axis (Fig. 10.29) The roll axis is the line joining the roll centres of the front and the rear suspension. Roll centre height for the front and rear
suspension will be quite different; usually the front
suspension has a lower roll centre than that at the
rear, causing the roll axis to slope down towards the
front of the vehicle. The factors which determine
the inclination of the roll axis will depend mainly
on the centre of gravity height and weight distribution between front and rear axles of the vehicle.
1 The swing arm pivot instantaneous centres IWB1
and IWB2 rotate about their instantaneous centres
IWG1 and IWG2 in proportion to the amount of
body roll.
2 The swing arm pivot instantaneous centres IWB1
and IWB2 move on a circular path which has a
centre derived by the intersecting projection lines
drawn through the tyre to ground instantaneous
centres IWG1 and IWG2.
10.2.1 Determination of roll centre height
(Fig. 10.18)
The determination of the roll centre height can be
best explained using the three instantaneous centre
method applied to the swing axle suspension, which
is the basic design used for the development of
almost any suspension geometry (Fig. 10.18).
A vehicle's suspension system involves three
principal items; the suspended body B, the supporting wheels W and the ground G which provides the
reaction to the downward load of the vehicle.
If a body which is suspended between two pairs
of wheels is to be capable of rolling relative to the
ground, then there must be three instantaneous
centres as follows:
The tilting, and therefore rotation, of both
swing arms about the tyre to ground instantaneous centres IWG1 and IWG2 will thus produce
an arc which is tangential to the circle on which
the swing arm pivot instantaneous centres IWB1
and IWB2 touch. Therefore, the intersecting point
IBG, where the projection lines which are drawn
through the wheel to ground contact points and
the swing arm pivots meet, is the instantaneous
centre of rotation for the body relative to the
ground. This point is usually referred to as the
body roll centre.
Thus the body roll centre may be found by drawing a straight line between the tyre contact centre
and swing arm pivot centre of each half suspension
and projecting these lines until they intersect somewhere near the middle of the vehicle. The point of
intersection becomes the body roll centre.
The roll centre height may be derived for a short
swing arm suspension by consideration of similar
1 IBG the instantaneous centre of the body relative
to the ground which is more commonly known
as the body roll centre,
2 IWB the instantaneous centre of the wheel relative
to the body which is the swing arm point of pivot,
3 IWG the instantaneous centre of the wheel relative to the ground which is the contact centre
between the tyre and ground. It therefore forms
a pivot permitting the top of the wheel to tilt
laterally inwards or outwards.
t=2 l
Fig. 10.18 Short swing axle
= Roll centre height
= Track width
= Wheel radius
= Swing arm length
10.2.3 Long swing arm suspension (Fig. 10.19)
The long swing arm suspension is very similar to
the short swing arm arrangement previously
described, but the arms extend to the opposite
side of the body relative to its wheel it supports
and therefore both arms overlap with each other
(Fig. 10.19).
The roll centre is determined by joining the tyre
contact centre and the swing arm pivot centre by a
straight line for each half suspension. The point
where these lines meet is the body roll centre and
its distance above or below the ground is known as
the roll centre height. Because the long swing arm
suspension has a much longer arm than used on the
short swing arm layout, the slope of the lines joining the tyre contact centre and swing arm pivot is
not so steep. Therefore the crossover point which
determines the body roll centre height is lower for
the long swing arm than for the short swing arm
The inherent disadvantage of the short swing
arm suspension is that there is too much camber
change with body roll and there is a tendency for
the axle arms to jack the body up when cornering.
Whereas the long swing arm suspension would
meet most of the requirements for a good quality
ride, it is impractical for a front suspension layout
as it would not permit the engine to be situated
relatively low between the two front wheels.
10.2.4 Transverse double wishbone suspension
(Figs 10.20, 10.21 and 10.22)
If lines are drawn through the upper and lower
wishbone arms and extended until they meet either
inwards (Fig. 10.20) or outwards (Fig. 10.21), their
intersection point becomes a virtual instantaneous
centre for an imaginary (virtual) triangular swing
arm suspension. The arc scribed by the wishbone
arms pivoting relative to the body is almost identical to that of the imaginary or virtual arm which
swings about the instantaneous virtual centres IBW1
Fig. 10.20
Inward converging transverse double
Fig. 10.21
Outward converging transverse double
Fig. 10.22
Parallel transverse double wishbone
and IBW2 for small movements of the suspension.
Therefore, the body roll centre for a transverse
double wishbone suspension can be derived similarly to a long swing arm suspension.
For inwardly converging transverse upper and
lower wishbone arm suspension (Fig. 10.20) the
body roll centre can be derived in two stages.
Firstly, extend straight lines through the wishbone
arms until they meet somewhere on the opposite
side of the body at their virtual instantaneous
centres IWB1 and IWB2. Secondly, draw straight lines
between the tyre contact centres IWG1 and IWG2 and
the virtual centres IBW1 and IBW2 for each half
suspension. The point where these inclined lines
intersect is therefore the body roll centre IBG.
For outward converging transverse upper and
lower wishbone arm suspension (Fig. 10.21) the
body roll centre is found again by drawing two
Fig. 10.19 Long swing axle
sets of lines. Firstly project straight lines through
the wishbone arms for each side of the vehicle until
they meet somewhere on the outside of each wheel
at their virtual instantaneous centres IWB1 and IWB2.
Next draw straight lines between the tyre contact
centres IWG1 and IWG2 and the virtual centres IWB1
and IWB2 for each half suspension, and at the same
time extend these lines until they intersect near the
middle of the vehicle. This point therefore becomes
the body roll centre IBG. It can be seen that inclining the wishbone arms so that they either converge
inward or outward produces a corresponding high
and low roll centre height.
With parallel transverse upper and lower wishbone arms suspension (Fig. 10.22) lines drawn
through the double wishbone arms would be parallel. They would never meet and so the virtual
instantaneous centres IWB1 and IWB2 would tend
to infinity 1. Under these circumstances, lines
normally drawn between the tyre contact centres
IWG1 and IWG2 and the virtual instantaneous
centres IWB1 and IWB2 would slope similarly to
the wishbone extended lines. Consequently, the
downwardly inclined parallel wishbone suspension
predicts the tyre contact centre to virtual centre
extended lines which meet at the roll centre would
meet just above ground level. Therefore if the parallel wishbone arms were horizontally instead of
downwardly inclined to the ground then the body
roll centre would be at ground level.
Fig. 10.24 Vertical pillar strut
verse swing tendency about some imaginary pivot.
Lines drawn through the two trailing arm pivot axes
or sliding pillar stub axle, which represent the principle construction points for determining the virtual
swing arm centres, project to infinity. The tyre contact centre to virtual instantaneous centre joining
lines projected towards the middle of the vehicle
will therefore meet at ground level, thus setting the
body roll centre position. Inclining the trailing arm
pivot axes or the vertical sliding pillar axis enables
the roll centre height to be varied proportionally.
10.2.5 Parallel trailing double arm and vertical
pillar strut suspension (Figs 10.23 and 10.24)
In both examples of parallel double trailing arm
(Fig. 10.23) and vertical pillar strut (Fig. 10.24)
suspensions their construction geometry becomes
similar to the parallel transverse double wishbone
layout, due to both vertical stub axle members moving parallel to the body as they deflect up and down.
Hence looking at the suspension from the front,
neither the double trailing arms (Fig. 10.23) nor
the sliding pillar (Fig. 10.24) layout has any trans-
10.2.6 MacPherson strut suspension (Fig. 10.25)
To establish the body roll centre height of any
suspension, two of the three instantaneous centres,
the tyre contact centre and the swing arm virtual
centre must first be found. If straight lines are
drawn between, and in some cases projected
beyond, these instantaneous centres the third
instantaneous centre which is the body roll centre
becomes the point where both lines intersect.
The tyre contact centres (instantaneous centres
IWG1 and IWG2) where the wheels pivot relative to
the ground are easily identified as the centres of the
tyre where they touch the ground, but the second
instantaneous virtual centre can only be found
once the virtual or imaginary equivalent swing
arm geometry has been identified.
For the MacPherson strut suspension (Fig.
10.25) the vertical swing arm and pivot centres
IBW1 and IBW2 are obtained for each half suspension
by projecting a line perpendicular to the direction
Fig. 10.23 Parallel trailing double arm
Fig. 10.25 MacPherson strut
of strut slide at the upper pivot. A second line is
then drawn through and beyond the lower control
arm until it intersects the first line. This point is the
instantaneous virtual centre about which the virtual swing arm pivots.
Straight lines are then drawn for each half suspension between the tyre contact centre and the
virtual swing arm centre. The point of intersection
of these two lines will then be the third instantaneous centre IBG, commonly referred to as the
body roll centre.
10.2.8 High load beam axle leaf spring sprung
body roll stability (Fig. 10.27)
The factors which influence the resistance to body
roll (Fig. 10.27) are as follows:
a) The centrifugal force acting through the centre
of gravity of the body load.
b) The arm length from the centre of load to the
effective roll centre h1 or h2.
c) The spring stiffness in Newtons/metre of vertical spring deflection.
d) The distance between the centres of both
springs known as the spring stability base ts.
e) The distance between road wheel centres known
as the tyre stability base tw.
10.2.7 Semi-trailing arm rear suspension
(Fig. 10.26)
A semi-trailing arm suspension has the rear wheel
hubs supported by a wishbone arm pivoted on an
inclined axis across the body (Fig. 10.26(a)).
If lines are projected through the wishbone arm
pivot axis and the wheel hub axis they will intersect
at the virtual instantaneous centres IBW1 and IBW2
(Fig. 10.26(a and b)). The distance between these
centres and the wheel hub is the transverse equivalent
(virtual) swing arm length a. Projecting a third line
perpendicular to the wheel hub axis so that it intersects the skewered wishbone arm axis produces the
equivalent fore and aft (trailing) swing arm length b
for the equivalent (virtual) semi-trailing triangular
arm (Fig. 10.26(c)). The movement of this virtual
swing arm changes the wheel camber and moves
the wheel hub axis forward as the wheel deflects in
bump or bounce from the horizontal position.
The body roll centre can now be determined by
drawing a rear view of both virtual swing arms
(Fig. 10.26(b)) and then drawing lines between
each half swing arm instantaneous pivot centres
IWB1 and IWB2 and the tyre contact centres IWG1
and IWG2. The point where these two sloping lines
cross over can then be defined as the body roll
centre IBG.
Considering the same side force acting through
the centre of gravity of the body load and similar
spring stiffness for both under- and over-slung
springs (Fig. 10.27), two fundamental observations
can be made.
Firstly it can be seen (Fig. 10.27) that with overslung springs the body roll centre RC1 is much
higher than that for underslung springs RC2 and
therefore the overslung springs provide a smaller
overturning arm length h1 as opposed to h2 for the
underslung springs. As a result, the high roll centre
with the small overturning arm length offers
a greater resistance to body roll than a low roll
centre with a long overturning arm.
Secondly it can be seen (Fig. 10.27) that the
triangular projection lines produced from the centre
of gravity through the centres of the springs to
Fig. 10.27 Effects of under- and over-slung springs on
the roll centre height
Fig. 10.26 Semi-trailing arm
the ground provide a much wider spring stability
base for the high mounted springs compared to
the low mounted underslung springs. In fact the
overslung spring centre projection lines nearly
approach the tyre stability base width tw which
is the widest possible for such an arrangement
without resorting to outboard spring seats.
Methods used to locate and control the axle
movement are considered as follows:
Longitudinally located semi-elliptic springs
(Fig. 10.28(a)) When semi-elliptic leaf springs
support the body, the pivoting point or body roll
centre will be roughly at spring-eye level but this
will become lower as the spring camber (leaves
bow) changes from positive upward bowed leaves
when unloaded to negative downward bowed
leaves with increased payload.
10.2.9 Rigid axle beam suspension
(Fig. 10.28(a±d))
An axle beam suspension is so arranged that both
wheel stub axles are rigidly supported by a common transverse axle beam member which may be a
steered front solid axle beam, a live rear axle hollow
circular sectioned casing or a DeDion tubular axle
With a rigid axle beam suspension there cannot
be any independent movement of the two stub axles
as is the case with a split swing axle layout. Therefore any body roll relative to the ground must take
place between the axle beam and the body itself.
Body roll can only take place about a mechanical
pivot axis or about some imaginary axis somewhere near mid-spring height level.
Fig. 10.28 (a±d)
Transverse located Panhard rod (Fig. 10.28(b)) The
use of coil springs to support the body requires
some form of lateral body to axle restraint if a
torque tube type axle is to be utilized. This may
be provided by a diagonally positioned Panhard
rod attached at its ends to both the axle and
body. When the body tilts it tends to move sideways and either lifts or dips depending which way
the side force is applied. Simultaneously the body
will roll about the mid-position of the Panhard rod.
Diagonally located tie rods (Fig. 10.28(c)) To provide both driving thrust and lateral support for
Body roll centres for rigid beam axle suspensions
a helical coil spring live axle layout, a trailing four
link suspension may be adopted which has a pair of
long lower trailing arms which absorb both the
driving and braking torque reactions and a pair of
short upper diagonally located tie rods to control
any lateral movement. Any disturbing side forces
which attempt to make the body tilt sideways will
cause it to roll about a centre roughly in line with
the upper tie rod height.
overturn. An overturning moment is therefore generated which tends to transfer weight from the
inner wheels to the outside wheels. At the same
time due to the flexibility and softness of the suspension, the body rolls so that in effect it overhangs
and imposes an additional load to the outer wheels.
The opposition to any body roll will be shared
out between the front and rear suspension according to their roll resistance. Thus if the front suspension roll stiffness with an anti-roll bar is twice that
of the rear, then the front wheels will sustain two
thirds of the roll couple while the rear ones only
carry one third.
Transverse Watt linkage (Fig. 10.28(d)) An alternative arrangement for controlling the sideways
movement for a coil spring suspension when used
in conjunction with either a live axle or a DeDion
tube is the Watt linkage. Suspension linkages of
this type consist of a pair of horizontal tie rods
which have their outer ends anchored to the body
and their inner ends coupled to a central balance
lever which has its pivot attachment to the axle
beam. If the body is subjected to an overturning
moment it will result in a body roll about the Watt
linkage balance lever pivot point. This instantaneous centre is therefore the body roll centre.
10.3.1 Body roll couple (Fig. 10.29)
The body roll couple (moment) M consists of two
Centrifugal moment about the roll centre ˆ
Fa …Nm†
Transverse displacement moment ˆ w a tan °Wa (Nm)
where1 F = centrifugal side force
a = distance between the centre of
gravity and roll centre
w = unsprung weight
= angle of body roll
Total roll movement or couple M ˆ Fa ‡ Wa
ˆ (F ‡ W) a (Nm)
10.3 Body roll stability analysis
When a vehicle turns a corner the centrifugal force
produced acts outwards through the centre of gravity of the sprung mass, but it is opposed by the tyre
to ground reaction so that the vehicle will tend to
Fig. 10.29 Body roll centres and roll axis
The sum of these couples are resisted by the
springs in proportion to their torsional stiffness at
the front and rear.
It should be noted that the centre of gravity
height h is made up from two measurements; the
distance between the ground and the body roll
centre b and the distance between the roll centre
and the centre gravity a.
Body roll stiffness (Fig. 10.29) The body roll stiffness is defined as the roll couple produced per
degree of body roll.
Total body roll couple ˆ Fh ˆ F(a ‡ b) (N)
M ˆ Fa ‡ Fb (N)
Roll couple
Roll angle
S ˆ (Nm=deg)
S = roll stiffness (Nm/deg)
M = roll couple (Nm)
= angle of roll (deg)
i:e: Roll stiffness ˆ
10.3.3 Body roll weight transfer (Fig. 10.31)
The product Fa is the overturning couple rotating
about the roll centre which causes the body to roll.
This couple is opposed by a reaction couple Rt
where R is the vertical reaction force due to the
weight transfer and t is the wheel track width.
Rt ˆ Fa
This shows that as the distance between the
ground and the body roll centre known as the
couple arm becomes smaller, the overturning couple
and therefore the body roll will also be reduced
in the same proportion. Thus if the couple arm a
is reduced to zero the reaction force R will likewise
approach zero. A small couple is desirable so that
the driver experiences a sense of body roll as
a warning for cornering stability. If both roll centre
and centre of gravity height coincided there would
be no indication to the driver that the lateral forces
acting on the body were reaching the limit of the
tyre to ground sideway grip. Consequently suspensions in which the centre of gravity and the roll
centre are at the same height can cause without
warning a sudden tyre to ground breakaway when
cornering at speed.
The fraction of torsional stiffness for the front
and rear suspensions will therefore be:
SF ˆ
SF ‡ SR Nm=deg
SR ˆ
SF = fraction of front torsional stiffness
SR = fraction of rear torsional stiffness
10.3.2 Body overturning couple (Fig. 10.30)
The centrifugal force F created when a vehicle is
travelling on a circular track acts through the
body's centre of gravity CG at some height h and
is opposed by the four tyre to ground reaction
forces F1, F2, F3 and F4.
Consequently an overturning couple Fh is produced which transfers weight W from the inside
wheels to the outer wheels which are spaced the
track width t apart. Thus the overturning couple
will also be equivalent to Wt, that is, Wt ˆ Fh.
i:e: Weight transferred W ˆ
10.3.4 Body direct weight transfer couple
(Fig. 10.32)
If the centrifugal force acted through the roll centre
axis instead of through the centre of gravity, a
Fig. 10.30 Overturning couple
Fig. 10.31 Body roll weight transfer
MF ˆ
(F ‡ W)a ‡ FF hF (Nm)
Roll couple on rear tyres
MR ˆ
(F ‡ W)a ‡ FR hR (Nm)
Body roll angle The body roll angle may be
defined as the roll couple per unit of roll stiffness
Roll couple
Roll stiffness Nm=deg
Fig. 10.32 Direct weight transfer
moment Fb about the ground would be produced
so that a direct transference of weight from the
inner to the outer wheels occurs. The reaction to
this weight transfer for a track width t is a resisting
moment Rt which is equal but opposite in sense to
the moment Fb.
10.3.6 Body roll weight transfer (Fig. 10.29)
The body roll weight transferred may be defined as
the roll couple per unit width of track
Total roll weight transfer
Roll couple Nm
Track width m
hence W ˆ (N)
Front suspension weight transfer
WF ˆ
Rt ˆ Fb,
If the fore and aft weight distribution is proportional between the front and rear axle roll centres,
the centrifugal force F acting through the roll centre axis would be split into two forces FF and FR
which act outwards from the front and rear roll
Rear suspension weight transfer
WR ˆ
F F bF
F R bR
RR ˆ
where R, RF and RR ˆ Total, front and rear
vertical reaction forces
Total roll angle ˆ
RF ˆ
W, WF and WR = Total, front and rear
weight transfer
respectively (N)
t = Wheel tract (m)
10.3.7 Lateral (side) force distribution
(Fig. 10.33)
The total lateral resisting forces generated at all
tyre to ground interfaces must equal the centrifugal
Thus lowering the body roll centre correspondingly reduces the vertical reaction force R and by
having the roll centre at ground level the direct
weight transfer couple will be eliminated.
Therefore if the roll axis slopes from the ground
upwards from front to rear, all the direct weight
transfer couple will be concentrated on the rear
10.3.5 Body roll couple distribution (Fig. 10.29)
The body roll couple on the front and rear tyres is
proportional to the front and rear suspension stiffness fraction.
Roll couple on front tyres
Fig. 10.33
Longitudinal weight distributions
Fig. 10.34 (a and b)
Comparison of rigid and independent suspension body roll stiffness
force acting through the body's centre of gravity.
Thus the fore and aft position of the centre of
gravity determines the weight distribution between
the front and rear wheels and therefore the proportion of cornering force necessary to be generated by
their respective tyres.
If FF and FR are the front and rear tyre to ground
cornering forces, then taking moments about FR
outside wheel reaction ‡W between the effective
spring span tw.
If the vertical spring stiffness is S N/m and the
vertical deflection at the extremes of the spring
span is x m then the angle of body roll Y degrees
can be derived as follows:
FF l ˆ Fb
FF ˆ (N)
FR l ˆ Fa
FR ˆ (N)
Thus the amount of load and cornering force
carried by either the front or rear tyres is proportional to the distance the centre of gravity is from
the one or the other axle. Normally there is slightly
more weight concentrated at the front half of the
vehicle so that greater cornering forces and slip
angles are generated at the front wheels compared
to the rear.
Weight transfer
tw =2 tw
W ˆxS
tan ˆ
Overturning couple ˆ Fh
Reaction couple ˆ Wt ˆ Sxt
(since W ˆ Sx)
From (1)
10.3.8 Comparison of rigid axle beam and
independent suspension body roll stiffness
(Fig. 10.24)
A comparison between roll stiffness of both rigid
axle beam and independent suspension can be
derived in the following manner:
Consider the independent suspension (Fig.
10.34(a)). Let the centrifugal force F act through
the centre of gravity CG at a height h above the roll
centre RC. The overturning couple Fh produced
must be equal and opposite to the reaction couple
Wtw created by a reduction in the inside wheel
reaction W and a corresponding increase in the
; Fh ˆ Sxtw
or x ˆ
tw S
tan ˆ
2 Fh
tan ˆ
t Stw
When is small, tan ° 2Fh
Stw 2
This formula shows that the body roll angle is
proportional to both centrifugal force F and the
couple arm height h but it is inversely proportional
to both the spring stiffness k and the square of the
spring span tw2, which in this case is the wheel
/ F, / h, /
and /
A similar analysis can be made for the rigid axle
beam suspension (Fig. 10.34(b)), except the spring
span then becomes the spring base ts instead of tw.
Because the spring span for a rigid axle beam suspension is much smaller than for an independent
suspension (tw 2 ts 2 ), the independent wide spring
span suspension offers considerably more roll resistance than the narrow spring span rigid axle beam
suspension and is therefore preferred for cars.
effective if one wheel is raised higher than the other
(Fig. 10.35) as the vehicle passes over a hump in the
road or the body commences to roll while cornering. Under these conditions, the suspension spring
stiffness (total spring rate) increases in direct proportion to the relative difference in deflection of
each pair of wheels when subjected to the bump
and rebound of individual wheels or body roll
when the vehicle is moving on a circular path.
10.4.2 Anti-roll bar construction (Fig. 10.36)
Generally the anti-roll bar is formed from a medium
carbon steel solid circular sectioned rod which is
positioned transversely and parallel to the track (Fig.
10.36) so that it nearly spans the distance between
the road wheels (Fig. 10.35). The bar is bent at both
ends in right angles to form cracked arms. These
arms can then be actuated by short link rods
attached to the unsprung portion of the suspension
such as the axle beam or transverse wishbone arms
for independent suspension. The main transverse
span of the rod is supported by rubber bearings
positioned on the inside of the cranked arms at
each end. These bush bearings are either mounted
directly onto the body structure when incorporated
10.4 Anti-roll bars and roll stiffness (Fig. 10.35)
10.4.1 Anti-roll bar function
A torsion anti-roll bar is incorporated into the
suspension of a vehicle to enable low rate soft
springs to be used which provides a more comfortable ride under normal driving conditions. The
torsion bar does not contribute to the suspension
spring stiffness (the suspension's resistance to vertical deflection) as its unsprung weight is increased
or when the driven vehicle is subjected to dynamic
shock loads caused possibly by gaps or ridges
where concrete sections of the road are joined
together. However, the anti-roll bar does become
Fig. 10.35 Transverse double wishbone coil spring independent suspension with anti-roll bar
Fig. 10.36 Transverse double wishbone torsion bar independent suspension with anti-roll bar
Fig. 10.37 Relationship of body roll and suspension
spring and anti-roll bar stiffness
Fig. 10.38 Relationship of body roll and the understeer
tendency with and without an anti-roll bar
on cars (Fig. 10.35) or indirectly for commercial
vehicles (Fig. 10.39) on short vertical arms which
provide a swing attachment to the chassis.
10.4.3 Anti-roll bar operating principle
When a pair of road wheels supported on an axle
travel over a bumpy road one or other wheel will lift
and fall as they follow the contour of the road
surface. If the springs were relatively hard, that is
they have a high spring rate, then the upthrust
caused by the bumps would be transmitted to the
body which would then lift on the side being disturbed. Thus the continuous vertical deflection of
either wheel when the vehicle moves forward would
tend to make the body sway from side to side producing a very uncomfortable ride. On the other hand
if softer springs were used for the suspension, the
small road surface irregularities would be adequately
absorbed by the springs and dampers, but when
cornering there would be insufficient spring stiffness
to resist the overturning moment; this would therefore permit excessive body roll which could not be
tolerated. Incorporating an anti-roll bar with relatively soft suspension springs mostly overcomes the
difficulties discussed and therefore greatly improves
the vehicle's ride. This is possible because the soft
springs improve the suspension's response on good
straight roadways (Fig. 10.37), with the benefits of
the anti-roll bar automatically increasing the suspension roll stiffness when the vehicle is cornering.
produces an overturning moment created by its
offset to the body's roll centre which will therefore
tend to make the body roll (Fig. 10.39(a and b)).
The rolling body will tilt the transverse span of the
roll bar with it so that the cranked arms on the
outside wheel to the turn will be depressed downward, whereas the cranked arm on the opposite end
near the inside wheel to the turn will tend to rise.
The consequence of this misalignment of the antiroll bar arms is that the two cranked arms will
rotate in opposite directions to each other and so
transmit a torque from the inside wheel which is
subjected to less load to the outside wheel which is
now more heavily loaded. The effect of the torsional wind-up in the bar is that it tries to rotate
the outside wheel cranked arm and since the arm is
attached to the axle or indirectly to the wishbone
arm it cannot move. The alternative is for the roll
bar and the rubber bearing mount near the outside
wheel to lift in proportion to the degree of twisting
torque. It therefore counteracts some of the downward push due to the increase in weight to the
outside wheel and as a result stiffens the roll resistance of the springing on the outside wheel as a
whole. Consequently a larger slip angle is generated
on the front outside wheel relative to the rear
wheel, and as a result, the vehicle will develop a
small initial but progressive understeer tendency
approximately proportional to the amount the
body rolls (Fig. 10.38).
10.4.4 Anti-roll bar action caused by the body
rolling (Fig. 10.39(a and b))
When cornering, the centrifugal force acting
through the centre of gravity of the sprung body
10.4.5 Anti-roll bar action caused by single wheel
lift (Fig. 10.39(c and d))
If one of a pair of axle wheels lifts as it climbs over a
bump (Fig. 10.39(c)) in the road, then the vertical
Fig. 10.39 (a±d)
Anti-roll bar action
deflection of the wheel and spring raises and rotates
the anti-roll bar's cranked arm on that side so that
the transverse span of the bar is twisted. The bar is
therefore subjected to a torque which is proportional to its angle of rotation.
This twisting torque is transferred to the opposite cranked arm which then applies a downward
force onto the axle and wheel. However, because
the wheel cannot sink into the ground, the reaction
occurs on the rubber bearing mount arm which
therefore tends to lift up the side of the chassis on
the opposite side to the vertically deflected wheel.
As a result, both sides of the chassis (body) will
have been raised, thereby enabling the vehicle's
body to remain upright instead of tilting to one
side. Similarly, if the opposite wheel hits an
obstacle in the road (Fig. 10.39(d)), the torsional
wind-up of the bar transfers an upward thrust to the
other side, which again tends to lift the chassis on
the undisturbed wheel side and so maintains
the sprung chassis and body on an even keel
(Fig. 10.39(c)).
10.5 Rubber spring bump or limiting stops
10.5.1 Bump stop function
(Figs 10.40 and 10.42)
Suspension bump and body roll control depends
upon the stiffness of both the springs and anti-roll
bar over the normal operating conditions, but if the
suspension deflection approaches maximum bump
or roll the bump stop (Fig. 10.40(a, b, c and d))
becomes active and either suddenly or progressively provides additional resistance to the full
deflection of the wheel and axle relative to the
body (Fig. 10.42). The bump stop considerably
stiffens the resisting spring rate near the limit of
its vertical movement to prevent shock impact and
damage to the suspension components. The stop
also isolates the sprung and unsprung members of
the suspension under full deflection conditions so
that none of the noise or vibrations are transmitted
through to the body structure. In essence the bump
stop enables an anti-roll bar to be used which has
a slightly lower spring rate, therefore permitting
a more cushioned ride for a moderate degree of
body roll.
Bump rubber spring stops may be solid and
conical in shape or they may be hollow and cylindrical or rectangular shaped with a bellow profile
(Fig. 10.40(a, b, c and d)) having either a single,
double or triple fold (known as convolutions). The
actual profile of the rubber bump stop selected will
depend upon the following:
10.5.2 Bump stop construction (Fig. 10.40(a±d))
Bump stops may be considered as limiting springs
as they have elastic properties in compression
similar to other kinds of spring materials. Solid
and hollow spring stops are moulded without
reinforcement from natural rubber compound
containing additives to increase the ozone resistance. The deflection characteristics for a given
size of rubber stop spring are influenced by the
hardness of the rubber, this being controlled to
a large extent by the proportion of sulphur and
carbon black which is mixed into the rubber compound. The most common rubber compound
hardness used for a rubber spring stop is quoted
as a shore hardness of 65; other hardness ranging
from 45 to 75 may be selected to match a particular operating requirement. A solid cylindrical
rubber block permits only 20% deflection when
loaded in compression, whereas hollow rubber
spring stops have a maximum deflection of 50±75%
of their free height. The actual amount of deflection for a given spring stop height and response
to load will depend upon a number of factors
such as the rubber spring stop size, outer profile,
wall thickness, shape of inner cavities, hardness of
rubber compound and the number of convolution
Fig. 10.40 (a±d)
1 How early in the deflection or load operating
range of the suspension the rubber begins to
compress and become active.
2 Over what movement and weight change the
bump stop is expected to contribute to the sudden or progressive stiffening of the suspension so
that it responds to any excessive payload, impact
load and body roll.
10.5.3 Bump stop characteristics
(Figs 10.41 and 10.42)
The characteristics of single, double and triple convolution rubber spring stops, all using a similar
rubber hardness, are shown in Fig. 10.41. It can
be seen that the initial deflection for a given load is
large but towards maximum deflection there is very
little compression for a large increase in load. The
relation between load and deflection for bump is
not quite the same on the release rebound so that
the two curves form what is known as a hysteresis
loop. The area of this loop is a measure of the
energy absorbed and the internal damping within
Suspension bump stop limiter arrangements
10.6 Axle location
10.6.1 Torque arms (Figs 10.28(c) and 10.44)
Torque arms, sometimes known as radius arms
or rods, are mounted longitudinally on a vehicle
between the chassis/body structure and axle or
unsprung suspension member. Its purpose is to
permit the axle to move up and down relative to
the sprung chassis/body and to maintain axle alignment as the torque arm pivots about its pin, ball
or conical rubber joint. Sometimes the upper
torque rods are inclined diagonally to the vehicle's
lengthwise axis to provide lateral axle stability
(Figs 10.28(c) and 10.44). These arms form the
link between the unsprung suspension members
and the sprung chassis/body frame and are therefore able to transmit both driving and braking
forces and to absorb the resulting torque reactions.
Fig. 10.41 Characteristics of hollow rubber single,
double and triple convolute progressive bump stops
10.6.2 Panhard rod (Fig. 10.28(b))
Panhard rods, also known as transverse control
rods or arms, are positioned across and between
both rear wheels approximately parallel to the axle
(Fig. 10.28(b)). One end of the rod is anchored to
one side of the axle span while the other end is
anchored to the body structure; both attachments
use either pin or ball type rubber joints. A Panhard
rod restrains the body from moving sideways as the
vehicle is subjected to lateral forces caused by sidewinds, inclined roads and centrifugal forces when
cornering. When the body is lowered, raised or
tilted relative to the axle, the Panhard rod is able
to maintain an approximate transverse axle alignment (Fig. 10.28(b)) relative to the chassis/body
thus relieving the suspension springs from side
10.6.3 Transverse located Watt linkage
(Fig. 10.43(a, b and c))
A Watt linkage (Fig. 10.43) was the original
mechanism adopted by James Watt to drive his
beam steam engine. This linkage is comprised of
two link rods pivoting on the body structure at
their outer ends and joined together at their inner
ends by a coupler or equalizing arm which is
pivoted at its centre to the middle of the rear axle.
When in mid-position the link rods are parallel
whereas the equalizing arm is perpendicular to
both (Fig. 10.43(b)).
If vertical movement of the body occurs either
towards bump (Fig. 10.43(c)) or rebound (Fig.
10.43(a)) the end of the link rods will deviate an
equal amount away from the central pivot point of
the coupling arm. Thus the left hand upper link rod
Fig. 10.42 Combined characteristics of suspension
spring and rubber bump stops
the rubber in one cycle of compression and expansion of the rubber spring stop. For hollow rubber
spring stops they always end in a point; this means
for any load change there will be some spring
Fig. 10.42 shows how the bump spring stop
deviates from the main spring load-deflection
curve at about two-thirds maximum deflection
and that the resultant stiffness (steepness of curve)
of the steel spring, be it leaf, coil or torsion bar, and
that of the bump spring stops considerably
increases towards full load.
10.7 Rear suspension arrangements
10.7.1 Live rigid axle rear suspension
Suspension geometry characteristics of a live axle
are as follows:
Fig. 10.43 (a±c)
1 Wheel camber is zero irrespective if the vehicle is
stationary or moving round a bend in the road.
2 If one wheel moves over a hump or dip in the
road then the axle will tilt causing both wheels to
become cambered.
3 Because both wheels are rigidly joined together
the wheel track remains constant under all driving conditions.
4 Because the axle casing, half shafts and final
drive are directly supported by the wheels, the
unsprung weight of a live axle is very high.
5 With a live rigid axle, which is attached to the
body by either leaf or coil springs, the body will
tilt about some imaginary roll centre roughly
mid-way between the upper and lower spring
anchorage points.
6 Horizontal fore and aft or lateral body location
is achieved by using the leaf springs themselves
as restraining members or, in the case of coil
springs which can only support the vehicle's vertical load and therefore cannot cope with driving
thrust and side loads, horizontally positioned
control rods.
Transversely located Watt linkage
will tend to pull towards the left and the right hand
lower link rod will apply an equal pull towards the
right. The net result will be to force the equalizing
arm to rotate anticlockwise to accommodate the
inclination to the horizontal of both link rods. If
the left hand link rod were made the lower link and
the right hand rod the upper link, then the direction
of tilt for the equalizing arm would now become
For moderate changes in the inclination of the
link rods, the body will move in a vertical straight
line, thus maintaining a relatively accurate body to
axle lateral alignment. Excessive up and down
movement of the body will cause the pivot centre
to describe a curve resembling a rough figure eight,
a configuration of this description being known as
a lemniscoid (Fig. 10.43(b)).
Under body roll conditions when cornering,
the whole body relative to the axle and wheels will
be restrained to rotate about the equalizing
arm pivot centre at mid-axle height; this point
therefore becomes the roll centre for the rear end
of the body.
A similar Watt linkage arrangement can be
employed longitudinally on either side of the wheels
to locate the axle in the fore and aft direction.
Without accurate control of horizontal body
movement relative to the axle casing caused by
vertical deflection of the springs or longitudinal
and transverse forces, the body's weight distribution would be unpredictable which would result in
poor road holding and steering response.
Hotchkiss drive suspension (Fig. 10.84(a)) This is
the conventional semi-elliptic spring suspension
which has each spring positioned longitudinally
on each side of the axle and anchored at the front
end directly to a spring hanger attached to the body
structure and at the rear end indirectly via swing
shackle plates to the rear spring hangers, the axle
being clamped to the springs somewhere near their
mid-span position. Thus fore and aft driving and
braking forces are transmitted through the front
half of the springs and lateral forces are accommodated by the rigidity of the spring leaves and spring
Four link coil spring live axle rear suspension
(Fig. 10.44) Substituting coil springs for semi-elliptic
springs requires a separate means of locating
and maintaining body and axle alignment when
Fig. 10.44 Four link coil spring live axle rear suspension
arms are normally inclined at 45 to the car's centre
line axis so that they can absorb any axle reaction
torque tending to rotate the axle, and at the same
time prevent relative lateral movement between the
body and axle. Body roll or axle tilt are permitted
due to the compliance of the rubber pin joints.
A relatively high roll centre is obtained which
will be roughly at the upper torque arm height.
subjected to longitudinal and transverse forces
caused by spring deflection, body roll or driving
and braking thrust loads.
The locating links are comprised of a pair of long
trailing lower arms and a pair of short diagonally
positioned upper torque arms (Fig. 10.44). Rubber
pin joints secure the forward ends of the arms to
the body structure but the lower rear ends are
attached underneath the axle tubes as far apart as
possible and the upper short torque arms attached
much closer together onto the final drive housing.
The coil springs are mounted between the upper
body structure and the lower pressed steel trailing
arms. These springs only provide vertical support
and cannot restrain any horizontal movement on
their own. Spring deflection due to a change in
laden weight causes both sets of arms to swivel
together, thereby preventing the axle assembly
rotating and possibly making the universal joints
operate with very large angles. Both driving and
braking thrust are transmitted through the lower
trailing arms which usually are of a length equal to
roughly half the wheel track so that when the arms
swing the change in wheelbase is small. The upper
Torque tube rear wheel drive suspension (Fig. 10.45)
One of the major problems with the Hotchkiss
drive layout is that the axle torque reaction tends
to spin the axle casing when transmitting drive
torque in the opposite direction to the rotating
wheels and when braking to twist the axle casing
in the same direction as the revolving wheels. The
result is a considerably distorted semi-elliptic
spring and body to axle misalignment. To overcome this difficulty, a rigid tube may be bolted to
the front of the final drive pinion housing
which extends to the universal joint at the rear of
the gearbox or a much shorter tube can be used
which is supported at its front end by a rubber pin
or ball joint attached to a reinforcing cross-member
Fig. 10.45 Torque tube with trailing arm and transverse Watt linkage live axle rear suspension
(Fig. 10.45). On either side of the torque tube
is a trailing arm which locates the axle and also
transmits the driving and braking thrust between
the wheels and body. Coil springs are mounted
vertically between the axle and body structure,
their only function being to give elastic support to
the vehicle's laden weight. Lateral body to axle
alignment is controlled by a transverse Watt
linkage. The linkage consists of an equalizing arm
pivoting centrally on the axle casing with upper and
lower horizontal link arms anchored at their outer
ends by rubber pin joints to the body structure.
Thus when the springs deflect or the body rolls,
the link arms will swing about their outer body
location centres causing the equalizing arm to
tilt and so restrain any relative lateral body to
axle movement without hindering body vertical
With the transversely located Watt linkage, the
body roll centre will be in the same position as the
equalizing arm pivot centre. The inherent disadvantages of this layout are still the high amount
of unsprung weight and the additional linkage
required for axle location.
10.7.2 Non-drive rear suspension
The non-drive (dead) rear axle does not have the
drawback of a large unsprung weight and it has the
merit of maintaining both wheels parallel at all
times. There is still the unwanted interconnection
therefore floats and equalizes the load applied by
each shoe to the drum.
clockwise about its pivot. This permits the springloaded pawl to ride over the sector teeth until the
shoe contacts the drum. At this point the pawl teeth
drop into corresponding teeth on the sector, locking the sector lever to the leading shoe.
When the brakes are released, the shoes are
pulled inwards by the retraction spring, but only
back to where the rectangular slot contacts the
outer edge of the strut (Fig. 11.10(a)). The next
time the brakes are applied the shoes will not
move far enough out for the slot to strut clearance
to be taken up. When the brakes are released the
shoe returns to the previous position and the sector
to pawl ratchet action does not occur.
As the lining wears, eventually there will be sufficient slot to strut end clearance for the ratchet
action to take place and for the pawl to slide over
an extra sector tooth before re-engaging the sector
in a more advanced position.
Automatic adjuster operation
Brake application with new linings (Fig. 11.10(b))
When the foot brake is applied, hydraulic pressure
forces the twin plungers apart so that the shoes are
expanded against the drum. If the linings are new
and there is very little lining to shoe clearance, then
the outward movement of the leading and trailing
shoes will not be sufficient for the clearance
between the rectangular slot in the sector lever
and the strut inner edge to be taken up. Therefore
the shoes will return to their original position when
the brakes are released.
Brake application with worn lining (Fig. 11.10(c))
Applying the foot brake with worn linings makes
the brake shoes move further apart. The first part
of the outward movement of the leading shoe takes
up the clearance between the strut's inner edge and
the adjacent side of the rectangular slot formed in
the sector lever. As the shoe moves further outwards, the strut restrains the sector lever moving
with the leading shoe, so that it is forced to swivel
Fig. 11.11 (a±d)
11.3.2 Self-adjusting quadrant and pinion brake
shoe mechanism (Fig. 11.11(a±d))
This rear wheel brake layout incorporates leading
and trailing brake shoes which have a hydraulically
operated foot brake system and a mechanically
actuated hand brake. The brake shoes are mounted
Self-adjusting sector and pinion brake shoes with cross-pull hand brake
Fig. 11.11 contd
on a back plate and the lower shoe tips are prevented from rotating by a fixed anchor abutment
plate riveted to the back plate. The upper shoe tips
are actuated by twin hydraulic plungers and a hand
brake strut and lever mechanism which has a built-in
automatic shoe clearance adjustment device.
The hand brake mechanism consists of a strut
linking the two shoes together indirectly via a hand
brake lever on the trailing shoe and a quadrant lever
on the leading shoe.
the inner edge of the quadrant lever and slot touch.
Further outward shoe movement disengages the
quadrant lever teeth from the adjacent pinion teeth
and at the same time twists the lever. When the
brakes are released the retraction spring pulls the
shoes together. Initially the leading shoe web slot
contacts the outer edge of the quadrant lever, and
then further shoe retraction draws the quadrant
lever teeth into engagement with the fixed pinion
teeth, but with half worn linings the quadrant will
mesh with the pinion somewhere mid-way between
the outer edges of the quadrant. Consequently the
shoes will only be allowed to move part of the way
back to maintain a constant predetermined lining to
drum clearance.
Brake application with new linings (Fig. 11.11(a))
When the foot pedal is depressed, the hydraulic plungers are pushed apart, forcing the shoes into contact
with the drum. When the brakes are released the
retraction spring pulls the shoe inwards until the
rectangular slot in the shoe web contacts the outer
edge of the quadrant lever. The lever is then pushed
back until the teeth on its quadrant near the end of
the quadrant mesh with the fixed pinion teeth (or serrations). The position on the quadrant teeth where it
meshes with the pinion determines the amount the
shoe is permitted to move away from the drum and
the gap between the quadrant's inner edge and the
slot contact (the lining to shoe clearance).
Brake application with fully worn linings
(Fig. 11.11(d)) When the brakes are operated
with fully worn linings the shoes move outwards
before they contact the drum. During this outward
movement the rear end of the slot contacts the
inner edge of the quadrant lever, disengaging it
from the pinion. At the same time the quadrant
lever rotates until the fingered end of the lever
touches the side of the shoe web. Releasing the
brakes permits the shoes to retract until the quadrant lever contacts the pinion at its least return
position near the quadrant's edge, furthest away
from the new lining retraction position.
If any more lining wear occurs, the quadrant is
not able to compensate by moving into a more
Brake application with half worn linings (Fig. 11.11
(b and c)) When the foot brake applied, hydraulic
pressure forces the shoe plungers outwards. The
leading shoe moves out until the clearance between
raised position and therefore the master cylinder
pedal movement will become excessive, providing a
warning that the linings need replacing.
tappet head abutment which guides and supports
the twin web shoes. This construction enables the
linings to follow the drum shape more accurately.
The tappet head abutments are inclined to provide
a means for self-centralizing the brake shoes after
each brake application. The cam strut lift relative
to the camshaft angular movement tends to give an
approximately constant lift rate for the normal
angular operating range of the cam between new
and worn linings conditions.
When the brakes are applied (Fig. 11.12(b)) the
camshaft is rotated, causing the struts to move
outwards against the hollow tapper plungers. The
tappet head abutments force the shoes into contact
with the drum, thereby applying the brakes.
11.3.3 Strut and cam brake shoe expander
(Fig. 11.12(a and b))
This type of shoe expander is used in conjunction
with leading and trailing shoe brakes normally
operated by air pressure-controlled brake actuators connected to a lever spline mounted on the
camshaft, which is itself supported on a pair of
plain bronze bushes.
The camshaft mounted in the expander housing
is splined at its exposed end to support and secure
the actuating lever (Fig. 11.12(a)). The other end,
which is enclosed, supports an expander cam which
has two spherical recesses to accommodate a pair
of ball-ended struts. The opposite strut ends are
located inside a hollow tappet plunger (follower).
Mounted on the end of each tappet plunger is a
Fig. 11.12 (a and b)
11.3.4 Wedge shoe adjuster unit (Fig. 11.13)
The adjuster housing is made from malleable iron
and is spigoted and bolted firmly to the back plate
(Fig. 11.13). A hardened steel wedge is employed
Strut and cam brake shoe expander
Fig. 11.13 Wedge shoe adjuster unit
with a screw adjuster stem rotating within the
wedge which does not rotate, but moves at right
angles to the inclined faced tappet plungers. So that
accurate adjustment for each brake assembly is
possible, a clicker spring is located between the
screwed stem and the wedge. This spring fits onto
two flats provided on the screw stem (not shown).
The clicker spring has two embossed dimples which
align and clip into shallow holes formed in the back
of the wedge when the shoes are being adjusted;
they therefore enable a fine adjustment to be made
while also preventing the adjuster screw stem
unwinding on its own while in service.
circumference along each tangent and then plot
a smooth curve passing through the extending tangential lengths. The locus generated is the involute
to base circle, this shape being the basic shape of
the so-called S cam.
Cam and shoe working conditions With new shoe
linings the leading shoe works harder than the
trailing shoe so that initially the leading shoe wear
will be higher than that of the trailing shoe. If there
is adequate camshaft to bush bearing clearance,
shoe wear will eventually be sufficient to permit
the camshaft to float between the shoe tips, allowing the trailing shoe to produce the same friction
drag as the leading shoe, thus producing the equal
work condition.
If the shoe tip force applied by the cam is equal,
then the camshaft floats on its bearing. In practice,
because the shoe tip force is not always equal, a
resultant reaction force input will be provided by
the camshaft to maintain equilibrium. Therefore
the frictional force between the shaft and bearing
can be significant in the mechanical losses between
the input camshaft torque and the shoe tip force.
The input camshaft torque may be derived from
both the shaft frictional torque and the cam to
roller contact torque (Fig. 11.14(c)).
11.3.5 S cam shoe expander (Fig. 11.14)
Cam expander requirements The object of a cam
brake shoe expander (Fig. 11.14(a and b)) is to convert an input camshaft leverage torque into a shoe
tip force. The shape of the cam profile plays a large
part in the effective expansion thrust imposed on the
shoe tips as the shoe linings wear and the clearance
between the drum and linings increase.
Early S-shaped cams were derived from an
Archimedean spiral form of locus which gives a
constant rate of lift per degree of cam rotation,
but varying cam radius. The present tendency is
for the S cam to be generated from an involute spiral
(Fig. 11.14(a)) which gives a slight reduction in lift
per degree of cam rotation, but maintains a constant cam effective radius so that the shoe tip force
always acts in the same direction relative to the cam
shoe roller, no matter which part of the cam
profile is in contact with the roller (Fig. 11.14(b)).
By these means the shoe tip force will remain
approximately constant for a given input torque
for the whole angular movement of the cam
between new and worn lining conditions. Note
this does not mean that the effective input torque
will be constant. This depends upon the push or
pull rod and the slack adjuster lever remaining
perpendicular to each other which is unlikely.
C ˆ coefficient of camshaft to
bearing friction
R ˆ resultant camshaft radial
load (N)
rc ˆ camshaft radius (m)
rb ˆ base circle radius (m)
F1 and F2 ˆ roller contact forces (N)
Camshaft frictional
torque ˆ c Rrc (Nm)
Cam to roller torque ˆ (F1 ‡ F2 )rb (Nm)
Input camshaft
torque Tc ˆ c Rrc ‡ (F1 ‡ F2 )rb (Nm)
Cam profile geometry The involute to a base circle is generated when a straight line is rolled round
a circle without slipping; points on the line will
trace out an involute. The involute profile may be
produced by drawing a base circle and a straight
line equal to its circumference and dividing both
into the same number of equal parts (Fig. 11.14(a)).
From the marked points on the circle draw tangents to represent successive positions of the generated line. Step off the unwrapped portion of the
Cam design considerations To give the highest
shoe factor, that is the maximum shoe frictional
drag to input torque, a low rate of cam lift is desirable. This conflicts with the large total lift needed
to utilize the full lining thickness which tends to be
limited to 19 mm.
Typical rates of cam lift vary from 0.2 to 0.4 mm/
deg which correspond to brake factors of about
12 to 16 with the involute cam profile.
Fig. 11.14 (a±c)
Air operated foundation brake assembly
As the cam lifts (Fig. 11.14(c)) the pressure angle
which is made between the cam and roller centre
lines and the base circle tangential line decreases.
For the cam to be self-returning the pressure angle
should not be permitted to be reduced below 10 .
One approach to maximize cam lift without the
rollers falling off the end of the cam in the extreme
wear condition is to use the involute cam up to the
point where the lining rivets would contact the
drum and relining would be required. Beyond this
point the cam is continued in a straight line, tangential to the cam profile (Fig. 11.14(c)). By this
method, total cam lift is achieved for the normal
thickness of lining within the designed angular
movement of the cam, which is not possible with
the conventional involute cam. Shoe tip force efficiency does drop off in the final tangential lift cam
range but this is not a serious problem as it is very
near the end of the linings' useful life. One important outcome of altering the final involute profile is
that the blunting of the cam tips considerably
strengthens the cam.
resulting pawl lift gradually increase until the
pawl climbs over and drops into the next tooth
space. The next time, when the brake is released
and the plunger is pushed back into its bore, the
upright face of the pawl teeth prevents the sleeve
moving directly back. It permits the sleeve to twist
as its outside helical teeth slide through the corresponding guide pawl teeth in its endeavour for the
whole plunger assembly to contract inwards to the
off position, caused by the inward pull of the shoe
retraction spring. The partial rotation of the
adjusting sleeve unscrews and advances the adjusting screw to a new position. This reduces the lining
clearance. This cycle of events is repeated as the
lining wears. The self-adjustment action only operates in the forward vehicle direction. Once the
brake shoes have been installed and manually
adjusted no further attention is necessary until the
worn linings are replaced.
11.3.6 Wedge type brake shoe expander with
automatic adjustment (Fig. 11.15)
The automatic brake shoe adjustment provides a
self-adjusting mechanism actuated by the expander
movement during the on/off brake application
cycle, enabling a predetermined lining to drum
clearance to be maintained. When a brake application takes place, an adjusting pawl mechanism
senses the movement of an adjusting sleeve located
in one of the wedge expander plungers. If the sleeve
travel exceeds 1.52 mm, the spring-loaded pawl acting on the sleeve teeth drops into the next tooth and
automatically makes the adjusting screw wind out
a predetermined amount. An approximate 1.14 mm
lining to drum clearance will be maintained when
the brakes are released, but if the adjusting sleeve
and plunger outward travel is less than 1.52 mm,
then the whole plunger assembly will move back to
its original position without any adjustment being
11.3.7 Manual slack adjuster (Fig. 11.17)
Purpose A slack adjuster is the operating lever
between the brake actuator chamber push rod
and the camshaft. It is used with `S' type cam
shoe expanders and features a built-in worm and
worm wheel adjustment mechanism enabling
adjustment to be made without involving the
removal and alteration of the push rod length.
Operation (Fig. 11.17) The slack adjuster lever
incorporates in its body a worm and worm wheel
type adjuster (Fig. 11.17). The slack adjuster lever
is attached indirectly to the splined camshaft via
the internal splines of the worm wheel located
inside the slack adjuster body. For optimum input
leverage the slack adjuster lever and the push rod
should be set to maintain an inside angle just
greater than 90 with the brakes fully applied.
Once the push rod length has been set, further
angular adjustment of the `S' cam is made by rotating the worm shaft so that the large gear reduction
between the meshing worm and worm wheel will
slowly turn the worm wheel and camshaft until the
cam flanks take up the excess shoe lining to drum
clearance. Owing to the low reverse efficiency of
the worm and worm wheel gearing, the worm and
shaft will not normally rotate on its own. To prevent the possibility of the worm and shaft unwinding, caused perhaps by transmitted oscillatory
movement of the slack adjuster during periods of
applying and releasing the brakes, a lock sleeve is
Description (Fig. 11.15) The automatic adjustment is built into one of the expander plungers.
With this construction the adjusting screw is
threaded into an adjusting sleeve, the sleeve being
a free fit inside the hollow plunger. A hollow cap
screw, spring, and an adjusting pawl are preassembled and act as a plunger guide. The end of
the adjusting pawl has sawtooth type teeth which
engage corresponding helical teeth on the outside
of the adjusting sleeve.
Operation (Fig. 11.16) As the brakes are applied,
the plunger sleeve and screw move outwards and
the sloping face of the teeth on the adjusting sleeve
lifts the adjusting pawl against the spring. When
the brake is released, the rollers move down both
the central wedge and the two outer plungers'
inclined planes to their fully released position. As
the linings wear, both the plunger strokes and
Fig. 11.15 Twin wedge foundation brake expander and automatic adjuster
Adjustment (Fig. 11.17) Cam adjustment is provided by the hexagon head of the worm shaft
situated on the side of the slack adjuster body
(Fig. 11.17). To adjust the cam relative to the
slack adjuster lever, the lock sleeve is depressed
against the worm lock spring by a suitable spanner
until the worm shaft is free to turn. The worm shaft
is then rotated with the spanner until all the excess
Fig. 11.16 Wedge expander and automatic clearance adjuster
shoe lining to drum clearance is eliminated. The
worm shaft is then prevented from unwinding by
the worm lock spring forcing the lock sleeve against
the hexagonal head of the worm shaft. The removal
of the spanner permits the worm lock spring to push
the internally serrated lock sleeve up to and over the
hexagonal bolt head. To prevent the lock sleeve
rotating, a guide pin fixed in the slack adjuster
body aligns with a slot machined on the sleeve.
Operation (Fig. 11.18)
Automatic adjustment When the foot or hand
brake valve is operated, the brake actuator chamber push rod connected to the slack adjuster will
rotate the slack adjuster body and camshaft in an
anticlockwise direction (Fig. 11.18(a)). The restraining link rod of the ratchet lever which is connected
to some fixed point on the axle will cause the ratchet
lever to rotate clockwise relative to the slack adjuster
body, and thus the ratchet pawl will ride up one of
the teeth of the ratchet wheel.
When adjustment is required, the relative movement between the ratchet lever and the main slack
adjuster body is sufficient for the pawl to ride over
one complete tooth space and engage the next
tooth, following which, on release of the brake
(Fig. 11.18(b)), as the slack adjuster moves to its
`off' position, the ratchet wheel will be rotated by
the amount of the one tooth space. As the ratchet
11.3.8 Automatic slack adjuster
(Fig. 11.18(a, b, c and d))
Purpose Once set up automatic slack adjusters
need no manual adjustment during the life of the
brake linings. Self-adjustment takes place whilst the
brakes are released (when the clearance between
lining and drum exceeds 1.14 mm). This designed
clearance ensures that there is no brake drag and
adequate cooling exists for both shoe and drum.
Fig. 11.17 Manual slack adjuster (Clayton Dewandre)
wheel rotates, so too will the ratchet worm, which
in turn transmits motion to the ratchet worm wheel
and splined main worm shaft. Thus the main
worm, and subsequently the main worm wheel,
are rotated, causing the slack adjuster lever to
take up a new position relative to the camshaft
splined into the main worm wheel.
Adjustment will only take place if the pawl
moves sufficiently around the ratchet wheel to collect the next tooth. This can only occur if the
relative movement between the ratchet lever and
the main slack adjuster body exceeds an angle
of 8.18 .
causes the worm wheel to slide outward along the
splines to the stop washer. A spanner is then
fitted over the square on the end of the ratchet
worm wheel and rotated until the shoe to drum
excess clearance has been taken up. After the
shoes have been manually adjusted, the ratchet
worm wheel should be pushed back until it meshes
with the ratchet worm. The end cap should then be
11.4 Disc brake pad support arrangements
11.4.1 Swing yoke type brake caliper (Fig. 11.19)
This is a lightweight, single cylinder, disc brake
caliper. The caliper unit consists of a yoke made
from a rigid steel pressing, a cylinder assembly, two
pads and a carrier bracket which is bolted to the
suspension hub carrier. The cylinder is attached by
a tongue and groove joint to one side of the yoke
frame whilst the yoke itself is allowed to pivot at
Initial manual adjustment (Fig. 11.18(c)) Manual
adjustment after fitting new shoe linings can be
made by unscrewing the hexagonal cap from
the extended slack adjuster body. This releases the
ratchet worm wheel from the ratchet worm by the
action of the spring behind the worm wheel, which
Fig. 11.18 (a±d)
Automatic slack adjuster
one end on it supporting carrier bracket. The disc is
driven by the transmission drive shaft hub on
which it is mounted and the lining pads are positioned and supported on either side of the disc by
the rectangular aperature in the yoke frame.
reduced respectively, which accordingly moves the
piston forwards and back.
11.4.2 Sliding yoke type brake caliper
(Fig. 11.20)
With this type of caliper unit, the cylinder body is
rigidly attached to the suspension hub carrier,
whereas the yoke steel pressing fits over the cylinder body and is permitted to slide between parallel
grooves formed in the cylinder casting.
Operation (Fig. 11.19) When the foot brake is
applied generated hydraulic pressure pushes the
piston and inboard pad against their adjacent disc
face. Simultaneously, the hydraulic reaction will
move the cylinder in the opposite direction away
from the disc. Consequently, as the outboard pad
and cylinder body are bridged by the yoke, the
latter will pivot, forcing the outboard pad against
the opposite disc face to that of the inboard pad.
As the pads wear, the yoke will move through an
arc about its pivot, and to compensate for this tilt
the lining pads are taper shaped. During the wear
life of the pad friction material, the amount of
taper gradually reduces so that in a fully worn
state the remaining friction material is approximately parallel to the steel backing plate.
The operating clearance between the pads and
disc is maintained roughly constant by the inherent
distortional stretch and retraction of the pressure
seals as the hydraulic pressure is increased and
Operation (Fig. 11.20) When the foot brake is
applied, hydraulic pressure is generated between
the two pistons. The hydraulic pressure pushes
the piston apart, the direct piston forces the direct
pad against the disc whilst the indirect piston forces
the yoke to slide in the cylinder in the opposite
direction until the indirect pad contacts the outstanding disc face.
Further pressure build-up causes an equal but
opposing force to sandwich the disc between the
friction pads. This pressure increase continues until
the desired retardation force is achieved.
During the pressure increase the pressure seals distort as the pistons move apart. When the hydraulic
pressure collapses the rubber pressure seals retract
Fig. 11.19 Swing yoke type brake caliper
Fig. 11.20 Sliding yoke type brake caliper
and withdraw the pistons and pads from the disc
surface so that friction pad drag is eliminated.
Yoke rattle between the cylinder and yoke frame
is reduced to a minimum by inserting either a wire
or leaf type spring between the sliding joints.
As with all other types of caliper units, pad to
disc free clearance is obtained by the pressure seals
which are fitted inside recesses in the cylinder wall
and grip the piston when hydraulic pressure forces
the piston outwards, causing the seal to distort.
When the brakes are released and the pressure is
removed from the piston crown, the strain energy
of the elastic rubber pulls back the piston until the
pressure seal has been restored to its original shape.
11.4.3 Sliding pin type brake caliper (Fig. 11.21)
The assembled disc brake caliper unit comprises
the following; a disc, a carrier bracket, a cylinder
caliper bridge, piston and seals, friction pads and
a pair of support guide pins.
The carrier bracket is bolted onto the suspension
hub carrier, its function being to support the cylinder caliper bridge and to absorb the brake torque
The cylinder caliper bridge is mounted on a pair
of guide pins sliding in matched holes machined in
the carrier bracket. The guide pins are sealed
against dirt and moisture by dust covers so that
equal frictional sliding loads will be maintained at
all times. On some models a rubber bush sleeve is
fitted to one of the guide pins to prevent noise and
to take up brake deflection.
Frictional drag of the pads is not taken by the
guide pins, but is absorbed by the carrier bracket.
Therefore the pins only support and guide the
caliper cylinder bridge.
Operation (Fig. 11.21) When the foot brake is
applied, the hydraulic pressure generated pushes
the piston and cylinder apart. Accordingly the
inboard pad moves up to the inner disc face,
whereas the cylinder and bridge react in the opposite sense by sliding the guide pins out from their
supporting holes until the outboard pad touches
the outside disc face. Further generated hydraulic
pressure will impose equal but opposing forces
against the disc faces via the pads.
11.4.4 Sliding cylinder body type brake caliper
(Fig. 11.22)
This type of caliper unit consists of a carrier
bracket bolted to the suspension hub carrier and
a single piston cylinder bridge caliper which straddles
Fig. 11.21 Slide pin type brake caliper
the disc and is allowed to slide laterally on guide
keys positioned in wedge-shaped grooves machined
in the carrier bracket.
11.4.5 Twin floating piston caliper disc brake
with hand brake mechanism (Fig. 11.23)
This disc brake unit has a pair of opposing pistons
housed in each split half-caliper. The inboard halfcaliper is mounted on a flanged suspension hub
carrier, whereas the other half straddles the disc
and is secured to the rotating wheel hub. Lining
pads bonded to steel plates are inserted on each
side of the disc between the pistons and disc rubbing face and are held in position by a pair of steel
pins and clips which span the two half-calipers.
Brake fluid is prevented from escaping between
the pistons and cylinder walls by rubber pressure
seals which also serve as piston retraction springs,
while dirt and moisture are kept out by flexible
rubber dust covers.
Operation (Fig. 11.22) When the foot brake is
applied, the generated hydraulic pressure enters
the cylinder, pushing the piston with the direct
acting pad onto the inside disc face. The cylinder
body caliper bridge is pushed in the opposite direction. As a result, the caliper bridge reacts and slides
in its guide groove at right angles to the disc until
the indirect pad contacts the outside disc face,
thereby equalling the forces acting on both sides
of the disc.
A pad to disc face working clearance is provided
when the brakes are released by the retraction of
the pressure seal, drawing the piston a small
amount back into the cylinder after the hydraulic
pressure collapes.
To avoid vibration and noise caused by relative
movement between the bridge caliper and carrier
bracket sliding joint, anti-rattle springs are normally incorporated alongside each of the twoedge-shaped grooves.
Foot brake application (Fig. 11.23) Hydraulic
pressure, generated when the foot brake is applied,
is transferred from the inlet port to the central halfcaliper joint, where it is then transmitted along
passages to the rear of each piston.
As each piston moves forward to take the clearance between the lining pads and disc, the piston
Fig. 11.22 Slide cylinder body brake caliper
Fig. 11.23 Twin floating piston caliper disc brake with hand brake mechanism
pressure seals are distorted. Further pressure buildup then applies an equal but opposite force by way
of the lining pads to both faces of the disc, thereby
creating a frictional retarding drag to the rotating
disc. Should the disc be slightly off-centre, the pistons will compensate by moving laterally relative to
the rubbing faces of the disc.
Releasing the brakes causes the hydraulic pressure
to collapse so that the elasticity within the distorted
rubber pressure seals retracts the pistons and pads
until the seals convert to their original shape.
The large surface area which is swept on each
side of the disc by the lining pads is exposed to
the cooling airstream so that heat dissipation is
Foot brake application (Fig. 11.24(a)) When the
hydraulic brakes are applied, the piston outward
movement is approximately equal to the predetermined clearance between the piston and nut with
the brakes off, but as the pads wear, the piston
takes up a new position further outwards, so that
the normal piston to nut clearance is exceeded.
If there is very little pad wear, hydraulic pressure
will move the piston forward until the pads grip the
disc without the thrust washer touching the ball
race. However, as the pads wear, the piston
moves forward until the thrust washer contacts
the ball race. Further outward movement of the
piston then forces the thrust washer ball race and
shouldered nut together in an outward direction.
Since the threaded shaft is prevented from rotating
by the strut and cam, the only way the nut can
move forward is by unwinding on the screw shaft.
Immediately the nut attempts to turn, the coil
spring uncoils and loses its grip on the nut, permitting the nut to screw out in proportion to the piston
On releasing the foot brake, the collapse of the
hydraulic pressure enables the pressure seals to
withdraw the pads from the disc. Because the
axial load has been removed from the nut, there is
no tendency for it to rotate and the coil spring
therefore contracts, gripping the nut so that it cannot rotate. Note that the outward movement of the
nut relative to the threaded shaft takes up part of
the slack in the mechanical linkage so that the hand
brake lever movement remains approximately constant throughout the life of the pads. The threaded
shaft and nut device does not influence the operating pad to disc clearance when the hydraulic brakes
are applied as this is controlled only by the pressure
seal distortion and elasticity.
Hand brake application (Fig. 11.23) The hand
brake mechanism has a long and short clamping
lever fitted with friction pads on either side of the
disc and pivots from the lower part of the caliper. A
tie rod with an adjusting nut links the two clamping
levers and, via an operating lever, provides the
means to clamp the disc between the friction
pads. Applying the hand brake pulls the operating
lever outwards via the hand brake cable, causing
the tie rod to pull the short clamp lever and pad
towards the adjacent disc face, whilst the long
clamp and pad is pushed in the opposite direction
against the other disc face. As a result, the lining
pads grip the disc with sufficient force to prevent
the car wheels rolling on relatively steep slopes.
To compensate for pad wear, the adjustment nut
should be tightened periodically to give a maximum
pad to disc clearance of 0.1 mm.
11.4.6 Combined foot and hand brake caliper
with automatic screw adjustment (Bendix)
This unit provides automatic adjustment for the
freeplay in the caliper's hand brake mechanism
caused by pad wear. It therefore keeps the hand
brake travel constant during the service life of the
The adjustment mechanism consists of a shouldered nut which is screwed onto a coarsely
threaded shaft. Surrounding the nut on one side
of the shoulder or flange is a coiled spring which is
anchored at its outer end via a hole in the piston.
On the other side of the shouldered nut is a ball
bearing thrust race. The whole assembly is enclosed
in the hollow piston and is prevented from moving
out by a thrust washer which reacts against the
thrust bearing and is secured by a circlip to the
interior of the piston.
Hand brake application (Fig. 11.24(b)) Applying
the hand brake causes the cable to rotate the camshaft via the cam lever, which in turn transfers
force from the cam to the threaded shaft through
the strut. The first part of the screwed shaft travel
takes up the piston to nut end-clearance. With
further screw shaft movement the piston is pushed
outwards until the pad on the piston contacts the
adjacent disc face. At the same time an equal and
opposite reaction causes the caliper cylinder to
move in the opposite direction until the outside
pad and disc face touch. Any further outward
movement of the threaded shaft subsequently
clamps the disc in between the pads. Releasing the
hand brake lever relaxes the pad grip on the disc
Fig. 11.24 (a and b)
Combined foot and hand brake caliper with automatic screw adjustment
with the assistance of the Belleville washers which
draws back the threaded shaft to the `off' position
to avoid the pads binding on the disc.
11.5.1 Front to rear brake line split
(Fig. 11.25(a))
With this arrangement, the two separate hydraulic
pipe lines of the tandem master cylinder are in
circuit with either both the front or rear caliper or
shoe expander cylinders. The weakness with this
pipe line split is that roughly two-thirds of the
braking power is designed to be absorbed by the
front calipers, and only one-third by the rear
brakes. Therefore if the front brakes malfunction,
the rear brake can provide only one-third of the
original braking capacity.
11.5 Dual- or split-line braking systems
Dual- or split-line braking systems are used on all
cars and vans to continue to provide some degree
of braking if one of the two hydraulic circuits
should fail. A tandem master cylinder is incorporated in the dual-line braking system, which is
in effect two separate master cylinder circuits
placed together end on so that it can be operated
by a common push rod and foot pedal. Thus, if
there is a fault in one of the hydraulic circuits, the
other pipe line will be unaffected and therefore will
still actuate the caliper or drum brake cylinders it
11.5.2 Diagonally front to rear brake split
(Fig. 11.25(b))
To enable the braking effort to be more equally
shared between each hydraulic circuit (if a fault
should occur in one of these lines), the one front
11.5.4 Compensating port type tandem master
cylinder (Fig. 11.26(a±d))
Tandem master cylinders are employed to operate
dual-line hydraulic braking systems. The master
cylinder is composed of a pair of pistons functioning within a single cylinder. This enables two independent hydraulic cylinder chambers to operate.
Consequently, if one of these cylinder chambers
or part of its hydraulic circuit develops a fault,
the other cylinder chamber and circuit will still
continue to effectively operate.
Brakes off (Fig. 11.26(a)) With brakes in the `off'
position, both primary and secondary pistons are
pushed outwards by the return springs to their
respective stops. Under these conditions fluid is
permitted to circulate between the pressure chambers and the respective piston recesses via the small
compensating port, reservoir supply outlet and the
large feed ports for both primary and secondary
brake circuits.
Fig. 11.25 (a±c)
Brakes applied (Fig. 11.26(b)) When the foot
pedal is depressed, the primary piston moves
inwards and, at the same time, compresses both
the intermediate and secondary return springs so
that the secondary piston is pushed towards the
cylinder's blanked end.
Initial movement of both pistons causes their
respective recuperating seals to sweep past each
compensating port. Fluid is trapped and, with
increased piston travel, is pressurized in both the
primary and secondary chambers and their pipe
line circuits, supplying the front and rear brake
cylinders. During the braking phase, fluid from
the reservoir gravitates and fills both of the annular
piston recesses.
Dual- or split-line braking systems
and one diagonally opposite rear wheel are connected together. Each hydraulic circuit therefore
has the same amount of braking capacity and the
ratio of front to rear braking proportions do not
influence the ability to stop. A diagonal split also
tends to retard a vehicle on a relatively straight line
on a dry road.
11.5.3 Triangular front to rear brake split
(Fig. 11.25(b))
This hydraulic pipe line system uses front calipers
which have two independent pairs of cylinders,
and at the rear conventional calipers or drum
brakes. Each fluid pipe line circuit supplies half
of each front caliper and one rear caliper or
drum brake cylinder. Thus a leakage in one or
the other hydraulic circuits will cause the other
three pairs of calipers or cylinders or two pairs of
caliper cylinders and one rear drum brake cylinder to provide braking equal to about 80% of
that which is possible when both circuits are
operating. When one circuit is defective, braking
is provided on three wheels; it is then known as
a triangular split.
Brakes released (Fig. 11.26(a)) When the foot
pedal effort is removed, the return springs rapidly
expand, pushing both pistons outwards. The speed
at which the swept volume of the pressure chambers increases will be greater than the rate at which
the fluid returns from the brake cylinders and pipe
lines. Therefore a vacuum is created within both
primary and secondary pressure chambers.
As a result of the vacuum created, each recuperating seal momentarily collapses. Fluid from the
annular piston recess is then able to flow through
the horizontal holes in the piston head, around the
inwardly distorted recuperating seals and into their
respective pressure chambers. This extra fluid
Fig. 11.26 (a±d)
Tandem master cylinder
entering both pressure chambers compensates for
any fluid loss within the brake pipe line circuits or
for excessive shoe to drum clearance. But, if too
much fluid is induced in the chambers, some of this
fluid will pass back to the reservoir via the compensating ports after the return springs have fully
retracted both pistons.
move inwards until the primary piston abuts the
secondary spring retainer. Further pedal effort will
move the secondary piston recuperating seal
beyond the compensating port, thereby pressurizing the fluid in the secondary chamber and subsequently transmitting this pressure to the secondary
circuit pipe line and the respective brake cylinders.
Failure in the primary circuit (Fig. 11.26(c))
Should a failure (leakage) occur in the primary
circuit, there will be no hydraulic pressure generated in the primary chamber. When the brake pedal
is depressed, the push rod and primary piston will
Failure in the secondary circuit (Fig. 11.26(d)) If
there is a failure (leakage) in the secondary circuit,
the push rod will move the primary piston inwards
until its recuperating seal sweeps past the compensating port, thus trapping the existing fluid
in the primary chamber. Further pedal effort
increases the pressure in the primary chamber and
at the same time both pistons, separated by the
primary chamber fluid, move inwards unopposed
until the secondary piston end stop contacts the
cylinder's blanked end. Any more increase in braking effort raises the primary chamber pressure,
which accordingly pressurizes the primary circuit
brake cylinders.
The consequence of a failure in the primary or
secondary brake circuit is that the effective push
rod travel increases and a greater pedal effort will
need to be applied for a given vehicle retardation
compared to a braking system which has both
primary and secondary circuits operating.
primary and secondary pipe line circuits to rise
and operate the brake cylinders.
Brakes released (Fig. 11.27(a)) Removing the
foot from the brake pedal permits the return spring
to push both pistons to their outermost position.
The poppet valve stem instantly contacts their
respective roll pins, causing both valves to open.
Since the return springs rapidly push back their
pistons, the volume increase in both the primary
and secondary chambers exceeds the speed of the
returning fluid from the much smaller pipe line
bore, with the result that a depression is created
in both chambers. Fluid from the reservoir flows
via the elongated slot and open poppet valve into
the primary and secondary chambers to compensate for any loss of fluid or excessive shoe to drum
or pad to disc clearance. This method of transferring fluid from the reservoir to the pressure chamber is more dynamic than relying on the collapse
and distortion of the rubber pressure seals as in the
conventional master cylinder.
Within a very short time the depression disappears and fluid is allowed to flow freely to and
fro from the pressure chambers to compensate for
fluid losses or fluid expansion and contraction
caused by large temperature changes.
11.5.5 Mecanindus (roll) pin type tandem
master cylinder incorporating a pressure
differential warning actuator (Fig. 11.27(a±d))
The tandem or split master cylinder is designed to
provide two separate hydraulic cylinder pressure
chambers operated by a single input push rod.
Each cylinder chamber is able to generate its own
fluid pressure which is delivered to two independent brake pipe line circuits. Thus if one hydraulic
circuit malfunctions, the other one is unaffected
and will provide braking to the wheel cylinders
forming part of its system.
11.5.6 Operation of the pressure differential
warning actuator
As a warning to the driver that there is a fault in
either the primary or secondary hydraulic braking
circuits of a dual-line braking system, a pressure
differential warning actuator is usually incorporated as an integral part of the master cylinder or
it may be installed as a separate unit (Fig. 11.27).
The switch unit consists of a pair of opposing
balance pistons spring loaded at either end so that
they are normally centrally positioned. Mounted
centrally and protruding at right angles into the
cylinder is an electrical conducting prod, insulated
from the housing with a terminal formed at its
outer end. The terminal is connected to a dashboard warning light and the electrical circuit is
completed by the earth return made by the master
Operation of tandem master cylinder
Brakes off (Fig. 11.27(a)) With the push rod fully
withdrawn, both primary and secondary pistons
are forced outwards by the return springs. This
outward movement continues until the central
poppet valve stems contact their respective
Mecanindus (roll) pins. With further withdrawal
the poppet valves start opening until the front end
of each elongated slot also contacts their respective
roll pins, at which point the valves are fully open.
With both valves open, fluid is free to flow between
the primary and secondary chambers and their
respective reservoirs via the elongated slot and
vertical passage in the roll pins.
Brakes applied (Fig. 11.27(b)) When the brake
pedal is applied, the push rod and the primary
return spring pushes both pistons towards the
cylinder's blank end. Immediately both recuperating poppet valves are able to snap closed. The fluid
trapped in both primary and secondary chambers
is then squeezed, causing the pressure in the
Operation (Fig. 11.27(b)) If, when braking, both
hydraulic circuits operate correctly, the opposing
fluid pressure imposed on the outer ends of the
balance piston will maintain the pistons in their
equilibrium central position.
Fig. 11.27 (a±d)
Tandem master cylinder with pressure differential warning actuator
Should one or the other of the dual circuits
develop a pressure drop fault due to fluid leakage
(Fig. 11.27(c and d)), then if the pressure difference
of 10 bar or more exists between the two circuits,
an imbalance of the fluid pressure applied against
the outer ends of the pistons will force both pistons
to move in the direction of the faulty circuit. The
sideways movement of the pistons will cause the
shoulder of the correctly operating circuit balance
piston to contact the protruding prod, thus automatically completing the dashboard warning light
electrical circuit, causing it to illuminate. Removing the brake pedal effort causes the fluid pressure
in the effective circuit to collapse, thereby enabling
the balance pistons to move back to their central
positions. This interrupts the electrical circuit so
that the warning light switches off.
Operation (Figs 11.28 and 11.29) Under light
brake pedal application, fluid pressure from the
master cylinder enters the valve inlet port and
passes through the centre and around the outside
of the plunger on its way to the outlet ports via the
wasted region of the plunger (Fig. 11.28(a)).
When heavy brake applications are made (Fig.
11.28(b)), the rising fluid pressure acting on the
large passage at the rear of the plunger displaces
the plunger assembly. Instantly the full crosssectional area equivalent to the reaction piston is
exposed to hydraulic pressure, causing the plungers
to move forward rapidly until the plunger end seal
contacts the valve seat in the body of the valve unit.
The valve closing pressure is known as the cut-off
pressure. Under these conditions the predetermined
line pressure in the rear pipe line will be maintained
constant (Fig. 11.29), whereas the front brake pipe
line pressure will continue to rise unrestricted,
according to the master cylinder pressure generated
by the depressed brake pedal.
11.6 Apportioned braking
11.6.1 Pressure limiting valve (Fig. 11.28)
The object of the pressure limiting valve is to interrupt the pressure rise of fluid transmitted to the
rear wheel brakes above some predetermined
value, so that the rear brakes will be contributing
a decreasing proportion of the total braking with
further increased pedal effort and master cylinder
generated line pressure. By imposing a maximum
brake line pressure to the rear brake cylinders, the
rear wheels will be subjected to far less overbraking
when the vehicle is heavily braked. It therefore
reduces the tendency for rear wheel breakaway
caused by wheels locking. Note that with this type
of valve unit under severe slippery conditions the
rear wheels are still subjected to lock-up.
Fig. 11.28 (a and b)
11.6.2 Load sensing pressure limiting valve
(Fig. 11.20)
To take into account the weight distribution
between the front and rear wheels between an
unladen and fully laden vehicle, a load sensing
valve may be incorporated in the pipe line connecting
the master cylinder to the rear wheel brakes. The
function of the valve is to automatically separate
the master cylinder to rear brake pipe line by closing
a cut-off valve when the master cylinder's generated
pressure reaches some predetermined maximum.
This cut-off pressure will vary according to the
weight imposed on the rear axle.
Pressure limiting valve
stiffen and increase the inward end thrust imposed
on the plunger.
With a light brake pedal application, fluid pressure generated by the master cylinder enters the inlet
port and passes around the wasted plunger on its
way out to the rear brake pipe line (Fig. 11.30(a)).
If a heavy brake application is made (Fig. 11.30(b)),
the rising fluid pressure from the master cylinder
passes through the valve from the inlet to the outlet
ports until the pressure creates a force at the end of
the plunger (Force ˆ Pressure Area) which opposes
the spring thrust, pushing back the plunger until
the face valve closes. Any further fluid pressure
rise will only be transmitted to the front brake
pipe lines, whereas the sealed-off fluid pressure in
the rear brake pipe lines remains approximately
If the load on the rear axle alters, the vertical
deflection height of the suspension will cause the
leaf spring to stiffen or relax according to any axle
load increase or decrease.
A change in leaf spring tension therefore alters
the established pressure (Fig. 11.31) (at which point
the cut-off valve closes) and the maximum attainable pressure trapped in the rear brake pipe lines.
Fig. 11.29 Pressure limiting valve front to rear brake line
Operation (Figs 11.30 and 11.31) This valve
device consists of a plunger supporting a rubber
face valve which is kept open by the tension of
a variable rate leaf spring. The inward thrust on the
plunger keeping the end face valve open is determined by the leaf spring pre-tension controlled by
the rear suspension's vertical deflection via the
interconnecting spring and rod link. When the
vehicle's rear suspension is unloaded, the leaf
spring will be partially relaxed, but as the load on
the rear axle increases, the link spring and rod pulls
the leaf spring towards the valve causing it to
Fig. 11.30 (a and b)
11.6.3 Load sensing progressive pressure limiting
valve (Fig. 11.32)
The load sensing progressive pressure limiting
valve regulates the fluid pressure transmitted to
the rear brake cylinders once the master cylinder's
generated pressure has risen above some predetermined value corresponding to the weight carried on
the rear axle.
Load sensing pressure limiting valve
the centre of the stepped reaction piston, between
the cone and seat and to the rear brake pipe lines.
If the brake pedal is further depressed (Fig.
11.32(b)), increased fluid pressure acting on the
large piston area produces an end force, which,
when it exceeds the opposing link spring tension
and fluid pressure acting on the annular piston
face, causes the stepped piston to move outwards.
This outward movement of the piston continues
until the valve stem clears the cylinder's blanked
end, thereby closing the valve. The valve closure is
known as the cut-off point since it isolates the rear
brake pipe lines from the master cylinder delivery.
Further generation of master cylinder pressure
exerted against the annular piston face produces an
increase in force which moves the piston inwards,
once again opening the valve. The hydraulic connection is re-established, allowing the rear brake
pipe line fluid pressure to increase. However, the
pressure exerted against the end face of the piston
immediately becomes greater than the spring force
and hydraulic force pushing on the annular piston
face, and so the piston moves outwards, again
closing the valve.
Every time the valve is opened with rising master
cylinder pressure, the rear brake pipe line pressure
increases in relation to the previous closing of the
valve. Over a heavy braking pressure rise phase the
piston oscillates around a position of balance,
causing a succession of valve openings and closings. It subsequently produces a smaller pressure
rise in the rear brake pipe line than with the directly
connected front brake pipe lines.
Fig. 11.31 Load sensing pressure limiting valve front to
rear brake line characteristics
The reduced rate of pressure increase, in proportion to the pedal effort in the rear brake pipe line,
provides a braking ability for both the front and
rear brakes which approximately matches the load
distribution imposed on the front and rear wheels,
so that the tendency for the rear wheels to be either
under or over braked is considerably reduced.
Operation (Fig. 11.32) When the foot pedal is
applied lightly (Fig. 11.32(a)), pressure generated
by the master cylinder will be transferred through
Fig. 11.32 (a and b)
Load sensing progressive pressure limiting valve
11.6.4 Inertia pressure limiting valve (Fig. 11.34)
The inertia pressure limiting valve is designed to
restrict the hydraulic line pressure operating the
rear wheel brakes when the deceleration of the
vehicle exceeds about 0.3 g. In preventing a further
rise in the rear brake line pressure, the unrestricted
front brake lines will, according to the hydraulic
pressure generated, increase their proportion of
braking relative to the rear brakes.
Operation (Figs 11.34 and 11.35) The operating
principle of the inertia valve unit relies upon the
inherent inertia of the heavy steel ball rolling up an
inclined ramp when the retardation of the vehicle
exceeds some predetermined amount (Fig. 11.34(b)).
When this happens, the weight of the ball is
removed from the stem of the disc valve, enabling
the return spring fitted between the inlet port and
valve shoulder to move the valve into the cut-off
At this point, the fluid trapped in the rear brake
pipe line will remain constant (Fig. 11.35), but fluid
flow between the master cylinder and front brakes
is unrestricted and therefore will continue to rise
with increased pedal force. As a result, the front
brakes will contribute a much larger proportion of
the total braking effort than the rear brakes.
When the vehicle has slowed down sufficiently or
even stopped, the steel ball will gravitate to its lowest
point, thereby pushing open the cut-off valve. Fluid
is now free again to move from the master cylinder
to the rear wheel brakes (Fig. 11.34(a)).
Fig. 11.33 Load sensing and progressive pressure
limiting valve front to rear brake line characteristics
The ratio of the stepped piston face areas
determines the degree of rear brake pipe line
increase with respect to the front brake pipe lines
(Fig. 11.33).
The cut-off or change point depends on the tensioning of the pre-setting spring which varies with
the rear suspension deflection. The brake force
distribution between the front and rear brakes is
not only affected by the static laden condition, but
even more so by the dynamic weight transference
from the rear to the front axle.
Fig. 11.34 (a and b)
Inertia pressure limiting valve
rolling up the inclined ramp until it seals off the
central piston passage. This state is known as the
point of cut-off.
Further foot pedal effort directly increases the
front brake pipe line pressure and the pressure in
the ball chamber. It does not immediately increase
the rear brake pipe line pressure on the output of
the valve.
Under these conditions, the trapped cut-off
pressure in the rear brake lines reacts against the
large piston cross-sectional area, whereas the small
piston cross-sectional area on the ball chamber side
of the piston is subjected to the master cylinder
hydraulic pressure.
As the master cylinder's generated pressure
rises with greater foot pedal effort, the input force
produced on the small piston side (Input force ˆ
Master cylinder pressure Small piston area) will
increase until it exceeds the opposing output force
produced on the large piston area (Output force ˆ
Rear brake line pressure Large piston area).
A further rise in master cylinder pressure which
will also be experienced in the ball chamber pushes
the stepped piston backwards. Again, the rear
brake line pressure will start to rise (Fig. 11.35),
but at a reduced rate determined by the ratio of the
small piston area to large piston area, i.e. AS/AL.
For example, if the piston area ratio is 2:1, then the
rear brake line pressure increase will be half the
input master cylinder pressure rise.
Fig. 11.35 Inertia pressure limiting valve and inertia
progressive pressure limiting front to rear brake line
11.6.5 Inertia and progressive pressure limiting
valve (Fig. 11.36)
The inertia and progressive pressure limiting valve
unit enables the braking power between the front
and rear wheels to be adjusted to match the weight
transference from the rear wheels to the front
wheels in proportion to the vehicle's deceleration
rate. This two stage valve unit allows equal fluid
pressure to flow between the front and rear brakes
for light braking, but above some predetermined
deceleration of the vehicle, the direct pressure
increase to the rear brakes stops. With moderate
to heavy braking, the front brake line pressure will
equal the master cylinder generated pressure. The
rear brake line pressure will continue to increase
but at a much slower rate compared to that of the
front brakes.
Pi ˆ input pressure
Po ˆ output pressure
i:e: Po ˆ
To safeguard the rear brake pipe lines, should
the piston reach its full extent of its travel, the
centre pin will stand out from the piston. Consequently the ball is dislodged from its seat so
that fluid pressure is permitted to pass to the rear
brake pipe lines.
If there are two separate rear brake pipe line
circuits, each line will have its own rear brake
pressure reducing valve.
Operation (Figs 11.35 and 11.36) This inertia and
progressive valve unit differs from the simple inertia pressure limiting valve because it incorporates a
stepped piston (two piston dimeters) and the ball
performs the task of the cut-off valve.
If the vehicle is lightly braked (Fig. 11.36(a)),
fluid will flow freely from the master cylinder inlet
port, through the dispersing diffuser, around the
ball, along the piston central pin passage to the
outlet port leading to the rear wheel brakes.
As the brake pedal force is increased (Fig.
11.36(b)), the vehicle's rate of retardation will
cause the ball to continue to move forward by
11.7 Antilocking brake system (ABS)
With conventional brake systems one of the road
wheels will always tend to lock sooner than the
other, due to the continuously varying tyre to
road grip conditions for all the road wheels. To
prevent individual wheels locking when braking,
the pedal should not be steadily applied but it
Fig. 11.36 (a±d)
Inertia and progressive pressure limiting valve
should take the form of a series of impulses caused
by rapidly depressing and releasing the pedal. This
technique of pumping and releasing the brake
pedal on slippery roads is not acquired by every
driver, and in any case is subjected to human error
in anticipating the pattern of brake pedal application to suit the road conditions. An antilock brake
system does not rely on the skill of the driver to
control wheel lock, instead it senses individual
wheel slippage and automatically superimposes a
brake pipe line pressure rise and fall which counteracts any wheel skid tendency and at the same time
provides the necessary line pressure to retard the
vehicle effectively.
When no slip takes place between the wheel and
road surface, the wheel's circumference (periphery)
speed and the vehicle's speed are equal. If, when the
brakes are applied, the wheel circumference speed
is less than the vehicle speed, the speed difference is
the slip between the tyre and road surface. When
the relative speeds are the same the wheels are in a
state of pure rolling. When the wheels stop rotating
with the vehicle continuing to move forward the
slip is 100%, that is, the wheel has locked.
To attain optimum brake retardation of the
vehicle, a small amount of tyre to ground slip is
necessary to provide the greatest tyre tread to road
surface interaction. For peak longitudinal braking
force an approximately 15% wheel slip is necessary (Fig. 11.37), whereas steerability when braking depends upon a maximum sideways tyre to
ground resistance which is achieved only with the
mines when the front wheel is approaching
a predetermined deceleration. In response to this
the modulator reduces the pressure in the respective brake circuits. When the wheel speeds up again,
the pump raises that pressure in order to bring the
braking force back to a maximum level. This
sequence of pressure reduction and build-up can
be up to five times a second to avoid the wheel
locking and also to provide the necessary deceleration of the car.
Braking as normal (Fig. 11.39(a)) Under normal
braking conditions, the master cylinder fluid output is conveyed to the wheel brakes through the
open cut-off valve. The dump valve is closed and
the pump piston is held out of engagement from the
rotating eccentric cam by the return spring.
Fig. 11.37 Relationship of braking force coefficient and
wheel slip
Brake pressure reducing (Fig. 11.39(b)) When the
deceleration of the front wheel, and therefore the
drive shaft, exceeds a predetermined maximum
(the wheels begin to lock), the flywheel overruns
the drive shaft due to its inertia. The clutch balls
then roll up their respective ramps, forcing the
flywheel to slide inwards and causing the dump
valve lever to tilt and open the dump valve. The
fluid pressure above the deboost piston drops
immediately. The much higher brake line pressure
underneath the deboost piston and the pump piston forces the pump piston against its cam and
raises the deboost piston. Fluid above both pistons
is displaced back to the reservoir via the dump
valve. The effect of the deboost piston rising is
to close the ball cut-off valve so that the master
cylinder pipe line fluid output and the wheel cylinder pipe line input become isolated from each
other. As a result, the sealed chamber space below
the deboost piston is enlarged, causing a rapid
reduction in the fluid pressure delivered to the
wheel cylinders and preventing the wheels connected to this brake circuit locking.
minimum of slip (Fig. 11.37). Thus there is conflict
between an increasing braking force and a decreasing sideways resistance as the percentage of wheel
slip rises initially. As a compromise, most anti-skid
systems are designed to operate within an 8±30%
wheel slip range.
11.7.1 Hydro-mechanical antilock brake system
(ABS) suitable for cars (SCS Lucas Girling)
(Figs 11.38 and 11.39)
This hydro-mechanical antilock braking system
has two modular units, each consisting of an integrated flywheel decelerating sensor, cam operated
piston type pump and the brake pressure modulator itself (Fig. 11.38). Each modulator controls the
adjacent wheel brake and the diagonally opposite
rear wheel via an apportioning valve. The modular
flywheel sensor is driven by a toothed belt at 2.8
times the wheel speed. The flywheel sensor deter-
Brake pressure increasing (Fig. 11.39(c)) The
pressure reduction resulting from the previous
phase releases the brakes and allows the wheel to
accelerate to the speed of the still decelerating flywheel. When the drive shaft and the flywheel are at
roughly equal speeds, the clutch balls roll down
their respective ramps, enabling the dump valve
lever return spring to slide the flywheel over. The
dump valve lever then pivots and closes the needletype dump valve. The flywheel is again coupled to
Fig. 11.38 Stop control braking system (SCS) layout
according to the loading imposed on the rear axle,
to the rear wheel service chamber actuators. Compressed air is also delivered to both the service and
the emergency line couplings via the relay valve and
the pressure protection valve. This therefore safeguards the tractor air supply should there be a hose
failure between the tractor and trailer. A differential protection valve is installed between the service
line and the secondary/park line to prevent both
systems operating simultaneously which would
overload the foundation brakes.
control valve lever to the `on' position progressively
reduces the secondary/park line pressure going to
the spring brake. The secondary line pressure going
to the trailer coupling increases, thereby providing
a tractor to trailer brake match. Moving the hand
control valve to the `park' position exhausts the air
from the trailer secondary line and the spring brake
secondary/park line. The tractor foundation brakes
are then applied by the thrust exerted by the power
spring within the actuator alone. The release of the
parking brake is achieved by delivering air to the
spring brake when the hand control valve is moved
to the `off' position again.
Secondary/park line circuit (Fig. 12.4) Air is supplied from the secondary/park reservoir to the
hand control valve and to a pair of relay valves.
One relay valve controls the air delivered to the
spring brake actuator, the other controls the service line air supply to the trailer brakes. With the
hand control valve in the `off' position, air is
delivered through the secondary/park line relay valve
to the spring brakes. The secondary/park spring
brakes are held in the released position due to the
compression of each power spring within the actuator. As the spring brakes are being released, the
secondary line to the trailer is exhausted of compressed air via its relay valve. Moving the hand
12.2.5 Towing truck or tractor spring brake
two line system (Fig. 12.5)
Compressed air supply (Fig. 12.5) The air supply
from the compressor passes through the air dryer
on its way to the multi-circuit protection. The output air supply is then shared between four reservoirs; two service, one trailer and one secondary/
park reservoirs.
Service line circuit (Fig. 12.5) The air delivered
to the service line wheel actuator chambers is
Fig. 12.5 Towing truck or spring brake two line system
provided by a dual foot valve which splits the
service line circuits between the tractor's front and
rear wheels. Therefore, if one or other service line
circuit should develop a fault, the other circuit
with its own reservoir will still function. At the
same time as the tractor service brakes are applied,
a signal pressure from the foot valve passes to the
multi-relay valve. This opens an inlet valve which
permits air from the trailer reservoir to flow to the
control line (service line Ð yellow) trailer coupling.
To prevent both service line and secondary/park
line supplies compounding, that is, operating at the
same time, and overloading the foundation brakes,
a differential protection valve is included for both
the front and rear axle brakes.
should one malfunction, so that trailer braking is
still provided. The multi-relay valve also enables
the hand control valve to operate the trailer brakes
so that the valve is designed to cope with three
separate signals; the two service line pressure signals controlled by the dual foot valve and the hand
valve secondary pressure signal.
Supply dump valve (Fig. 12.26(a, b and c)) The
purpose of the supply dump valve is to automatically reduce the trailer emergency line pressure to
1.5 bar should the trailer service brake line fail after
the next full service brake application within two
seconds. This collapse of emergency line pressure
signals to the trailer emergency valve to apply the
trailer brakes from the trailer reservoir air supply,
overriding the driver's response.
Secondary/park line circuit (Fig. 12.5) A secondary braking system which incorporates a parking
brake is provided by spring brakes which are
installed on both front and rear axles. Control of
the spring brakes is through a hand valve which
provides an inverse signal to the multi-relay valve
so that the trailer brakes can also be applied by the
hand control valve.
With the hand control valve in the `off' position
the secondary line from the hand valve to the multirelay valve, and the secondary/park line, also from
the hand valve, going to the spring brake actuators
via the differential protection valves, are both
pressurized. This compresses the power springs,
thereby releasing the spring brakes. During this
period no secondary line pressure signal is passed
to the trailer brakes via the multi-relay valve.
When the hand valve is moved towards the
`applied' position, the secondary line feeding the
multi-relay valve and the secondary/park line
going to the spring brakes reduces their pressures
so that both the tractor's spring brakes and the
trailer brakes are applied together in the required
tractor to trailer proportions.
Moving the hand valve lever to the `park' position exhausts the secondary/park line going to the
spring brakes and pressurizes the secondary line
going to the multi-relay valve. As a result, the
power springs within the spring actuators exert
their full thrust against the foundation brake cam
lever and at the same time the trailer control line
(service line) is exhausted of compressed air. Thus
the vehicle is held stationary solely by the spring
12.2.6 Trailer two line brake system (Fig. 12.6)
The difference with the two and three line trailer
braking systems is that the two line only has a
single control service line, whereas the three line
has both a service line and a secondary line.
Control (service) line circuit (Fig. 12.6) On making a brake application, a pressure signal from
the tractor control (service) line actuates the relay
Multi-relay valve (Fig. 12.25(a±d)) The purpose
of the multi-relay valve is to enable each of the
two service line circuits to operate independently
Fig. 12.6
Trailer two line brake system
portion of the emergency relay valve to deliver air
pressure from the trailer reservoir to each of the
single diaphragm actuator chambers. In order to
provide the appropriate braking power according
to the trailer payload, a variable load sensing valve
is installed in the control line ahead of the emergency relay valve. This valve modifies the control
line signal pressure so that the emergency relay
valve only supplies the brake actuators with sufficient air pressure to retard the vehicle but not to
lock the wheels. A quick-release valve may be
included in the brake actuator feed line to speed
up the emptying of the actuator chambers to
release the brakes but usually the emergency relay
valve exhaust valve provides this function adequately. If the supply (emergency) line pressure
Fig. 12.7 (a and b)
drops below a predetermined value, then the emergency portion of the emergency relay valve automatically passes air from the trailer reservoir to the
brake actuators to stop the vehicle.
12.3 Air operated power brake equipment
12.3.1 Air dryer (Bendix) (Fig. 12.7(a and b))
Generally, atmospheric air contains water vapour
which will precipitate if the temperature falls low
enough. The amount of water vapour content of
the air is measured in terms of relative humidity.
A relative humidity of 100% implies that the air is
saturated so that there will be a tendency for the
air to condensate. The air temperature and pressure
Air dryer (Bendix)
determines the proportion of water vapour retained
in the air and the amount which condenses.
If the saturation of air at atmospheric pressure
occurs when the relative humidity is 100% and the
output air pressure from the compressor is 8 bar,
that is eight times atmospheric pressure (a typical
working pressure), then the compressed air will
have a much lower saturation relative humidity
ˆ 12:5%.
equal to
Comparing this 12.5% saturation relative
humidity, when the air has been compressed, to
the normal midday humidity, which can range
from 60% in the summer to over 90% in the winter,
it can be seen that the air will be in a state of
permanent saturation.
However, the increase in air temperature which
will take place when the air pressure rises will raise
the relative humidity somewhat before the air actually becomes saturated, but not sufficiently to
counteract the lowering of the saturation relative
humidity when air is compressed.
The compressed air output from the compressor
will always be saturated with water vapour. A safeguard against water condensate damaging the air
brake equipment is obtained by installing an air
dryer between the compressor and the first reservoir.
The air dryer unit cools, filters and dries all the air
supplied to the braking system. The drying process
takes place inside a desiccant cartridge which consists
of many thousands of small microcrystalline pellets.
The water vapour is collected in the pores of these
pellets. This process is known as absorption. There is
no chemical change as the pellets absorb and release
water so that, provided that the pores do not become
clogged with oil or other foreign matter, the pellets
have an unlimited life. The total surface area of the
pellets is about 464 000 m2. This is because each pellet
has many minute pores which considerably increase
the total surface area of these pellets.
Dry, clean air is advantageous because:
Charge cycle (Fig. 12.7(a)) Air from the compressor is pumped to the air dryer inlet port where it
flows downwards between the dryer body and the
cartridge wall containing the desiccant. This cools
the widely but thinly spread air, causing it to condense onto the steel walls and drip to the bottom of
the dryer as a mixture of water and oil (lubricating
oil from the compressor cylinder walls). Any carbon and foreign matter will also settle out in this
phase. The cooled air charge now changes its direction and rises, passing through the oil filter and
leaving behind most of the water droplets and oil
which were still suspended in the air. Any carbon
and dirt which has remained with the air is now
separated by the filter.
The air will now pass through the desiccant so
that any water vapour present in the air is progressively absorbed into the microcrystalline pellet
matrix. The dried air then flows up through both
the check valve and purge vent into the purge air
chamber. The dryness of the air at this stage will
permit the air to be cooled at least 17 C before
any more condensation is produced. Finally the
air now filling the purge chamber passes out to
the check valve and outlet port on its way to the
brake system's reservoirs.
Regeneration cycle (Fig. 12.7(b)) Eventually the
accumulated moisture will saturate the desiccant,
rendering it useless unless the microcrystalline
pellets are dried. Therefore, to enable the pellets
to be continuously regenerated, a reverse flow of
dry air from the purge air chamber is made to occur
periodically by the cut-out and cut-in pressure cycle
provided by the governor action.
When the reservoir air pressure reaches the maximum cut-out pressure, the governor inlet valve
opens, allowing pressurized air to be transferred
to the unloader plunger in the compressor cylinder
head. At the same time, this pressure signal is
transmitted to the purge valve relay piston which
immediately opens the purge valve. The accumulated condensation and dirt in the base of the dryer
is then rapidly expelled due to the existing air pressure in the lower part of the dryer. The sudden drop
in air pressure in the desiccant cartridge chamber
allows the upper purge chamber to discharge dry
air back through the purge vent into the desiccant
cartridge, downwards through the oil filter, finally
escaping through the open purge valve into the
During the reverse air flow process, the expanding dry air moves through the desiccant and effectively absorbs the moisture from the crystals on its
1 the absence of moisture prevents any lubricant in
the air valves and actuators from being washed
2 the absence of moisture reduces the risk of the
brake system freezing,
3 the absence of oil vapour in the airstream caused
by the compressor's pumping action extends the
life of components such as rubber diaphragms,
hoses and `O' rings,
4 the absence of water and oil vapour prevents
sludge forming and material accumulating in
the pipe line and restricting the air flow.
way out into the atmosphere. Once the dryer has
been purged of condensation and moisture, the
purge valve will remain open until the cylinder
head unloader air circuit is permitted to exhaust
and the compressor begins to recharge the reservoir. At this point the trapped air above the purge
relay piston also exhausts, allowing the purge valve
to close. Thus with the continuous rise and fall of
air pressure the charge and regeneration cycles will
be similarly repeated.
A 60 W electric heater is installed in the base of
the dryer to prevent the condensation freezing during cold weather.
from 150 L/min to 500 L/min for a small to large
size compressor. This corresponds to a power loss
of something like 1.5 kW to 6 kW respectively.
Compressor operation When the crankshaft rotates, the piston is displaced up and down causing
air to be drawn through the inlet port into the
cylinder on the down stroke and the same air to
be pushed out on the upward stroke through the
delivery port. The unidirectional flow of the air
supply is provided by the inlet and delivery valves.
The suction and delivery action of the compressor
may be controlled by either spring loaded disc valves
(Fig. 12.9) or leaf spring (reed) valves (Fig. 12.8).
For high speed compressors the reed type valve
arrangements tend to be more efficient.
On the downward piston stroke the delivery
valve leaf flattens and closes, thus preventing the
discharged air flow reversing back into the cylinder
(Fig. 12.8). At the same time the inlet valve is
drawn away from its seat so that fresh air flows
through the valve passage in its endeavour to fill
the expanding cylinder space.
On the upward piston stroke the inlet valve leaf
is pushed up against the inlet passage exit closing
the valve. Consequently the trapped pressurized air
is forced to open the delivery valve so that the air
charge is expelled through the delivery port to the
The sequence of events is continuous with a corresponding increase in the quantity of air delivered
and the pressure generated.
The working pressure range of a compressor
may be regulated by either an air delivery line
mounted unloader valve (Figs 12.10 and 12.11) or
an integral compressor unloader mechanism controlled by an external governor valve (Fig. 12.9). A
further feature which is offered for some applications is a multiplate clutch drive which reduces
pumping and frictional losses when the compressor
is running light (Fig. 12.8).
12.3.2 Reciprocating air compressors
The source of air pressure energy for an air brake
system is provided by a reciprocating compressor
driven by the engine by either belt, gear or shaftdrive at half engine speed. The compressor is usually
base- or flange-mounted to the engine.
To prevent an excessively high air working temperature, the cast iron cylinder barrel is normally
air cooled via the enlarged external surface area
provided by the integrally cast fins surrounding
the upper region of the cylinder barrel. For low to
moderate duty, the cylinder head may also be air
cooled, but for moderate to heavy-duty high speed
applications, liquid coolant is circulated through
the internal passages cast in the aluminium alloy
cylinder head. The heat absorbed by the coolant is
then dissipated via a hose to the engine's own cooling system. The air delivery temperature should not
exceed 220 C.
Lubrication of the crankshaft plain main and bigend bearings is through drillings in the crankshaft,
the pressurized oil supply being provided by the
engine's lubrication system, whereas the piston and
rings and other internal surfaces are lubricated by
splash and oil mist. Surplus oil is permitted to drain
via the compressor's crankcase back to the engine's
sump. The total cylinder swept volume capacity
needed for an air brake system with possibly auxiliary equipment for light, medium and heavy commercial vehicles ranges from about 150 cm3 to
500 cm3, which is provided by either single or twin
cylinder reciprocating compressor. The maximum
crankshaft speed of these compressors is anything
from 1500 to 3000 rev/min depending upon maximum delivery air pressure and application. The
maximum air pressure a compressor can discharge
continuously varies from 7 to 11 bar. A more typical
maximum pressure value would be 9 bar.
The quantity of air which can be delivered at
maximum speed by these compressors ranges
Clutch operation (Fig. 12.8) With the combined
clutch drive compressor unit, the compressor's
crankshaft can be disconnected from the engine
drive once the primary reservoir has reached its
maximum working pressure and the compressor is
running light to reduce the wear of the rotary bearings and reciprocating piston and rings and to
eliminate the power consumed in driving the compressor.
The clutch operates by compressed air and is
automatically controlled by a governor valve similar to that shown in Fig. 12.9.
Fig. 12.8
Single cylinder air compressor with clutch drive
The multiplate clutch consists of four internally
splined sintered bronze drive plates sandwiched
between a pressure plate and four externally
splined steel driven plates (Fig. 12.8). The driven
plates fit over the enlarged end of the splined input
shaft, whereas the driven plates are located inside
the internally splined clutch outer hub thrust plate.
The friction plate pack is clamped together by
twelve circumferentially evenly spaced compression springs which react between the pressure
plate and the outer hub thrust plate. Situated
between the air release piston and the outer hub
thrust plate are a pair of friction thrust washers
which slip when the clutch is initially disengaged.
When the compressor air delivery has charged
the primary reservoir to its preset maximum,
the governor valve sends a pressure signal to the
clutch air release piston chamber. Immediately the
friction thrust washers push the clutch outer hub
thrust plate outwards, causing the springs to
become compressed so that the clamping pressure
between the drive and driven plates is relaxed.
As a result, the grip between the plates is removed.
This then enables the crankshaft, pressure plate,
outer hub thrust plate and the driven plates to
rapidly come to a standstill.
As the air is consumed and exhausted by brake or
air equipment application, the primary reservoir pressure drops to its lower limit. At this point the governor exhausts the air from the clutch release piston
chamber and consequently the pressure springs are
free to expand, enabling the drive and driven plates
once again to be squeezed together. By these means
the engagement and disengagement of the compressor's crankshaft drive is automatically achieved.
the reservoir, builds up pressure and then passes
to the braking system (Fig. 12.9(a)). A small sample
of air from the reservoir is also piped to the underside of the governor piston via the governor inlet
When the pressure in the reservoir is low, the
piston will be in its lowest position so that there is
a gap between the plunger's annular end face and
the exhaust disc valve. Thus air above the unloader
plunger situated in the compressor's cylinder head
is able to escape into the atmosphere via the governor plunger tube central passage.
Compressor unloaded (Fig. 12.9(b)) As the reservoir pressure rises the control spring is compressed
lifting the governor piston until the exhaust disc
valve contacts the plunger tube, thereby closing the
exhaust valve. A further air pressure increase from
the reservoir will lift the piston seat clear of the inlet
disc valve. Air from the reservoir now flows around
the inlet disc valve and plunger tube. It then passes
though passages to the unloader plunger upper
chamber. This forces the unloader plunger down,
thus permanently opening the inlet disc valve situated in the compressor's cylinder head (Fig.
12.9(b)). Under these conditions the compressor
will draw in and discharge air from the cylinder
head inlet port, thereby preventing the compressor pumping and charging the reservoir any
further. At the same time, air pressure acts on
the annular passage area around the governor
plunger stem. This increases the force pushing
the piston upwards with the result that the inlet
disc valve opens fully. When the brakes are used,
the reservoir pressure falls and, when this pressure
reduction reaches 1 bar, the control spring downward force will be sufficient to push down the
governor piston and to close the inlet disc valve
Instantly the reduced effective area acting on
the underside of the piston allows the control
spring to move the piston down even further
until the control exhaust valve (tube/disc) opens.
Compressed air above the unloader plunger will
flow back to the governor unit, enter the open
governor plunger tube and exhaust into the atmosphere. The unloader plunger return spring now
lifts the plunger clear of the cylinder head inlet
disc, permitting the compressor to commence
charging the reservoir.
The compressor will continue to charge the system until the cut-out pressure is reached and once
again the cycle will be repeated.
12.3.3 Compressor mounted unloader with
separate governor (Fig. 12.9(a and b))
Purpose The governor valve unit and the unloader
plunger mechanism control the compressed air output which is transferred to the reservoir by causing
the compressor pumping action to `cut-out' when
the predetermined maximum working pressure is
attained. Conversely, as the stored air is consumed,
the reduction in pressure is sensed by the governor
which automatically causes the compressor to `cutin', thus restarting the delivery of compressed air to
the reservoir and braking system again.
Compressor charging (Fig. 12.9(a)) During the
charging phase, air from the compressor enters
Fig. 12.9
Compressor mounted unloader with separate governor
12.3.4 Unloader valve (diaphragm type)
(Fig. 12.10(a and b))
plunger will also be exposed to the air pressure, so
that the additional force produced fully opens the
inlet valve. Air now passes through the centre of
the plunger and is directed via a passage to the head
of the relay piston.
Eventually a predetermined maximum cut-out
pressure is reached, at which point the air pressure
acting on the relay piston crown overcomes the
relay return spring, causing the relay exhaust
valve to open, expelling the compressed air into
the atmosphere. This enables the compressor to
operate under no-load conditions while the reservoir and braking system is sufficiently charged.
Compressor charging (Fig. 12.10(a)) When air is
initially pumped from the compressor to the reservoir, the unloader valve unit non-return valve
opens and air passes from the inlet to the outlet
port. At the same time, air flows between the neck
of the exhaust valve and the shoulder of the relay
valve piston, but since they both have the same
cross-sectional area, the force in each direction is
equalized. Therefore, the relay piston return spring
is able to keep the exhaust valve closed. Air will
also move through a passage on the reservoir side
of the non-return valve to the chamber on the
plunger side of the diaphragm.
Compressor commences charging (Fig. 12.10
(a and b)) As the stored air is consumed during
a braking cycle, the pressure falls until the cut-in
point (minimum safe working pressure) is reached.
At this point the control spring force equals and
exceeds the opposing air pressure force acting on
the diaphragm on the plunger side. The diaphragm
and plunger will therefore tend to move away from
the control spring until the plunger stem closes the
inlet valve. Further plunger movement pushes the
exhaust valve open so that trapped air in the relay
Compressor unloaded (Fig. 12.10(b)) As the reservoir pressure rises, the diaphragm will move
against the control spring until the governor plunger has shifted sufficiently for the exhaust valve to
close (Fig. 12.10(b)). Further pressure build-up
moves the diaphragm against the control spring
so that the end of the plunger enters its bore and
opens the inlet valve. The annular end face of the
Fig. 12.10 (a and b)
Unloader valve (diaphragm type)
piston crown chamber is able to escape to the
atmosphere. The relay piston return spring closes
the relay exhaust valve instantly so that compression of air again commences, permitting the reservoir to recharge to the pressure cut-out setting.
the pilot piston reaches a maximum (cut-out setting), the pilot piston pushes away from its inlet
seat. A larger piston area is immediately exposed to
the air pressure, causing the pilot piston to rapidly
move over to its outlet seat, thereby sealing the
upper relay piston chamber atmospheric vent. Air
will now flow along the space made between the
pilot piston and its sleeve to act on the upper face of
the relay piston. Consequently, the air pressure on
both sides of the relay piston will be equalized
momentarily. Air pressure acting down on the
exhaust valve overcomes the relay piston return
spring force and opens the compressor's discharge
to the atmosphere. The exhaust valve will then be
held fully open by the air pressure acting on the
upper face of the relay piston. Compressed air from
the compressor will be pumped directly to the
atmosphere and so the higher pressure on the reservoir side of the non-return valve forces it to close,
thereby preventing the stored air in the reservoir
12.3.5 Unloader valve (piston type)
(Fig. 12.11(a and b))
Purpose The unloader valve enables the compressor to operate under no-load conditions, once the
reservoir is fully charged, by automatically discharging the compressor's output into the atmosphere,
and to reconnect the compressor output to the
reservoir once the air pressure in the system drops
to some minimum safe working value.
Compressor charging (Fig. 12.11(a)) When the
compressor starts to charge, air will flow to the
reservoir by way of the horizontal passage between
the inlet and outlet ports.
The chamber above the relay piston is vented to
the atmosphere via the open outlet pilot valve so
that the return spring below the relay piston is able
to keep the exhaust valve closed, thus permitting
the reservoir to become charged.
Compressorcommencescharging (Fig. 12.11(a and b))
As the air pressure in the reservoir is discharged
and lost to the atmosphere during brake applications the reservoir pressure drops. When the pressure has been reduced by approximately one bar
below the cut-out setting (maximum pressure), the
control spring overcomes the air pressure acting on
the right hand face of the pilot piston, making it
shift towards its inlet seat. The pilot piston outlet
Compressor unloaded (Fig. 12.11(b)) As the reservoir pressure acting on the right hand end face of
Fig. 12.11 (a and b)
Unloader valve (piston type)
valve opens, causing the air pressure above the
relay piston to escape to the atmosphere which
allows the relay piston return spring to close
the exhaust valve. The discharged air from the
compressor will now be redirected to recharge
the reservoir.
The difference between the cut-out and cut-in
pressures is roughly one bar and it is not adjustable, but the maximum (cut-out) pressure can be
varied over a wide pressure range by altering the
adjustment screw setting.
it is able to unseat the non-return disc valve against
the closing force of the setting spring. Air will now
pass between the valve disc and its seat before it
enters the delivery port passage on its way to the
reservoir. A larger area of the disc valve is now
exposed to air pressure which forces the disc valve
and piston to move further back against the
already compressed setting spring. As the charging
pressure in the reservoir increases, the air thrust on
the disc and piston face also rises until it eventually
pushes back the valve to its fully open position.
When the air pressure in the reservoir reaches its
predetermined maximum, the governor or unloader
valve cuts out the compressor. The light return
spring around the valve stem, together with air pressure surrounding the disc, now closes the non-return
valve, thereby preventing air escaping back through
the valve. Under these conditions, the trapped air
pressure keeps the disc valve on its seat and holds the
setting spring and piston in the loaded position,
away from the neck of the valve stem. As air is
consumed from the reservoir, its pressure drops so
that the compressor is signalled to cut in again
(restarting pumping). The pressure on the compressor side of the non-return valve then builds up and
opens the valve, enabling the reservoir to recharge.
12.3.6 Single- and multi-circuit protection valve
(Fig. 12.12a)
Purpose Circuit protection valves are incorporated in the brake charging system to provide an
independent method of charging a number of reservoirs to their operating minimum. Where there is a
failure in one of the reservoir circuits, causing loss
of air, they will isolate the affected circuit so that
the remaining circuits continue to function.
Single element protection valve (Fig. 12.12(a))
When the compressor is charging, air pressure is
delivered to the supply port where it increases until
Fig. 12.12 Quadruple circuit protection valve
Should the air pressure in one of the reservoir
systems drop roughly 2.1 bar or more, the setting
spring stiffness overcomes the air pressure acting
on the piston so that it moves against the disc valve
to close the inlet passage. The existing air pressure
stored in the reservoir will still impose a thrust
against the piston, but because the valve face area
exposed to the charge pressure is reduced by the
annular seat area and is therefore much smaller, a
pressure increase of up to 1.75 bar may be required
to re-open the valve.
A total loss of air from one reservoir will automatically cause the setting spring of the respective
protection valve to close the piston against the nonreturn valve.
The protection valves open and close according to
the governor or unloader valve cutting in or cutting
out the pumping operation of the compressor.
Internal passages within the multi-element valve
body, protected by two non-return valves, connect
the delivery from the first and second valve elements to the inlet of the third and fourth valve
elements, which control the delivery to the secondary/park and the trailer reservoir supplies respectively. Delivery to the third and fourth valve
elements is fed from the reservoir connected to the
first and second valve element through passages
within the body.
The additional check valves located in the body of
the multi-protection valve act as a safeguard against
cross-leakage between the front and rear service
reservoirs. Failure of the front reservoir or circuit
still permits the rear service reservoir to supply the
third and fourth element valve. Alternatively, if the
rear service reservoir should fail, the front service
reservoir can cope adequately with delivering air
charge to the third and fourth reservoir.
Multi-element protection valve (Fig. 12.12) Multielement protection valves are available in triple and
quadruple element form. Each element contains
the cap, piston, setting spring and non-return
valve, similar to the single element protection
Charging air from the compressor enters the
supply port of the multi-element protection valve,
increasing the pressure on the inlet face of the first
and second valve element and controlling the delivery to the front and rear service reservoirs respectively. When the predetermined setting pressure is
reached, both element non-return valves open, permitting air to pass through the valve to charge both
service reservoirs.
12.3.7 Pressure reducing valve (piston type)
(Fig. 12.13(a, b and c))
Various parts of an air brake system may need to
operate at lower pressures than the output pressure
delivered to the reservoirs. It is therefore the function of the pressure reducing valve to decrease,
adjust and maintain the air line pressure within
some predetermined tolerance.
Fig. 12.13 (a±c) Pressure reducing valve (piston type)
Operation When the vehicle is about to start a
journey, the compressor charges the reservoirs and
air will flow through the system to the various components. Initially, air flows through to the pressure
reducing valve supply port through the open inlet
valve and out to the delivery port (Fig. 12.13(a)). As
the air line pressure approaches its designed working
value, the air pressure underneath the piston overcomes the stiffness of the control spring and lifts the
piston sufficiently to close the inlet valve and cut off
the supply of air passing to the brake circuit it
supplies (Fig. 12.13(b)).
If the pressure in the delivery line exceeds the
predetermined pressure setting of the valve spring,
the extra pressure will lift the piston still further
until the hollow exhaust stem tip is lifted clear of its
seat. The surplus of air will now escape through the
central exhaust valve stem into the hollow piston
chamber where it passes out into the atmosphere
via the vertical slot on the inside of the adjustable
pressure cap (Fig. 12.13(c)). Delivery line air will
continue to exhaust until it can no longer support
the control spring. At this point, the spring pushes
the piston down and closes the exhaust valve. After
a few brake applications, the delivery line pressure
will drop so that the control spring is able to
expand further, thereby unseating the inlet valve.
Hence the system is able to be recharged.
restrict the air flow in the forward direction, but
to prevent any air movement in the reverse or
opposite direction.
Operation (Fig. 12.14(a)) When compressed air
is delivered to a part of the braking system via
the non-return valve, the air pressure forces the
spherical valve (sometimes disc) head of its seat
against the resistance of the return spring. Air is
then permitted to flow almost unrestricted through
the valve. Should the air flow in the forward direction cease or even reverse, the return spring quickly
closes to prevent air movement in the opposite
direction occurring.
12.3.9 Safety valve (Fig. 12.14(b))
Purpose To protect the charging circuit of an air
braking system from excessive air pressure, safety
valves are incorporated and mounted at various
positions in the system, such as on the compressor
cylinder head, on the charging reservoir or in the
pipe line between the compressor and reservoir.
Operation (Fig. 12.14(b)) If an abnormal pressure
surge occurs in the charging system, the rise in air
pressure will be sufficient to push the ball valve
back against the regulating spring. The unseated
ball now permits the excess air pressure to escape
into the atmosphere. Air will exhaust to the atmosphere until the pressure in the charging system has
been reduced to the blow-off setting determined by
the initial spring adjustment. The regulating spring
then forces the ball valve to re-seat so that no more
air is lost from the charging system.
12.3.8 Non-return (check) valve (Fig. 12.14(a))
Purpose A non-return valve, sometimes known as
a check valve, is situated in an air line system where
it is necessary for the air to flow in one direction
only. It is the valve's function therefore not to
Fig. 12.14 (a and b)
Non-return and safety valves
12.3.10 Dual concentric foot control valve
(Fig. 12.15(a and b))
foot pedal, the piston simultaneously unseats the
inlet/exhaust valves and compressed air from the
reservoirs passes through the upper and lower
valves to the front and rear brake actuators respectively (or to the tractor and trailer brake actuators
Purpose The foot control valve regulates the air
pressure passing to the brake system from the reservoir according to the amount the foot treadle is
depressed. It also imparts a proportional reaction
to the movement of the treadle so that the driver
experiences a degree of brake application.
Balancing (Fig. 12.15(a and b)) With the compressed air passing to the brake actuator chambers,
air pressure is built up beneath the upper and lower
pistons. Eventually the upthrust created by this air
pressure equals the downward spring force; the
pistons and valve carrier lift and the inlet valves
close, thus interrupting the compressed air supply
to the brake actuators. At the same time, the
exhaust valves remain closed. The valves are then
in a balanced condition with equal force above and
Applying brakes (Fig. 12.15(a)) Depressing the
foot treadle applies a force through the graduating
springs to the pistons, causing the exhaust hollow
stem seats for both pistons to close the inlet/
exhaust valves. With further depression of the
Fig. 12.15 (a and b)
Dual concentric foot valve
below the upper piston and with equal air pressure
being held in both halves of the brake line circuits.
Pushing the treadle down still further applies an
additional force on top of the graduating spring.
There will be a corresponding increase in the air
pressure delivered and a new point of balance will
be reached.
Removing some of the effort on the foot treadle
reduces the force on top of the graduating spring.
The pistons and valve carrier will then lift due to
the air pressure and piston return springs. When
this occurs the inlet valves remain closed and the
exhaust valves open to exhausting air pressure
from the brake actuators until a state of balance
is obtained at lower pressure.
assembly both exhaust valves open. Air from both
brake circuits will therefore quickly escape to the
atmosphere thus fully releasing the brakes.
12.3.11 Dual delta series foot control valve
(Fig. 12.16)
Purpose The delta series of dual foot valves provide the braking system with two entirely separate
foot controlled air valve circuits but which operate
simultaneously with each other. Thus, if one half of
the dual foot valve unit should develop a fault then
the balance beam movement will automatically
ensure that the other half of the twin valve unit
continues to function.
Releasing brakes (Fig. 12.15(b)) Removing the
driver's force from the treadle allows the upper
and lower piston and the valve carrier to rise to
the highest position. This initially causes the inlet/
exhaust valves to close their inlet seats, but with
further upward movement of the pistons and valve
Brakes released (12.16(a)) When the brakes are
released, the return springs push up the piston,
graduating spring and plunger assemblies for each
half valve unit. Consequently the inlet disc valves
close and the control tube shaped exhaust valves
Exhaust spring
(a) Brake released
Fig. 12.16 (a and b)
(b) Brake applied
Dual delta foot control valve
open. This permits air to exhaust through the
centre of the piston tube, upper piston chamber
and out to the atmosphere.
The amount the inlet valve opens will be proportional to the graduating spring load, and the
pressure reaching the brake actuator will likewise
depend upon the effective opening area of the
inlet valve. Immediately the braking effort to the
foot treadle is charged, a new state of valve lap will
exist so that the braking power caused by the air
operating on the wheel brake actuator will be progressive and can be sensed by the driver by the
amount of force being applied to the treadle.
When the driver reduces the foot treadle load, the
inlet valve closes and to some extent the exhaust
valve will open, permitting some air to escape
from the actuator to the atmosphere via central
tube passages in the dual piston tubes. Thus the
graduating spring driver-controlled downthrust
and the reaction piston air-controlled upthrust
will create a new state of valve lap and a corresponding charge to the braking power.
Brakes applied (12.16(b)) When the foot treadle
is depressed, a force is applied centrally to the
balance beam which then shares the load between
both plunger spring and piston assemblies. The
downward plunger load initially pushes the piston
tubular stem on its seat, closing the exhaust disc
valve, and with further downward movement
unseats and opens the inlet disc valve. Air from
the reservoirs will now enter the lower piston
chambers on its way to the brake actuators via
the delivery ports.
As the air pressure builds up in the lower piston
chambers it will oppose and compress the graduating springs until it eventually closes the inlet valve.
The valve assembly is then in a lapped or balanced
position where both exhaust and inlet valves are
closed. Only when the driver applies an additional
effort to the treadle will the inlet valve again open
to allow a corresponding increase in pressure to
pass through to the brake actuator.
Fig. 12.17 (a and b)
12.3.12 Hand control valve (Fig. 12.17(a and b))
Purpose These valves are used to regulate the
secondary brake system on both the towing tractor
Hand control valve
and on the trailer. Usually only the tractor front
axle has secondary braking to reduce the risk of
a jack-knife during heavy emergency braking.
ward load of the reaction spring to fully raise the
piston. As a result, the inlet valve closes and the
exhaust valve is unseated, so that the air pressure in
the brake actuator chambers collapses as the air is
permitted to escape to the atmosphere.
Applying brakes (Fig. 12.17(a)) Swivelling the
handle from the released position enables the cam
follower to slide over the matching inclined cam
profile, thereby forcing the cam plate downwards
against the graduating (reaction) spring. The stiffening of the reaction spring forces the piston to
move downwards until the exhaust valve passage
is closed. Further downward movement of the piston unseats the inlet valve, permitting compressed
air from the reservoir to flow through the valve
underneath the piston and out of the delivery
port, to the front brake actuator and to the trailer
brake actuator via the secondary line (blue) coupling to operate the brakes.
12.3.13 Spring brake hand control valve
(Fig. 12.18(a, b and c))
Balancing (Fig. 12.17(a and b)) The air supply
passing through the valve gradually builds up an
opposing upthrust on the underside of the piston
until it eventually overcomes the downward force
caused by the compressed reaction spring. Subsequently the piston lifts, causing the inlet valve to
close so that the compressed air supply to the brake
actuators is interrupted. The exhaust valve during
this phase still remains seated, thereby preventing
air exhaustion. With both inlet and exhaust valves
closed, the system is in a balanced condition, thus
the downward thrust of the spring is equal to the
upthrust of the air supply and the predetermined
air pressure established in the brake actuators.
Rotating the handle so that the reaction spring is
further compressed, opens the inlet valve and
admits more air at higher pressure, producing
a new point of balance.
Partially rotating the handle back to the released
position reduces some of the reaction spring downward thrust so that the existing air pressure is able
to raise the piston slightly. The raised piston results
in the inlet valve remaining seated, but the exhaust
valve opens, permitting a portion of the trapped air
inside the brake actuator to escape into the atmosphere. Therefore the pressure underneath the piston will decrease until the piston upthrust caused
by the air pressure has decreased to the spring
downthrust acting above the piston. Thus a new
state of balance again is reached.
Purpose This hand control valve unit has two
valve assemblies which, due to the cam profile
design, is able to simultaneously deliver an
`upright' and an `inverse' pressure. The valve unit
is designed to provide pressure signals via the delivery
of small volumes of air to the tractor spring brakes
and the trailer's conventional diaphragm actuators. The required full volume of air is then able
to pass from the secondary/park reservoir to the
brake actuators via the relay valves to apply or
release the brakes.
Spring brake release (Fig. 12.18(a)) When the vehicle is in motion with the brakes released, the upright
valve assembly inlet valve is closed and the exhaust
valve is unseated, permitting all the air in the trailer
brake actuators to be expelled. Conversely the
inverse valve assembly delivers a signal pressure
to the spring brake relay valve. This results in
the line from the secondary/park reservoir to
the tractor spring brake actuators to be open.
Thus a large volume of air will be delivered to
the air chambers controlling the compression of
the power springs and the releasing of the tractor
Secondary brake application (Fig. 12.18(b)) As the
handle is moved across the gate to make a secondary
brake application it rotates the cam, depressing the
upright plunger. The exhaust valve closes and the
inlet valve is unseated, causing compressed air to
pass to the trailer brake actuator chambers. As the
pressure in the brake actuators increases, the air
pressure acting on top of the upright piston causes
it to move down against the upthrust exerted by the
graduating spring, closing the inlet valve. This procedure is repeated for further handle movement
until the full secondary brake position is reached
when the air pressure delivered to the trailer brake
chamber is at a maximum.
During this operation the inverse valve assembly, which was delivering maximum pressure when
the handle was in the `off' position, is exhausting
Releasing brakes (Fig. 12.17(b)) Returning the
handle to the released position reduces the down530
Fig. 12.18 (a±c) Spring brake hand control valve
until with the secondary brake position the
delivered pressure is zero.
In other words, the upright valve delivers a gradually increasing pressure to the trailer brake
actuators and, at the same time, the inverse valve
assembly allows the air pressure on the tractor
spring brake actuators to be gradually released.
valve (or hand control valve) reacts on the large
control piston which responds by moving downwards rapidly until the centre stem of the piston
closes the exhaust passage. The downwards movement of the piston pushes open the inlet valve. Air
will now be admitted to the underside of the piston
as it flows through to the service line and brake
actuator. Movement of air from the service reservoir to the service line continues until the combined
upthrust of both piston and valve springs and the
air pressure balances the air signal pressure force,
pushing the piston downwards. The piston now
rises, closing the inlet valve so that both inlet and
exhaust valves are in the lapped condition.
Park brake application (Fig. 12.18(c)) When the
handle is moved from the secondary brake position
to the park position, the cam lifted by the leverage
of the handle about its pivot allows the upright
plunger and the inverse plunger to be raised. The
air pressure in both tractor and trailer brake actuators then exhaust into the atmosphere. The tractor
brakes are now applied in the park position by the
mechanical force exerted by the spring actuators.
Brakes hold (Fig. 12.19(a and b)) A reduction in
signal pressure now produces a greater force, pushing the piston upwards rather than downwards.
The piston rises, closing the inlet valve, followed
by the opening of the exhaust valve. The trapped
air in the service line and actuator will now exhaust
through the hollow valve stem to the atmosphere.
The exhaustion of the service line air continues
until the upward piston force balances the downward force caused by signal pressure. Both inlet
and exhaust valves will subsequently close. These
cycles of events are repeated the instant there is
a change in signal pressure, be it an increasing or
decreasing one, the valve being self-lapping under
all conditions.
12.3.14 Relay valve (piston type) (Bendix)
(Fig. 12.19(a and b))
Purpose The relay valve is used to rapidly operate
a part of a braking system when signalled by either
a foot or hand control valve. This is achieved by
a small bore signal line feeding into the relay valve
which then controls the air delivery to a large bore
output service line. As a result, a small variation in
signal pressure from the foot or hand valve will
produce an instant response by the relay valve to
admit air from the service reservoir directly to the
service line brake system.
Brakes released (Fig. 12.19(b)) When the brakes
are released, the signal pressure collapses, permitting the piston return spring to raise the piston; first
closing the inlet valve, and then opening the
exhaust valve. Air in the service line then escapes
Brakes applied (Fig. 12.19(a)) When the brakes
are applied, a signal pressure from the foot control
Fig. 12.19 (a and b)
Relay valve
through the lower piston chamber and out into the
atmosphere through the hollow valve stem.
has equalized, the diaphragm return spring
upthrust pushes the outer diaphragm rim up onto
its seat whilst the centre of the diaphragm and stem
still seal off the exhaust port. Under these conditions, both inlet and exhaust passages are closed,
preventing any additional air flow to occur to or
from the brake actuators. The diaphragm is therefore in a state of `hold'.
12.3.15 Quick release valve
(Fig. 12.20(a, b and c))
Purpose The quick release valve (QRV) shortens
the brake release time by speeding up the exhaustion of air from the brake actuator chambers, particularly if the actuators are some distance from the
foot, hand or relay valve.
Released position (Fig. 12.20(c)) Releasing the air
pressure above the diaphragm allows the trapped
and pressurized air below the diaphragm to raise
the central region of the diaphragm and stem. The
trapped air in the brake lines and actuator chambers escape into this atmosphere.
Reducing the brake load slightly decreases the
air pressure above the diaphragm, so that some of
the air in the brake lines is allowed to escape before
the pressure on both sides of the diaphragm balances again. The central region of the diaphragm
moves down to close the exhaust port which moves
the diaphragm into its `hold' condition again.
The quick release valve therefore transfers any
increased foot or hand valve control pressure
through it to the brake actuators and quickly
releases the air pressure from the brake actuators
when the brake control valve pressure is reduced.
Applied position (Fig. 12.20(a)) When the brakes
are applied, the air pressure from the foot or hand
control valve enters the upper diaphragm chamber,
forcing the diaphragm and its central stem down
onto the exhaust port seat. The air pressure buildup then deflects downwards the circumferential
diaphragm rim, thereby admitting air to the brake
actuators via the pipe lines.
Hold position (Fig. 12.20(b)) Movement of air
from the inlet port to the outlet ports permits air
to occupy the underside of the diaphragm. Once
the air pressure above and below the diaphragm
Fig. 12.20 (a±c) Quick release valve
By these means the air pressure in the brake actuators will always be similar to the delivery air pressure from the brake control valve.
state. Until a larger service line pressure is applied
to the relay piston, the central stem will not move
further down to open the inlet valve again and
permit more air to pass to the brake actuator chambers. Conversely, if the foot brake is slightly
released, initially the relay piston is permitted to
rise, closing the inlet valve, followed by opening of
the exhaust valve to release some of the air pressure
acting on the brake actuator chambers.
12.3.16 Relay emergency valve
(Fig. 12.21(a±d))
Charging (Fig. 12.21(a)) Air delivery from the
emergency line (red) enters the inlet port and
strainer. The compressed air then opens the check
valve, permitting air to flow across to and around
the emergency piston, whence it passes to the outlet
port leading to the trailer reservoir, enabling it to
become charged.
If the reservoir is completely empty, both the
relay piston and the emergency piston will be in
their uppermost position. Under these conditions,
the exhaust valve will be closed and the inlet valve
open. Therefore some of the air flowing to the
trailer reservoir will be diverted through the inlet
valve to the brake actuator chambers, thereby
operating the brakes. When the trailer reservoir
charge pressure reaches 3.5 bar, air fed through a
hole from the strainer pushes down on the annular
area of the emergency piston causing the inlet to
close. As the reservoir stored pressure rises to 4.2
bar, the downward air pressure force on the emergency piston moves the inlet/exhaust valve stem
away from its exhaust seat, enabling the trapped
air in the brake actuator chambers to escape to the
atmosphere. The brakes will then be released.
Releasing brakes (Fig. 12.21(c)) Removing the
load on the foot control valve first closes off the
air supply to the service line and then releases the
remaining air in the service line to the atmosphere.
The collapse of service line pressure allows the relay
piston to rise due to the existing brake actuator
pressure acting upwards against the relay piston.
The hollow valve stem immediately closes the inlet
valve passage, followed by the relay piston centre
stem exhaust seat lifting clear of the exhaust valve.
Air is now free to escape underneath the relay
piston through the central hollow inlet/exhaust
valve inlet stem and out to the exhaust vent flap
to the atmosphere. The brake actuators now move
to the `off' position, permitting the `S' cam expanders to release the brake shoes from their drums.
Emergency position (Fig. 12.21(d)) If the air pressure in the emergency line (red) should drop below
a predetermined minimum (normally 2 bar), due to
air leakage or trailer breakaway, then the air pressure around the upper shoulder of the emergency
piston will collapse, causing the emergency piston
return spring to rapidly raise the piston. As the
emergency piston rises, the hollow inlet/exhaust
valve stem contacts and closes the relay piston
exhaust stem seat. Further piston lift then opens
the inlet valve. Air from the trailer reservoir is now
admitted through the control inlet valve to the
underside of the relay piston where it then passes
out to the trailer brake actuator chambers. The
trailer brakes are then applied automatically and
independently to the demands of the driver.
A trailer which has been braked to a standstill,
caused by a failure in the emergency line pressure,
can be temporarily moved by opening the trailer's
reservoir drain cock to exhaust the trailer brake
actuators of pressurized air.
Applying brakes (Fig. 12.21(b)) When the brakes
are applied, a signal pressure is passed through the
service line (yellow) to the upper relay piston
chamber, forcing the piston downwards. The lowering of the relay piston and its central exhaust seat
stem first closes the exhaust valve. It then opens the
inlet valve which immediately admits compressed
air from both the emergency line via the check
valve (non-return valve) and the trailer reservoir
through the central inlet valve, underneath the
relay piston and out to the brake actuator chambers. The expanding brake actuator chambers subsequently press the brake shoes into contact with
the drums.
Balancing brakes (Fig. 12.21(b and c)) As the air
pressure in the actuator chambers builds up, the
pressure underneath the relay piston increases its
upthrust on the piston until it eventually equals the
downward relay piston force created by the service
line pressure. At this point the inlet valve also
closes, so that both valves are now in a balanced
12.3.17 Differential protection valve
(Fig. 12.22(a, b and c))
Purpose The differential protection valve prevents both service brakes and secondary brakes
applying their full braking force at any one time.
The valve is designed to supply secondary line
pressure to the spring brake release chambers
when the service brakes are operating or to allow
the service line pressure supplying the service brake
chambers to decrease as the spring brakes are
applied. By these means the spring and diaphragm
actuator forces are prevented from compounding
and overloading the combined spring and diaphragm actuator units and the foundation brakes
which absorb the braking loads.
Fig. 12.21 (a±d)
Brakes in off position (Fig. 12.22(a)) Releasing
both the foot and hand brakes exhausts air from
the service line. Air from the secondary line enters
the secondary inlet port of the valve and flows
between the outer piston and the casing to the
spring brake output ports. It then passes to the
actuator air chambers. The compressed air now
holds the secondary springs in compression,
thereby releasing the brake shoes from the drums.
Relay emergency valve
Secondary (spring) brake application (Fig.
12.22(b)) When the secondary (spring) brakes
are applied, following the initial application and
holding of the service (foot) brakes, the compressed
air in the spring actuator chambers and in the
secondary line is exhausted via the differential protection valve to the atmosphere through the hand
control valve. As the secondary line pressure
reduces, the pressure trapped in the service line
due to the previous foot brake application becomes
greater than the decreasing pressure in the secondary line. It therefore causes the inner piston to be
pushed across to block the secondary port air exit.
Immediately afterwards, the outer piston is
unseated so that service line air now flows through
the valve from the service line inlet port to the
spring delivery ports and from there to the spring
actuator chambers. The service line air which has
entered the secondary line now holds the springs so
that they are not applied whilst the driver is still
applying the foot brake.
As the driver reduces the foot pedal pressure, the
corresponding reduction in service line pressure permits the outer piston, followed by the inner piston,
to move away from the secondary line inlet port,
closing the service line inlet port and opening the
secondary inlet port. The compressed air occupying
the spring brake actuator chambers is now permitted to fully exhaust so that the expanding springs
re-apply the brakes simultaneously as the service
(foot) brakes are being released.
spring brake application, the secondary line will be
exhausted of compressed air, which was essential
for the spring brakes to operate. Therefore, as the
service line pressure rises, it pushes the inner piston
against its seat, closing the secondary line inlet
port. With a further increase in service line pressure, the outer piston becomes unseated so that
service line pressure can now flow through the
valve and pass on to the spring brake actuators.
This withdraws the spring brake force, thereby
preventing the compounding of both spring and
service chamber forces.
While the differential protection valve is in
operation, an approximate 2.1 bar pressure differential between the service pressures and the
delivered effective anti-compounding pressure will
be maintained across the valve.
12.3.18 Double check valve (Fig. 12.23)
Purpose When two sources of charging a pipe line
are incorporated in a braking system such as the
service (foot) line and secondary (hand) line circuits, a double check valve is sometimes utilized
to connect whichever charging system is being
used to supply the single output circuit and to
isolate (disconnect) the charging circuit which is
not being operated at that time.
Operation (Fig. 12.23(a and b)) The two separate
charging circuits (service and secondary lines) are
joined together by the end inlet ports of the double
check valve. When one of the brake systems is
applied, air charge will be delivered to its double
Service (foot) brake application (Fig. 12.22(c))
When the service (foot) brakes are applied after a
Fig. 12.22 (a±c)
Differential protection valve
(c) Front axle – foot brake applied under ABS/TCS conditions
To rear
(d) Front axle – foot brake applied with a fault in the electronic-pneumatics
Fig. 12.41 (c and d)
To rear
(a) Rear axle – foot brake released
Dual circuit
spring brake
actuator (SBA)
2-way valve
Inlet / exhaust
solenoid valve
(in/ex SV)
2/2 solenoid
(2/2 SV)
3/2-way valve
(3/2 WV)
Redundancy valve
To trailer brakes
To front brakes
Axle modulator for
drive axle (AM)
(b) Rear axle – foot brake applied (normal brake operation)
in/ex SV
3/2 WV
2/2 SV
To front brakes
in/ex SV
To trailer brakes
Fig. 12.42 (a and b)
Electronic-pneumatic rear brake system
(c) Rear axle – foot brake applied under ABS/TCS conditions
Right hand
in/ex SV
2/2 SV
brake SBA
3/2 WV
Left hand
To trailer brakes
(d) Rear axle – foot brake applied with a fault in the electronic-pneumatics
in/ex SV
To front brakes
Right hand
in/ex SV
3/2 WV
2/2 SV
To front brakes
in/ex SV
Left hand
To trailer brakes
Fig. 12.42 (c and d)
(a) Trailer axles – foot brake released
head for
hand control
valve (P-HCV)
Proportional valve
head for
EPB trailer control valve
To rear brakes
To front brakes
(b) Trailer axles – foot brake applied (normal brake operation)
To rear brakes
To front brakes
Fig. 12.43 (a and b)
Electronic-pneumatic tractor unit brakes coupled to towed trailer
(c) Trailer axles – parking brake applied
To rear brakes
To front brakes
(d) Trailer axles – foot brake applied with a fault in the electronic-pneumatics
To rear brakes
To front brakes
Fig. 12.43 (c and d)
pressure in the wheel spring brake actuator brake
lines. Consequently the axle modulator (AM)
de-energizes the inlet and exhaust solenoid valves,
causing the inlet valve to close and the exhaust valve
to open, hence brake pressure will be prevented
from reaching the spring brake actuator and any
existing air pressure in the spring brake actuator
(SBA) will be expelled via the exhaust solenoid
valve. Air pressure in the pipe lines between the
brake value sensor, the redundancy valve and axle
modulator unit will also be exhausted by way of the
foot control valve exit and the axle modulator (AM)
exhaust solenoid valve exit.
and the feed-back from one of the rear axle speed
sensors registers excessive wheel lock/spin, then the
electronic control module (ECM) will actuate the
axle modulator (AM). Instantly the unstable
breakaway wheel inlet/exhaust solenoid valves are
energized, that is, the inlet solenoid valve closes
and the corresponding exit valve opens, therefore
blocking the air pressure passage leading to the
wheel brake actuator and causing the air pressure
in the actuator diaphragm chamber to collapse.
The energizing and de-energizing (opening and
closing) of each pair of solenoid valves is repeated
continuously to adjust the magnitude of the wheel
braking as demanded by the driver without the
individual wheel locking when braking or for one
of the wheels to spin due to poor road grip when
the vehicle is accelerating.
During ABS/TCS braking conditions the redundancy (pneumatic circuit) brake system must not
become active; to achieve this, the 2/2 solenoid
valve is energized and closes so that air pressure is
maintained underneath the 3/2-way valve piston.
Accordingly the space above the relay valve
remains open to the atmosphere and the 3/2-way
valve inlet remains shut, hence locking out the
redundancy brake circuit. Braking will revert to
normal electronic-pneumatic control when the difference in wheel speeds is relatively small.
Rear axle Ð foot brake applied (Fig. 12.42(b))
The rear axle reservoir tank delivers maximum
supply pressure to the brake value sensor (BVS),
the redundancy valve (RDV) and to the rear
axle modulator (AM).
When the foot brake pedal is applied the travel
sensor within the brake value sensor unit monitors
the pedal movement and relays this information to
the electronic control module. The brake switches
will also close thereby informing the electronic control module (ECM) to operate the stop lights. The
electronic control module (ECM) responds by signalling the axle modulators (AMs) to energize their
corresponding inlet/exhaust solenoid valves. The
exhaust valve will therefore close, whereas the inlet
valves now open to permit rear reservoir tank supply
pressure to flow via the 2-way valve to the dual
circuit spring brake actuators (SBA) thereby
operating the brakes. In addition, brake pressure is
conveyed to the redundancy valve (RDV) where it
flows though the 2/2 solenoid valve and then
actuates the 3/2-way valve. This closes the 3/2-way
valve, preventing redundancy circuit (pneumatic
control pressure) control pressure reaching the axle
modulator solenoid valves and at the same time
exhausts the air holding down the relay valve's
piston, hence it causes the relay valve to block the
rear axle redundancy valve reservoir tank supply
pressure entering the redundancy brake circuit.
Control of rear axle braking is achieved via the
speed sensors giving feed-back on each wheel retardation or acceleration to the axle modulator and
with the calculated brake pressure needs derived by
the electronic control module (ECM) delivers the
appropriate brake pressure to the wheel spring
brake actuators.
Rear axle Ð foot brake applied with a fault in the
electronic-pneumatics (Fig. 12.42(d)) Should the
electronic-pneumatic brake circuit fail, the axle
modulator (AM) solenoid valve de-energizes so
that the solenoid inlet valves close, whereas the
exit valves open. Consequently compressed air
under the 3/2-way valve piston escapes from the
left hand solenoid exit valve thereby permitting the
3/2-way valve inlet valve to open. Foot control
valve modulated brake pressure now enters the
relay valve's upper piston chamber where it controls the delivery of redundancy circuit pneumatic
pressure to both wheel brake spring actuators via
the 2-way valves which are now positioned to block
compressed air reaching the axle modulator solenoid valves. Note with the redundancy circuit operating there will be no active ABS at the front and
rear axles.
12.5.4 Trailer axle braking (Fig. 12.43(a±d))
Rear axles Ð foot brake applied under ABS/TCS
conditions (Fig 12.42(c)) If the brakes are applied
Trailer axle Ð foot brake released (Fig. 12.43(a))
When the brake pedal is released the foot travel
sensor signals the electronic control module (ECM)
to release the brakes by de-energizing the proportional valve (PV). As a result the proportional valve
(PV) inlet closes and its exit opens so that air
pressure in the throttle valve and relay valve piston
chamber is exhausted. Accordingly the relay valve
inlet closes and its exit opens, thus permitting the
brake pressure leading to the coupling head to
collapse and for the brakes to be released.
lower control piston within the relay valve. Supply
pressure acting beneath the reaction piston will
now be able to lift the reaction piston and inner
valve assembly until the upper control piston plunger closes the exit. Further upward movement then
opens the inlet valve, thus permitting supply air
pressure to flow through the partially open inlet
valve to the `coupling head for brake', and hence to
the trailer attached relay emergency valve where it
modulates the supply pressure reaching the wheel
brake actuators.
Trailer axles Ð foot brake applied (Fig. 12.43(b))
Maximum supply pressure from the rear axle and
trailer reservoirs is routed to both the brake value
sensor (BVS) and to the proportional relay valve
(PRV) respectively.
When the driver operates the foot brake pedal,
the travel sensors located inside the brake value
sensor (BVS) unit measure the pedal downward
movement and feed this information to the central
electronic control module (ECM). At the same time
the brake switch closes thereby instructing the electronic control module to switch on the stop light.
The electronic control module (ECM) responds
by directing a calculated variable control current to
the proportional valve (PV) which forms part of
the EPB-trailer control valve unit. Energizing the
proportional valve's solenoid, closes the exit valve
and opens the inlet valve in proportion to the
amount of braking requested. Controlled air pressure now enters the relay valve piston upper chamber; this closes the exit valve and opens the inlet
valve in proportion to the degree of braking
demanded. Modulated brake pressure will now
pass to the coupling head brake circuit where it is
then relayed to the trailer wheel brake actuators via
the trailer-mounted relay emergency valve.
Trailer axle Ð foot brake applied with a fault in the
electronic-pneumatics (Fig. 12.43(d)) If there is a
fault in the electronic-pneumatic system the proportional valve (PV) is de-energized, causing its
inlet valve to close and its exit to open; pressurized
air is therefore able to exhaust from the relay
valve's upper control piston chamber via the proportional valve exit. Note the relay valve consists of
an upper control piston and a lower assembly with
upper and lower piston regions and which incorporates an internal double seat inlet and exit valve. As
a result the upper control piston moves to its
uppermost position, the inlet valve initially closes
and the trapped supply pressure and the redundancy brake pressure (foot control valve pneumatic
pressure) acting underneath the inner assembly's
upper and lower piston region, pushes up the
assembly against the hand control valve park pressure sufficiently to close the exit valve and to open
the inlet valve. Brake pressure will hence be delivered to the trailer wheel brake actuators via the
`coupling head for brakes'.
Brake line to trailer defective (Fig. 12.43(b)) If the
brake line to the trailer fractures the output pressure from the relay valve drops, causing the pressure above the throttle valve piston to collapse; this
forces the throttle valve piston to rise and partially
close the throttle valve, thereby causing a rapid
reduction in the supply pressure flowing to the
coupling head supply line (diagram not shown
with throttle valve in defective pipe line position).
As a result the relay emergency valve mounted on
the trailer switches into braking mode and hence
overrides the electronic-pneumatic circuit brake
control to bring the vehicle to rest.
Trailer axle Ð parking brake applied (Fig. 12.43(c))
With the `park' hand control valve in the `off '
position the hand control valve central plunger
closes the exit and pushes open the inlet valve.
Compressed air from the parking reservoir tank is
therefore able to flow to the relay valve part of the
EPB trailer control valve via the open inlet valve
inside the hand control valve.
When the `park' control valve lever moves
towards `park' position the central plunger rises,
causing the exit to open and the inlet valve to close.
Air pressure therefore exhausts from above the
13 Vehicle refrigeration
Refrigeration transport is much in demand to
move frozen or chilled food from storage centres
to shops and supermarkets. Thermally insulated
body containers used for frozen and chilled food
deliveries for both small rigid trucks and large
articulated vehicles are shown in Figs 13.1 and 13.2
respectively. Refrigeration systems designed for
motor vehicle trucks are basically made up of two
parts supported on an aluminium alloy or steel
frame. The condenser unit which is mounted outside the thermally insulated cold storage compartment, comprises a diesel engine (and optional
electric motor) compressor, condenser coil, fan
thermostat and accessories. The evaporator unit
protruding inside the cold storage body contains
the evaporator coil, evaporator fan expansion
valve remote feeler bulb and any other accessories.
For certain applications a standby electric motor is
incorporated to drive the compressor when the
truck is being parked at the loading or delivery
site for a long period of time such as overnight,
the electricity supply being provided by the local
premises' mains power. Typical self-contained
refrigeration unit arrangements incorporating an
engine, compressor, evaporator, condenser, fans
and any other accessories for small to medium
and large frozen storage compartments are shown
in Figs 13.3 and 13.4 respectively. Temperature
control is fully automatic on a start±stop cycle.
With small and medium size refrigeration systems
the engine runs at full governed speed until the thermostat temperature setting is reached. It then automatically reduces speed and disconnects the magnetic
or centrifugal clutch which stops the compressor.
A slight increase of temperature will return the engine
to full speed and again driving the compressor.
The cold storage compartment temperature for
frozen food is usually set between 22 C and 25 C
whereas the chilled compartment temperature is set
between ‡3 C and ‡5 C.
13.1 Refrigeration terms (Fig. 13.5)
To understand the operating principles of a refrigeration system it is essential to appreciate the following terms:
Refrigerant This is the working fluid that circulates though a refrigeration system and produces
both cooling and heating as it changes state. The
desirable properties of a refrigerant fluid are such
that it flows through the evaporator in its vapour
state, absorbs heat from its surroundings, then
transfers this heat via the flow of the refrigerant
Air circulation
Overcab mounted
Thermally insulated
box body frozen/
cold storage
Fig. 13.1 Overcab mounted self-contained refrigeration system for small and medium rigid trucks
Cold air distribution ducting
Nose mounted
Thermally insulated
box body frozen/cold
storage compartment
Fig. 13.2
Nose mounted self-contained refrigeration system for large articulated truck
to the condenser; the refrigerant then condenses
to a liquid state and in the process dissipates heat
taken in by the evaporator to the surrounding
atmosphere. Refrigerants are normally in a vapour
state at atmospheric pressure and at room temperature because they boil at temperatures below zero
on the celsius scale; however, under pressure the
refrigerant will convert to a liquid state.
Saturated temperature (Fig. 13.5) This is the temperature at which a liquid converts into vapour or
a vapour converts into liquid, that is, the boiling
point temperature.
Subcooled liquid (Fig. 13.5) This is a liquid at
any temperature below its saturated (boiling)
Saturated vapour (Fig. 13.5) This is the vapour
which is formed above the surface of a liquid
when heated to its boiling point.
Saturated liquid (Fig. 13.5) This is a liquid heated
to its boiling point, that is, it is at the beginning of
Sight glass
Diesel engine
Fig. 13.3
Vee cylinder compressor
Light to medium duty diesel engine and standby electric motor belt driven compressor refrigeration unit
Drier Expansion
Sight glass
and clutch
Vee cylinder compressor
In-line four cylinder
diesel engine
Fig. 13.4 Heavy duty diesel engine shaft driven compressor refrigeration unit
Latent heat of evaporation
Refrigerant rejecting
heat, converting
to liquid
begins to
boil (vaporize)
Refrigerant temperature (°c)
Refrigerant absorbs
heat, converting
to vapour
boiled to
a saturated
Heat increase (J)
Fig. 13.5 Illustrative relationship between the refrigerant's temperature and heat content during a change of state
Latent heat of evaporation (Fig. 13.5) This is
the heat needed to completely convert a liquid to
a vapour and takes place without any temperature rise.
10 bar respectively flows from the receiver to the
expansion valve via the sight glass and drier. The
refrigerant then rapidly expands and reduces its
pressure as it passes out from the valve restriction and in the process converts the liquid into
a vapour flow.
2 The refrigerant now passes into the evaporator
as a mixture of liquid and vapour, its temperature
being lowered to something like 10 C with
a corresponding pressure of 2 bar (under these
conditions the refrigerant will boil in the evaporator). The heat (latent heat of evaporation)
necessary to cause this change of state will
come from the surrounding frozen compartment
in which the evaporator is exposed.
Superheated vapour (Fig. 13.5) This is a vapour
heated to a temperature above the saturated
temperature (boiling point); superheating can
only occur once the liquid has been completely
13.2 Principles of a vapour±compression cycle
refrigeration system (Fig. 13.6)
1 High pressure subcooled liquid refrigerant at
a typical temperature and pressure of 30 C and
40 °C 10 bar
(high pressure)
(high pressure) vapour
60 °C 10 bar
(high pressure)
40 °C 10 bar
Fig. 13.6
30 °C 10 bar
40 °C 10 bar
Refrigeration vapour±compression cycle
–10 °C 2 bar
(low pressure)
(low pressure)
8 °C 2 bar
(low pressure)
– 10 °C 2 bar
(low pressure)
erant's temperature down to something like 30 C
but without changing pressure which still
remains at 10 bar.
3 As the refrigerant moves through the evaporator
coil it absorbs heat and thus cools the space
surrounding the coil. Heat will be extracted
from the cold storage compartment until its
pre-set working temperature is reached, at this
point the compressor switches off. With further
heat loss through the storage container insulation leakage, doors opening and closing and
additional food products being stored, the compressor will automatically be activated to restore
the desired degree of cooling. The refrigerant
entering the evaporator tube completes the
evaporation process as it travels through the
evaporator coil so that the exit refrigerant from
the evaporator will be in a saturated vapour
state but still at the same temperature and
pressure as at entry, that is, 10 C and 2 bar
4 The refrigerant is now drawn towards the
compressor via the suction line and this causes
the heat from the surrounding air to superheat
the refrigerant thus raising its temperature to
something like 8 C; however, there is no change
in the refrigerant's pressure.
5 Once in the compressor the superheated vapour
is rapidly compressed, consequently the superheated vapour discharge from the compressor is
at a higher temperature and pressure in the order
of 60 C and 10 bar respectively.
6 Due to its high temperature at the exit from the
compressor the refrigerant quickly loses heat to
the surrounding air as it moves via the discharge
line towards the condenser. Thus at the entry to
the condenser the refrigerant will be in a saturated vapour state with its temperature now lowered to about 40 C; however, there is no further
change in pressure which is still therefore 10 bar.
7 On its way through the condenser the refrigerant
saturated vapour condenses to a saturated liquid
due to the stored latent heat in the refrigerant
transferring to the surrounding atmosphere via
the condenser coil metal walls. Note the heat
dissipated to the surrounding atmosphere by
the condenser coil is equal to the heat taken in
by the evaporator coil from the cold storage
compartment and the compressor.
8 After passing through the condenser where heat
is given up to the surrounding atmosphere the
saturated liquid refrigerant now flows into the
receiver. Here the increased space permits a
small amount of evaporation to occur, the refrigerant then completes the circuit to the expansion
valve though the liquid line where again heat is
lost to the atmosphere, and this brings the refrig-
13.3 Refrigeration system components
A description and function of the various components incorporated in a refrigeration system will be
explained as follows:
13.3.1 Reciprocating compressor cycle of
operation (Fig. 13.7(a±d))
Circulation of the refrigerant between the evaporator and the condenser is achieved by the pumping
action of the compressor. The compressor draws in
low pressure superheated refrigerant vapour from
the evaporator and discharges it as high pressure
superheated vapour to the condenser. After flowing
through the condenser coil the high pressure refrigerant is now in a saturated liquid state; it then flows
to the expansion valve losing heat on the way and
thus causing the liquid to become subcooled.
Finally the refrigerant expands on its way through
the expansion valve causing it to convert into a
liquid-vapour mix before re-entering the evaporator coil.
The reciprocating compressor completes a suction and discharge cycle every revolution; the outward moving piston from TDC to BDC forms the
suction-stroke whereas the inward moving piston
from BDC to TDC becomes the discharge stroke.
Suction stroke (Fig. 13.7(a and b)) As the crank
shaft rotates past the TDC position the piston commences its suction stroke with the discharge reed
valve closed and the suction reed valve open (Fig.
13.7(a and b)). The downward sweeping piston now
reduces the cylinder pressure from P1 to P2 as its
volume expands from V1 to V2, the vapour refrigerant in the suction line is now induced to enter the
cylinder. The cylinder continues to expand and to be
filled with vapour refrigerant at a constant pressure
P1 to the cylinder's largest volume of V3, that is the
piston's outermost position BDC, see Fig. 13.8.
Discharge stroke (Fig. 13.9(c and d)) As the
crankshaft turns beyond BDC the piston begins its
upward discharge stroke, the suction valve closes and
the discharge valve opens (see Fig. 13.7(c and d)).
The upward moving piston now compresses the
refrigerant vapour thereby increasing the cylinder
pressure from P1 to P2 through a volume reduction
from V3 to V4 at which point the cylinder pressure
Low pressure
vapour refrigerant
from evaporator
High pressure
vapour refrigerant
to condenser
(a) Piston at TDC both valves
(b) Piston on downward
closed high pressure vapour
suction stroke vapour
trapped in discharge line and
refrigerant drawn into
clearance volume
Fig. 13.7 (a±d)
(c) Piston at BDC both valves
closed, cylinder filled with
fresh vapour refrigerant
(d) Piston on upward discharge
stroke, suction valve closed
discharged valve open,
compressed vapour refrigerant
pumped into discharge line
Reciprocating compressor cycle of operation
Vapour discharge
Pressure (bar)
r ex
Vapour intake
V1 (TDC)
Swept volume
Fig. 13.8
Reciprocating compressor pressure-volume cycle
V3 (BDC)
(attached to output
side of evaporator)
to suction
(attached to output
side of evaporator)
(a) Valve closed
Fig. 13.9 (a and b)
to suction
(b) Valve open
Thermostatic expansion valve
equals the discharge line pressure; the final cylinder
volume reduction therefore from V4 back to V1 will
be displaced into the high pressure discharge line at
the constant discharge pressure of P2 (see Fig. 13.8).
accordingly the refrigerant condenses and reverts
to a liquid state. Heat will be rejected from the
refrigerant during this phase change via conduction
though the metal walls of the tubing and convection to the surrounding atmosphere.
A condenser consists of a single tube shaped
so that there are many parallel lengths with semicircular ends which therefore form a continuous
winding or coil. Evenly spaced cooling fins are
normally fixed to the tubing, this greatly increases
the surface area of the tubing exposed to the convection currents of the surrounding atmosphere,
see Fig. 13.15(a and b).
Fans either belt driven or directly driven by an
electric motor are used to increase the amount of
air circulation around the condenser coil, this
therefore improves the heat transfer taking place
between the metal tube walls and fins to the surrounding atmosphere. This process is known as
forced air convection.
13.3.2 Evaporator (Fig. 13.6)
The evaporator's function is to transfer heat from
the food being stored in the cold compartment
into the circulating refrigerant vapour via the
fins and metal walls of the evaporator coil tubing
by convection and conduction respectively. The
refrigerant entering the evaporator is nearly all
liquid but as it moves through the tube coil, it
quickly reaches its saturation temperature and is
converted steadily into vapour. The heat necessary
for this change of state comes via the latent heat
of evaporation from the surrounding cold chamber atmosphere.
The evaporator consists of copper, steel or stainless steel tubing which for convenience is shaped in
an almost zigzag fashion so that there are many
parallel lengths bent round at their ends thus
enabling the refrigerant to flow from side to side.
To increase the heat transfer capacity copper fins
are attached to the tubing so that relatively large
quantities of heat surrounding the evaporator coil
can be absorbed through the metal walls of the
tubing, see Fig. 13.15(a and b).
13.3.4 Thermostatic expansion valve
(Fig. 13.9(a and b))
An expansion valve is basically a small orifice which
throttles the flow of liquid refrigerant being
pumped from the condenser to the evaporator;
the immediate exit from the orifice restriction will
then be in the form of a rapidly expanding refrigerant, that is, the refrigerant coming out from
the orifice is now a low pressure continuous liquidvapour stream. The purpose of the thermostatic
valve is to control the rate at which the refrigerant
passes from the liquid line into the evaporator and
13.3.3 Condenser (Fig. 13.6)
The condenser takes in saturated refrigerant
vapour after it has passed though the evaporator
and compressor, progressively cooling then takes
place as it travels though the condenser coil,
to keep the pressure difference between the high
and low pressure sides of the refrigeration system.
The thermostatic expansion valve consists of a
diaphragm operated valve (see Fig. 13.9(a and b)).
One side of the diaphragm is attached to a spring
loaded tapered/ball valve, whereas the other side of
the diaphragm is exposed to a refrigerant which
also occupies the internal space of the remote feeler
bulb which is itself attached to the suction line tube
walls on the output side of the evaporator. If the
suction line saturated/superheated temperature
decreases, the pressure in the attached remote
feeler bulb and in the outer diaphragm chamber
also decreases. Accordingly the valve control
spring thrust will partially close the taper/ball
valve (see Fig. 13.9(a)). Consequently the reduced
flow of refrigerant will easily now be superheated
as it leaves the output from the evaporator. In
contrast if the superheated temperature rises, the
remote feeler bulb and outer diaphragm chamber
pressure also increases, this therefore will push the
valve further open so that a larger amount of refrigerant flows into the evaporator, see Fig. 13.9(b).
The extra quantity of refrigerant in the evaporator
means that less superheating takes place at the output from the evaporator. This cycle of events is
a continuous process in which the constant superheated temperature control in the suction line
maintains the desired refrigerant supply to the
A simple type of thermostatic expansion valve
assumes the input and output of an evaporator are
both working at the same pressure; however, due to
internal friction losses the output pressure will be
slightly less than the input. Consequently the lower
output pressure means a lower output saturated
temperature so that the refrigerant will tend to
vaporize completely before it reaches the end of
the coil tubing. As a result this portion of tubing
converted completely into vapour and which is in a
state of superheat does not contribute to the heat
extraction from the surrounding cold chamber so
that the effective length of the evaporator coil is
reduced. To overcome early vaporization and
superheating, the diaphragm chamber on the
valve-stem side is subjected to the output side of
the evaporator down stream of the remote feeler
bulb. This extra thrust opposing the remote feeler
bulb pressure acting on the outer diaphragm chamber now requires a higher remote feeler bulb pressure to open the expansion valve.
13.3.5 Suction pressure valve (throttling valve)
(Fig. 13.10(a and b))
This valve is incorporated in the compressor
output suction line to limit the maximum suction
Valve seat
Flat valve
(a) Valve fully open
Outlet vapour
to compressor
suction valves
(b) Valve partially open
Fig. 13.10 Suction pressure regulating valve (throttling valve)
pressure generated by the compressor thereby safeguarding the compressor and drive engine/motor
from overload. If the maximum suction pressure is
exceeded when the refrigeration system is switched
on and started up (pull down) excessive amounts of
vapour or vapour/liquid or liquid refrigerant may
enter the compressor cylinder, which could produce
very high cylinder pressures; this would therefore
cause severe strain and damage to the engine/electric
motor components, conversely if the suction line
pressure limit is set very low the evaporator may
not operate efficiently.
The suction pressure valve consists of a combined piston and bellows controlled valve subjected
to suction vapour pressure.
When the compressor is being driven by the
engine/motor the output refrigerant vapour from
the evaporator passes to the intake port of the
suction pressure valve unit; this exposes the bellows
to the refrigerant vapour pressure and temperature.
Thus as the refrigerant pressure rises the bellows
will contract against the force of the bellows spring;
this restricts the flow of refrigerant to the compressor (see Fig. 13.10(a)). However, as the bellows
temperature rises its internal pressure also increases
and will therefore tend to oppose the contraction of
the bellows. At the same time the piston will be
subjected to the outlet vapour pressure from the
suction pressure valve before entering the compressor cylinders, see Fig. 13.10(b). If this becomes
excessive the piston and valve will move towards
the closure position thus restricting the entry of
refrigerant vapour or vapour/liquid to the compressor cylinders. Hence it can be seen that the
suction pressure valve protects the compressor
and drive against abnormally high suction line
13.3.6 Reverse cycle valve (Fig. 13.11(a and b))
The purpose of this valve is to direct the refrigerant
flow so that the refrigerant system is in either a
cooling or heating cycle mode.
Refrigerant cycle mode (Fig. 13.11(a)) With the
pilot solenoid valve de-energized the suction passage to the slave cylinder of the reverse cycle valve
is cut off whereas the discharge pressure supply
from the compressor is directed to the slave piston. Accordingly the pressure build-up pushes the
piston and both valve stems inwards; the left
hand compressor discharge valve now closes the
To compressor
(a) Cooling cycle
Fig. 13.11 (a and b)
(b) Heating cycle
Reverse cycle valve
To compressor
compressor discharge passage to the evaporator
and opens the compressor discharge passage to
the condenser whereas the right hand double compressor discharge valve closes the condenser to
compressor suction passage and opens the evaporator to the compressor suction pressure.
impurities. Liquid line driers are plumbed in on
the output side of the receiver, this is because the
moisture is concentrated in a relatively small space
when the refrigerant is in a liquid state.
A liquid line drier usually takes the form of
a cylindrical cartridge with threaded end connections so that the drier can be replaced easily when
necessary. Filter material is usually packed in at
both ends; in the example shown Fig. 13.12 there
are layers, a coarse filter, felt pad and a fine filter.
In between the filter media is a desiccant material,
these are generally of the adsorption desiccant kind
such as silica gel (silicon dioxide) or activated
alumna (aluminium oxide). The desiccant substance has microscopic holes for the liquid refrigerant to pass through; however, water is attracted
to the desiccant and therefore is prevented from
moving on whereas the dry (free of water) clean
refrigerant will readily flow through to the expansion valve.
Heat/defrost cycle mode (Fig. 13.11(b)) Energizing the pilot solenoid valve cuts off the compressor
discharge pressure to the slave cylinder of the
reverse cycle valve and opens it to the compressor
suction line. As a result the trapped refrigerant
vapour in the slave cylinder escapes to the compressor suction line thus permitting the slave piston
and both valves to move to their outermost position.
The left hand compressor discharge valve now
closes the compressor discharge to the condenser
passage and opens the compressor discharge to the
evaporator passage whereas the right hand compressor suction double valve closes the evaporator
to the compressor suction passage and opens the
condenser to compressor suction pressure.
13.3.8 Oil separator (Fig. 13.13)
Oil separators are used to collect any oil entering
the refrigeration system through the compressor
and to return it to the compressor crankcase and
sump. The refrigerant may mix with the compressor's lubrication oil in the following way:
13.3.7 Drier (Fig. 13.12)
Refrigerant circulating the refrigerator system
must be dry, that is, the fluid, be it a vapour or a
liquid, should not contain water. Water in the form
of moisture can promote the formation of acid
which can attack the tubing walls and joints and
cause the refrigerant to leak out. It may initiate the
formation of sludge and restrict the circulation of
the refrigerant. Moisture may also cause ice to form
in the thermostatic expansion valve which again
could reduce the flow of refrigerant. To overcome
problems with water contamination driers are normally incorporated in the liquid line; these liquid
line driers not only remove water contained in the
refrigerant, they also remove sludge and other
mixture from
1 During the cycle of suction and discharge refrigerant vapour periodically enters and is displaced
from the cylinders; however, if the refrigerant
flow becomes excessive liquid will pass through
the expansion valve and may eventually enter the
suction line via the evaporator. The fluid may
then drain down the cylinder walls to the crankcase and sump. Refrigerant mixing with oil
dilutes it so that it loses its lubricating properties:
the wear and tear of the various rubbing components in the compressor will therefore increase.
dehydrating material
Fine filter
Felt pad
Coarse filter
Fig. 13.12 Adsorption type liquid line drier
Dry clean
of a cylindrical chamber with a series of evenly
spaced perforated baffle plates or wire mesh partitions attached to the container walls; each baffle
plate has a small segment removed to permit the
flow of refrigerant vapour (Fig. 13.13), the input
from the compressor discharge being at the lowest
point whereas the output is via the extended tube
inside the container. A small bore pipe connects the
base of the oil separator to the compressor crankcase to provide a return passage for the liquid oil
accumulated. Thus when the refrigerator is operating, refrigerant will circulate and therefore passes
though the oil separator. As the refrigerant/oil mix
zigzags its way up the canister the heavier liquid oil
tends to be attracted and attached to the baffle
plates; the accumulating oil then spreads over the
plates until it eventually drips down to the base of
the canister, and then finally drains back to the
compressor crankcase.
Vapour + oil
flow path
High pressure
High pressure
vapour refrigerant
Separated oil return to
compressor crankcase
13.3.9 Receiver (Fig. 13.6)
The receiver is a container which collects the condensed liquid refrigerant and any remaining
vapour from the condenser; this small amount of
vapour will then have enough space to complete the
condensation process before moving to the expansion valve.
Fig. 13.13 Oil separator
2 When the refrigerator is switched off the now
static refrigerant in the evaporator may condense
and enter the suction line and hence the compressor cylinders where it drains over a period of
time into the crankcase and sump.
3 Refrigerant mixing with the lubricant in the
crankcase tends to produce oil frothing which
finds its way past the pistons and piston rings
into the cylinders; above each piston the oil will
then be pumped out into the discharge line with
the refrigerant where it then circulates. Oil does
not cause a problem in the condenser as the
temperature is fairly high so that the refrigerant
remains suspended; however, in the evaporator
the temperature is low so that the liquid oil separates from the refrigerant vapour, therefore
tending to form a coating on the inside bore of
the evaporator coil. Unfortunately oil is a very
poor conductor of heat so that the efficiency of
the heat transfer process in the evaporator is very
much impaired.
13.3.10 Sight glass (Fig. 13.14)
This device is situated in the liquid line on the output side of the receiver; it is essentially a viewing
port which enables the liquid refrigerant to be seen.
Refrigerant movement or the lack of movement
due to some kind of restriction, or bubbling caused
by insufficient refrigerant, can be observed.
13.4 Vapour±compression cycle refrigeration
system with reverse cycle defrosting
(Fig. 13.15(a and b))
A practical refrigeration system suitable for road
transportation as used for rigid and articulated
vehicles must have a means of both cooling and
After these observations it is clear that the refrigerant must be prevented from mixing with the oil
but this is not always possible and therefore an oil
separator is usually incorporated on the output
side of the compressor in the discharge line which
separates the liquid oil from the hot refrigerant
vapour. An oil separator in canister form consists
Liquid line
To drier
Inspection glass
Fig. 13.14 Sight glass
Reverse expansion
valve – cold
Condenser coil
Remote feeler bulb
Remote feeler bulb
Evaporator coil
cvc 2
cvc 3
cvo 4
Reverse cycle
Suction valve
(a) Refrigeration cycle
Fig. 13.15 (a and b)
Check valve
cvo 1
Refrigeration system with reverse cycle defrosting
Reverse expansion
valve – hot
Condenser coil
Remote feeler bulb
Remote feeler bulb
Evaporator coil
valve – cool
cvo 2
Reverse cycle
Check valve
cvc 1
(b) Heating and defrost cycle
Fig. 13.15 Contd
defrosting the cold compartment. The operation of
such a system involving additional valves enables
the system to be switched between cooling and
heat/defrosting, which will now be described.
movement so that the refrigerant in the evaporator
is unable to absorb the heat from the surrounding
atmosphere in the cold storage compartment,
therefore a time will come when the evaporator
must be defrosted.
Heating/defrosting is achieved by temporarily
reversing the refrigerant flow circulation so that
the evaporator becomes a heat dissipating coil and
the condenser converts to a heat absorbing coil.
To switch to the heat/defrosting cycle the pilot
solenoid valve is energized; this causes the solenoid
valve to block the discharge pressure and connect
the suction pressure to the servo cylinder reverse
cycle valve, see Fig. 13.15(b). Subcooled high pressure liquid refrigerant is permitted to flow from the
receiver directly to the now partially opened reverse
thermostatic expansion valve (due to the now
hot remote feeler bulb's increased pressure). The
refrigerant expands in the reverse expansion valve
and accordingly converts to a liquid/vapour; this
then passes through the condenser via the open
check valve (3) in the reverse direction to the normal refrigeration cycle and in the process absorbs
heat from the surroundings so that it comes out as
a low pressure saturated vapour. The refrigerant
then flows to the compressor suction port via the
reverse cycle valve and suction pressure valve but
due to the high surrounding atmospheric temperature it is now superheated vapour. The compressor
then transforms the low pressure superheated
vapour into a high pressure superheated vapour
and discharges it to the evaporator via the oil
separator and reverse cycle valve. Hence the saturated vapour stream dissipates its heat through the
tubing walls to the ice which is encasing the tubing
coil until it has all melted. The refrigerant at the exit
from the evaporator will now be in a saturated
liquid state and is returned to the receiver via the
open check valve (2), sight glass, and open check
valve (5) for the heating/defrosting cycle to be
repeated. Note during the refrigeration cycle the
condenser's reverse expansion valve and remote
feeler bulb sense the reduction in temperature at
the exit from the condenser, thus the corresponding
reduction in internal bulb pressure is relayed to the
reverse expansion valve which therefore closes
during the defrosting cycle. Defrosting is fully automatic. A differential air pressure switch which
senses any air circulation restriction around the
evaporator coil automatically triggers defrosting
of the evaporator coil before ice formation can
reduce its efficiency. A manual defrost switch is
also provided.
13.4.1 Refrigeration cooling cycle (Fig. 13.15(a))
With the pilot solenoid valve de-energized and the
compressor switched on and running the refrigerant commences to circulate through the system
between the evaporator and condenser.
Discharge line pressure from the right hand
compressor cylinder is transferred via the pilot
valve to the reverse cycle valve; this pushes the
slave piston and valves inwards to the left hand
side into the `cooling' position, see Fig. 13.15(a).
Low pressure refrigerant from the receiver flows
via the open check valve (1), sight glass and drier
to the thermostatic expansion valve where rapid
expansion in the valve converts the refrigerant to
a liquid/vapour mixture. Low pressure refrigerant
then passes through the evaporator coil where it
absorbs heat from the cold storage compartment:
the refrigerant then comes out from the evaporator
as low pressure saturated vapour. Refrigerant now
flows to the compressor suction port via the reverse
cycle valve and suction pressure valve as superheated vapour. The compressor now converts the
refrigerant to a high pressure superheated vapour
before pumping it to the condenser inlet via the oil
separator and reverse cycle valve; at this point the
refrigerant will have lost heat to the surroundings
so that it is now in a high pressure saturated vapour
state. It now passes through the condenser where it
gives out its heat to the surrounding atmosphere;
during this process the high pressure refrigerant is
transformed into a saturated liquid. Finally the
main liquid refrigerant flows into the receiver via
the open check valve (4) where there is enough
space for the remaining vapour to condense. This
cycle of events will be continuously repeated as the
refrigerant is alternated between reducing pressure
in the expansion valve before passing through the
evaporator to take heat from the cold chamber, to
increasing pressure in the compressor before passing through the condenser to give off its acquired
heat to the surroundings. Note check valves (1) and
(4) are open whereas check valves (2), (3) and (5)
are closed for the cool cycle.
13.4.2 Heating and defrosting cycle
(Fig. 13.15(b))
With constant use excessive ice may build up
around the evaporator coil; this restricts the air
14 Vehicle body aerodynamics
The constant need for better fuel economy,
greater vehicle performance, reduction in wind
noise level and improved road holding and stability
for a vehicle on the move, has prompted vehicle
manufacturers to investigate the nature of air resistance or drag for different body shapes under
various operating conditions. Aerodynamics is the
study of a solid body moving through the atmosphere
and the interaction which takes place between the
body surfaces and the surrounding air with varying
relative speeds and wind direction. Aerodynamic
drag is usually insignificant at low vehicle speed but
the magnitude of air resistance becomes considerable with rising speed. This can be seen in Fig. 14.1
which compares the aerodynamic drag forces of a
poorly streamlined, and a very highly streamlined
medium sized car against its constant rolling resistance over a typical speed range. A vehicle with a
high drag resistance tends only marginally to
hinder its acceleration but it does inhibit its maximum speed and increases the fuel consumption with
increasing speed.
Body styling has to accommodate passengers and
luggage space, the functional power train, steering,
suspension and wheels etc. thus vehicle design will
conflict with minimizing the body surface drag so
that the body shape finally accepted is nearly always
a compromise.
An appreciation of the fundamentals of aerodynamics and the methods used to counteract
high air resistance for both cars and commercial
vehicles will now be explained.
14.1 Viscous air flow fundamentals
14.1.1 Boundary layer (Fig. 14.2)
Air has viscosity, that is, there is internal friction
between adjacent layers of air, whenever there
is relative air movement, consequently when there
is sliding between adjacent layers of air, energy is
dissipated. When air flows over a solid surface a
thin boundary layer is formed between the main
airstream and the surface. Any relative movement
between the main airstream flow and the surface of
Resisting forces opposing motion (N)
ir r
rolling resistance
Vehicle speed (km/h)
Fig. 14.1 Comparison of low and high aerodynamic drag forces with rolling resistance
Full velocity of air flow
Thickness of boundary
Parabolic rise
in air layer
velocity from inner
to outer boundary
Surface of body
Fig. 14.2
Boundary layer velocity gradient
Direction of
plate drag
Flat plate
Fig. 14.3
Apparatus to demonstrate viscous drag
the body then takes place within this boundary
layer via the process of shearing of adjacent layers
of air. When air flows over any surface, air particles
in intimate contact with the surface loosely attach
themselves so that relative air velocity at the surface becomes zero, see Fig. 14.2. The relative speed
of the air layers adjacent to the arrested air surface
film will be very slow; however, the next adjacent
layer will slide over an already moving layer so that
its relative speed will be somewhat higher. Hence
the relative air velocity further out from the surface
rises progressively between air layers until it attains
the unrestricted main airstream speed.
layer of air moving near the surface. However,
there will be a gradual increase in air speed from
the inner to the outer boundary layer. The thickness of the boundary layer is influenced by the
surface finish. A smooth surface, see Fig. 14.4(b),
allows the free air flow velocity to be reached
nearer the surface whereas a rough surface, see
Fig. 14.4(a), widens the boundary so that the full
air velocity will be reached further out from the
surface. Hence the thicker boundary layer associated with a rough surface will cause more adjacent
layers of air to be sheared, accordingly there
will be more resistance to air movement compared
with a smooth surface.
14.1.2 Skin friction (surface friction drag)
(Fig. 14.3)
This is the restraining force preventing a thin flat
plate placed edgewise to an oncoming airstream
being dragged along with it (see Fig. 14.3), in other
words, the skin friction is the viscous resistance
generated within the boundary layer when air flows
over a solid surface. Skin friction is dependent on the
surface area over which the air flows, the degree of
surface roughness or smoothness and the air speed.
14.1.4 Venturi (Fig. 14.5)
When air flows through a diverging and converging
section of a venturi the air pressure and its speed
changes, see Fig. 14.5. Initially at entry the unrestricted air will be under atmospheric conditions
where the molecules are relatively close together,
consequently its pressure will be at its highest and
its speed at its minimum.
As the air moves into the converging section
the air molecules accelerate to maintain the
volume flow. At the narrowest region in the
venturi the random air molecules will be drawn
14.1.3 Surface finish (Fig. 14.4(a and b))
Air particles in contact with a surface tend to be
attracted to it, thus viscous drag will retard the
Boundary layer
(a) Rough surface
Fig. 14.4 (a and b)
(b) Smooth surface
Influence of surface roughness on boundary layer velocity profile
Accelerating Low
Fig. 14.5 Venturi
Fig. 14.6 Streamline air flow around car
apart thus creating a pressure drop and a faster
movement of the air. Further downstream
the air moves into the diverging or expanding section of the venturi where the air flow decelerates,
the molecules therefore are able to move
closer together and by the time the air reaches the
exit its pressure will have risen again and its
movement slows down.
14.1.5 Air streamlines (Fig. 14.6)
A moving car displaces the air ahead so that the air
is forced to flow around and towards the rear. The
pattern of air movement around the car can be
visualized by airstreamlines which are imaginary
lines across which there is no flow, see Fig. 14.6.
These streamlines broadly follow the contour of
the body but any sudden change in the car's shape
High pressure
low speed
Low pressure
high speed
High pressure
low speed
Low pressure (subatmospheric pressure) high speed
Fig. 14.7
Relative air speed and pressure conditions over the upper profile of a moving car
compels the streamlines to deviate, leaving zones of
stagnant air pockets. The further these streamlines
are from the body the more they will tend to
straighten out.
region it decelerates to cope with the enlarged flow
As can be seen in Fig. 14.8 the pressure conditions over and underneath the car's body can be
plotted from the data; these graphs show typical
pressure distribution trends only. The pressure
over the rear half of the bonnet to the mid-front
windscreen region where the airstream speed is
slower is positive (positive pressure coefficient
Cp), likewise the pressure over the mid-position of
the rear windscreen and the rear end of the boot
where the airstream speed has been reduced is also
positive but of a lower magnitude. Conversely the
pressure over the front region of the bonnet and
particularly over the windscreen/roof leading edge
and the horizontal roof area where the airstream
speed is at its highest has a negative pressure (negative pressure coefficient Cp). When considering the
air movement underneath the car body, the
restricted airstream flow tends to speed up the air
movement thereby producing a slight negative
pressure distribution along the whole underside of
the car. The actual pattern of pressure distribution
above and below the body will be greatly influenced by the car's profile style, the vehicle's speed
and the direction and intensity of the wind.
14.1.6 Relative air speed and pressure conditions
over the upper profile of a moving car (Figs 14.7
and 14.8)
The space between the upper profile of the horizontal outer streamlines relative to the road surface
generated when the body is in motion can be considered to constitute a venturi effect, see Fig. 14.7.
Note in effect it is the car that moves whereas air
remains stationary; however, when wind-tunnel
tests are carried out the reverse happens, air is
drawn through the tunnel with the car positioned
inside on a turntable so that the air passes over and
around the parked vehicle. The air gap between the
horizontal airstreamlines and front end bonnet
(hood) and windscreen profile and the back end
screen and boot (trunk) profile produces a diverging and converging air wedge, respectively. Thus
the air scooped into the front wedge can be considered initially to be at atmospheric pressure and
moving at car speed. As the air moves into the
diverging wedge it has to accelerate to maintain
the rate of volumetric displacement. Over the roof
the venturi is at its narrowest, the air movement
will be at its highest and the air molecules will be
stretched further apart, consequently there will be
a reduction in air pressure in this region. Finally the
relative air movement at the rear of the boot will
have slowed to car speed, conversely its pressure
will have again risen to the surrounding atmospheric pressure conditions, thus allowing the random network of distorted molecules to move closer
together to a more stable condition. As the air
moves beyond the roof into the diverging wedge
14.1.7 Lamina boundary layer (Fig. 14.9(a))
When the air flow velocity is low sublayers within
the boundary layer are able to slide one over the
other at different speeds with very little friction;
this kind of uniform flow is known as lamina.
14.1.8 Turbulent boundary layer (Fig. 14.9(b))
At higher air flow velocity the sublayers within the
boundary layer also increase their relative sliding
speed until a corresponding increase in interlayer
friction compels individual sublayers to randomly
Over top
Pressure coefficient (Cp)
pressure line
Under floor
Fig. 14.8 Pressure distribution above and below the body structure
Outer layer
High velocity
Boundary layer
Thickness of boundary layer
Low velocity
Inner layer
(a) Lamina air flow (low velocity)
Fig. 14.9 (a and b)
(b) Turbulent air flow (high velocity)
Lamina and turbulent air flow
L = Lamina
T = Turbulent
TP = Transition point
(a) Low speed
(b) High speed
Fig. 14.10 (a and b)
Lamina/turbulent boundary layer transition point
break away from the general direction of motion;
they then whirl about in the form of eddies, but still
move along with the air flow.
14.1.10 Flow separation and reattachment
(Fig. 14.11(a and b))
The stream of air flowing over a car's body tends to
follow closely to the contour of the body unless
there is a sudden change in shape, see Fig.
14.11(a). The front bonnet (hood) is usually slightly
curved and slopes up towards the front windscreen,
from here there is an upward windscreen tilt (rake),
followed by a curved but horizontal roof; the rear
windscreen then tilts downwards where it either
merges with the boot (trunk) or continues to slope
gently downwards until it reaches the rear end of
the car.
The air velocity and pressure therefore reaches
its highest and lowest values, respectively, at the top
of the front windscreen; however, towards the rear
of the roof and when the screen tilts downwards
14.1.9 Lamina/turbulent boundary layer
transition point (Fig. 14.10(a and b))
A boundary layer over the forward surface of a
body, such as the roof, will generally be lamina,
but further to the rear a point will be reached called
the transition point when the boundary layer
changes from a lamina to a turbulent one, see Fig.
14.10(a). As the speed of the vehicle rises the transition point tends to move further to the front, see
Fig. 14.10(b), therefore less of the boundary layer
will be lamina and more will become turbulent;
accordingly this will correspond to a higher level
of skin friction.
(a) Notch front and rear windscreens
Attached flow
(b) Very streamlined shape
Fig. 14.11 (a and b)
Flow separation and reattachment
there will be a reduction in air speed and a rise in
pressure. If the rise in air pressure towards the rear
of the car is very gradual then mixing of the airstream with the turbulent boundary layers will be
relatively steady so that the outer layers will be
drawn along with the main airstream, see Fig.
14.11(b). Conversely if the downward slope of the
rear screen/boot is considerable, see Fig. 14.11(a),
the pressure rise will be large so that the mixing rate
of mainstream air with the boundary layers cannot
keep the inner layers moving, consequently the
slowed down boundary layers thicken. Under
these conditions the mainstream air flow breaks
away from the contour surface of the body, this
being known as flow separation. An example of
flow separation followed by reattachment can be
visualized with air flowing over the bonnet and
front windscreen; if the rake angle between the
bonnet and windscreen is large, the streamline
flow will separate from the bonnet and then
reattach itself near the top of the windscreen or
front end of the roof, see Fig. 14.11(a). The space
between the separation and reattachment will then
be occupied by circulating air which is referred to
as a separation bubble, and if this rotary motion is
vigorous a transverse vortex will be established.
14.2 Aerodynamic drag
14.2.1 Pressure (form) drag (Figs 14.12(a±e)
and 14.13)
When viscous air flows over and past a solid form,
vortices are created at the rear causing the flow
of air
– Ve
(a) Flat plate
(b) Circular section
(c) Circular /lobe section
(d) Aerofoil section
(e) Fineness ratio (b/a)
Fig. 14.12 (a±e) Air flow over various shaped sections
to deviate from the smooth streamline flow, see
Fig. 14.12(a). Under these conditions the air flow
pressure in front of the solid object will be higher
than atmospheric pressure while the pressure behind
will be lower than that of the atmosphere, consequently the solid body will be dragged (sucked) in
the direction of air movement. Note that this effect
is created in addition to the skin friction drag. An
extreme example of pressure drag (sometimes
known as form drag) can be seen in Fig. 14.13
where a flat plate placed at right angles to the
air movement will experience a drag force in the
direction of flow represented by the pulley weight
which opposes the movement of the plate.
Pressure drag can be reduced by streamlining any
solid form exposed to the air flow, for instance a
round tube (Fig. 14.12(b)) encourages the air to flow
smoothly around the front half and part of the rear
before flow separation occurs thereby reducing the
resistance by about half that of the flat plate. The
resistance of a tube can be further reduced to about
15% of the flat plate by extending the rear of the
circulating tube in the form of a curved tapering
lobe, see Fig. 14.12(c). Even bigger reductions in
resistance can be achieved by proportioning the
tube section (see Fig. 14.12(d)) with a fineness ratio
a/b of between 2 and 4 with the maximum thickness
b set about one-third back from the nose, see
Fig. 14.12(e). This gives a flow resistance of roughly
one-tenth of a round tube or 5% of a flat plate.
Then let
Mass ˆ m kg
Volume ˆ Q m3
Density ˆ kg=m
kg=m 3
kg 3
m ˆ Q
Density of air flow ˆ kg=m3
Frontal area of plate ˆ A m2
Velocity of air striking surface ˆ v m=s
Volume of air striking
plate per second ˆ Q ˆ vA m3
14.2.2 Air resistance opposing the motion
of a vehicle (Fig. 14.13)
The formula for calculating the opposing resistance
of a body passing though air can be derived as
Let us assume that a flat plate body (Fig. 14.13)
is held against a flow of air and that the air particles
are inelastic and simply drop away from the
perpendicular plate surface. The density of air is
the mass per unit volume and a cubic metre of air
at sea level has an approximate mass of 1.225 kg,
therefore the density of air is 1.225 kg/m3.
Mass movement of air per second ˆ Q ˆ vA
Q ˆ vA
Momentum of this air (mv) ˆ vA v
therefore momentum lost by
air per second ˆ Av2
From Newton's second law the rate at which the
movement of air is changed will give the force
exerted on the plate.
of force
Fig. 14.13 Pressure drag apparatus
force on plate ˆ Av2
tance; conversely a high drag coefficient is caused
by poor streamlining of the body profile so that
there is a high air resistance when the vehicle is in
motion. Typical drag coefficients for various
classes of vehicles can be seen as follows:
However, the experimental air thrust against a
flat plate is roughly 0.6 of the calculated Av2 force.
This considerable 40% error is basically due to the
assumption that the air striking the plate is brought
to rest and falls away, where in fact most of the air
escapes round the edges of the plate and the flow
then becomes turbulent. In fact the theoretical air
flow force does not agree with the actual experimental force (F ) impinging on the plate, but it has
been found to be proportional to Av2
Vehicle type drag coefficient CD
Saloon car
Sports car
Light van
Buses and coaches
Articulated trucks
Ridged truck and draw bar trailer
F / A v2
14.2.5 Drag coefficients and various body shapes
(Fig. 14.15(a±f ))
A comparison of the air flow resistance for different shapes in terms of drag coefficients is presented
as follows:
therefore air resistance F ˆ CD A v2 where CD
is the coefficient of proportionality.
The constant CD is known as the coefficient of
drag, it has no unity and its value will depend upon
the shape of the body exposed to the airstream.
(a) Circular plate (Fig. 14.15(a)) Air flow is head
on, and there is an immediate end on pressure
difference. Flow separation takes place at the
rim; this provides a large vortex wake and a
correspondingly high drag coefficient of 1.15.
14.2.3 After flow wake (Fig. 14.14)
This is the turbulent volume of air produced at the
rear end of a forward moving car and which tends
to move with it, see Fig. 14.14. The wake has a
cross-sectional area equal approximately to that
of the rear vertical boot panel plus the rearward
projected area formed between the level at which
the air flow separates from the downward sloping
rear window panel and the top edge of the boot.
(b) Cube (Fig. 14.15(b)) Air flow is head on but
a boundary layer around the sides delays the
flow separation; nevertheless there is still a large
vortex wake and a high drag coefficient of 1.05.
(c) Sixty degree cone (Fig. 14.15(c)) With the
piecing cone shape air flows towards the
cone apex and then spreads outwards parallel
to the shape of the cone surface. Flow separation however still takes place at the periphery thereby producing a wide vortex wake.
This profile halves the drag coefficient to
about 0.5 compared with the circular plate
and the cube block.
14.2.4 Drag coefficient
The aerodynamic drag coefficient is a measure of
the effectiveness of a streamline aerodynamic body
shape in reducing the air resistance to the forward
motion of a vehicle. A low drag coefficient implies
that the streamline shape of the vehicle's body is
such as to enable it to move easily through the
surrounding viscous air with the minimum of resis-
Fig. 14.14 After flow wake
(a) Circular disc (CD = 1.15)
(d) Sphere (CD = 0.47)
Fig. 14.15 (a±e)
(b) Cube (CD = 1.05)
(e) Hemisphere (CD = 0.42)
(c) 60° cone (CD = 0.5)
(f) Tear drop (CD = 0.05)
Drag coefficient for various shaped solids
(d) Sphere (Fig. 14.15(d)) Air flow towards the
sphere, it is then diverted so that it flows outwards from the centre around the diverging surface and over a small portion of the converging
rear half before flow separation occurs. There is
therefore a slight reduction in the vortex wake
and similarly a marginal decrease in the drag
coefficient to 0.47 compared with the 60 cone.
tion. With these contours the drag coefficient
can be as low as 0.05.
14.2.6 Base drag (Fig. 14.16(a and b))
The shape of the car body largely influences the
pressure drag. If the streamline contour of
the body is such that the boundary layers cling to
a converging rear end, then the vortex area is considerably reduced with a corresponding reduction
in rear end suction and the resistance to motion. If
the body was shaped in the form of a tear drop, the
contour of the body would permit a boundary layer
to continue a considerable way towards the tail
before flow separation occurs, see Fig. 14.16(a),
consequently the area heavily subjected to vortex
swirl and negative pressure will be at a minimum.
However, it is impractical to design a tear drop
body with an extended tapering rear end, but if
the tail is cut off (bobtailed) at the point where
the air flow separates from the contour of the
body (see Fig. 14.16(b)), the same vortex (negative
pressure) exists as if the tail was permitted to converge. The cut off cross-section area where flow
separation would occur is known as the `base
area' and the negative vortex pressure produced is
referred to as the `base drag'. Thus there is a trend
(e) Hemisphere (Fig. 14.15(e)) Air flow towards
and outwards from the centre of the hemisphere. The curvature of the hemisphere gradually aligns with the main direction of flow after
which flow separation takes place on the periphery. For some unknown reason (possibly
due to the very gradual alignment of surface
curvature with the direction of air movement
near the rim) the hemisphere provides a lower
drag coefficient than the cone and the sphere
shapes this, being of the order of 0.42.
(f ) Tear drop (Fig. 14.15(f )) If the proportion of
length to diameter is well chosen, for example
0.25, the streamline shape can maintain a
boundary layer before flow separation occurs
almost to the end of its tail. Thus the resistance
to body movement will be mainly due to viscous
air flow and little to do with vortex wake suc594
Point of
– Ve
+ Ve
(a) Tear drop shaped body
+ Ve
– Ve
(b) Bobtailed tear drop
Fig. 14.16 (a and b)
Base drag
for car manufacturers to design bodies that taper
slightly towards the rear so that flow separation
occurs just beyond the rear axle.
from the front to the rear of the car, the air
moves between the underside and ground, and
over the raised upper body profile surfaces. Thus
if the upper and lower airstreams are to meet at the
rear at a common speed the air moving over the
top must move further and therefore faster than
the more direct underfloor airstream. The air
pressure will therefore be higher in the slower
underfloor airstream than that for the faster airstream moving over the top surface of the car.
Now air moves from high to low pressure regions
so that the high pressure airstream underneath the
car will tend to move diagonally outwards and
upwards towards the low pressure airstream flowing over the top of the body surface (see Fig.
14.18(b)). Both the lower and the upper airstreams
eventually interact along the side-to-top profile
edges on opposite sides of the body to form an
inward rotary air motion that continues to whirl
for some distance beyond the rear end of the forward moving car, see Fig. 14.18(a and b). The
magnitude and intensity of these vortices will to a
great extent depend upon the rear styling of the
14.2.7 Vortices (Fig. 14.17)
Vortices are created around various regions of a
vehicle when it is in motion. Vortices can be
described as a swirling air mass with an annular
cylindrical shape, see Fig. 14.17. The rotary speed
at the periphery is at its minimal, but this
increases inversely with the radius so that its
speed near the centre is at a maximum. However,
there is a central core where there is very little
movement, consequently viscous shear takes place
between adjacent layers of the static core and the
fast moving air swirl; thus the pressure within
the vortex will be below atmospheric pressure,
this being much lower near the core than in the
peripheral region.
14.2.8 Trailing vortex drag (Fig. 14.18(a and b))
Consider a car with a similar shape to a section
of an aerofoil, see Fig. 14.18(a), when air flows
Outer region
Inner region
Inner core
air mass
∴ω = 1
or ωr = a constant
Where V = Liner velocity
ω = Angular velocity
r = Radius
Fig. 14.17 The vortex
Faster airstream low pressure
over upper body surface
Ideal aerofoil car shape
Air moving from
low to high
pressure region
of motion
vortex cone
Slower airstream
and higher air pressure
underneath body
(a) Pictorial view
Trailing vortex
Diagonal airstream
of motion
(b) Plan view
Fig. 14.18 (a and b)
Establishment of trailing vortices
car. The negative (below atmospheric) pressure
created in the wake of the trailing vortices at the
rear of the car attempts to draw it back in the
opposite direction to the forward propelling
force; this resistance is therefore referred to as the
`trailing vortex drag'.
boot (trunk) lid and the boot and rear light panel
tend to generate attached transverse vortices (see
Fig. 14.19(a and b)). The front attached vortices
work their way around the `A' post and then
extend along the side windows to the rear of the
car and beyond. Any overspill from the attached
vortices in the rear window and rear light panel
regions merges and strengthens the side panel vortices (see Fig. 14.19(b)); in turn the products of
these secondary transverse vortices combine and
enlarge the main trailing vortices.
14.2.9 Attached transverse vortices
(Fig. 14.19(a and b))
Separation bobbles which form between the bonnet
(hood) and front windscreen, the rear screen and
Separation bubble
transforms into
transverse vortex
'A' post
(a) Front and side vortices
Side vortex
(b) Rear and side vortices
Fig. 14.19 (a and b)
Notch back transverse and trailing vortices
14.3 Aerodynamic lift
therefore produces a reduction in the tyre to
ground grip. If the uplift between the front and
rear of the car is different, then the slip-angles
generated by the front and rear tyres will not be
equal; accordingly this will result in an under- or
over-steer tendency instead of more neutral-steer
characteristics. Thus uncontrolled lift will reduce
the vehicle's road holding and may cause steering
14.3.1 Lift coefficients
The aerodynamic lift coefficient CL is a measure of
the difference in pressure created above and below
a vehicle's body as it moves through the surrounding viscous air. A resultant upthrust or downthrust
may be produced which mainly depend upon the
body shape; however, an uplift known as positive
lift is undesirable as it reduces the tyre to ground
grip whereas a downforce referred to as negative
lift enhances the tyre's road holding.
14.3.3 Underbody floor height versus
aerodynamic lift and drag
(Figs 14.21(a and b) and 14.22)
With a large underfloor to ground clearance the car
body is subjected to a slight negative lift force
(downward thrust). As the underfloor surface
moves closer to the ground the underfloor air
space becomes a venturi, causing the air to move
much faster underneath the body than over it, see
Figs 14.21(a) and 14.22. Correspondingly with
these changing conditions the air flow pressure on
top of the body will be higher than for the underbody reduced venturi effect pressure, hence there
will be a net down force (negative lift) tending to
increase the contact pressure acting between the
wheels and ground. Conversely a further reduction
in underfloor to ground clearance makes it very
restrictive for the underbody air flow (see Figs
14.21(b) and 14.22), so that much of the airstream
is now compelled to flow over the body instead of
underneath it, which results in an increase in air
speed and a reduction in pressure over the top to
cope with the reduction in the underfloor air
14.3.2 Vehicle lift (Fig. 14.20)
When a car travels along the road the airstream
moving over the upper surface of the body from
front to rear has to move further than the underside
airstream which almost moves in a straight line (see
Fig. 14.20). Thus the direct slower moving underside and the indirect faster moving top side airstream produces a higher pressure underneath the
car than over it, consequently the resultant vertical
pressures generated between the upper and under
surfaces produce a net upthrust or lift. The magnitude of the lift depends mainly upon the styling
profile of both over and under body surfaces, the
distance of the underfloor above the ground, and
the vehicle speed. Generally, the nearer the underfloor is to the ground the greater the positive lift
(upward force); also the positive lift tends to
increase with the square of the vehicle speed. Correspondingly a reduction in wheel load due to the
lift upthrust counteracts the downward load; this
pressure (+ve)
(positive lift)
Faster moving air
greater reduction
in pressure
Higher stagnant
air pressure
Slower moving air Direction
of motion
slight reduction
in pressure
Fig. 14.20 Aerodynamic lift
Low pressure
wake (–ve)
(a) Large ground clearance (negative lift downthrust)
Small reduction in air speed
small increase in pressure
(compared with Fig. 14.21(b))
Small increase in air speed
small reduction in pressure
(compared with Fig. 14.21(b))
Venturi effect
(b) Small ground clearance (positive lift upthrust)
Fast air flow
low pressure
Slow air flow
high pressure
Fig. 14.21 (a and b)
Effects of underfloor to ground clearance on the surrounding air speed, pressure and aerodynamic
movement. Thus the over and under pressure conditions have been reversed which subsequently now
produces a net upward suction, that is, a tendency
toward a positive lift.
will increase, see Fig. 14.23(c and d). Conversely as
the angle of inclination decreases, the lift increases
and the drag decreases; however, as the angle of
inclination is reduced so does the resultant reaction
force. If an aerofoil profile is used instead of the flat
plate, (see Fig. 14.24(a and b)), the airstream over the
top surface now has to move further and faster than
the underneath air movement. This produces a
greater pressure difference between the upper and
lower surfaces and consequently greatly enhances
the aerodynamic lift and promotes a smooth air
flow over the upper profiled surface. A typical relationship between the CL, CD and angle of attack
(inclination angle) is shown for an aerofoil section
in Fig. 14.25.
14.3.4 Aerofoil lift and drag
(Figs 14.23(a±d), 14.24(a and b) and 14.25)
Almost any object moving through an airstream
will be subjected to some form of lift and drag.
Consider a flat plate inclined to the direction of
air flow, the pressure of air above the surface of
the plate is reduced while that underneath it is
increased. As a result there will be a net pressure
on the plate striving to force it both upwards and
backwards, see Fig. 14.23(a). It will be seen that
the vertical and horizontal components of the
resultant reaction represents both lift and drag
respectively, see Fig. 14.23(b). The greater the
angle of inclination, the smaller will be the upward
lift component, while the backward drag component
14.3.5 Front end nose shape (Fig. 14.26(a±c))
Optimizing a protruding streamlined nose profile
shape influences marginally the drag coefficient
and to a greater extent the front end lift coefficient.
Small ground
Lift coefficient (CL)
Car height range
Venturi effect range
Free stream range
Negative lift downthrust
Positive lift upthrust
Fig. 14.22 Aerodynamic lift versus ground, floor height
With a downturned nose (see Fig. 14.26(a)) the
streamlined nose profile directs the largest proportion of the air mass movement over the body, and
only a relatively small amount of air flows underneath the body. If now a central nose profile is
adopted (see Fig. 14.26(b)) the air mass movement
is shared more evenly between the upper and lower
body surfaces; however, the air viscous interference with the underfloor and ground still causes the
larger proportion of air to flow above than below
the car's body. Conversely a upturned nose (see
Fig. 14.26(c)) induces still more air to flow
beneath the body with the downward curving
entry gap shape producing a venturi effect.
Consequently the air movement will accelerate
before reaching its highest speed further back at
its narrowest body to ground clearance. Raising
the mass airflow in the space between the body
and ground increases the viscous interaction of the
air with the under body surfaces and therefore
forces the air flow to move diagonally out and
upward from the sides of the car. It therefore
strengthens the side and trailing vortices and as a
result promotes an increase in front end aerodynamic lift force.
The three basic nose profiles discussed showed,
under windtunnel tests, that the upturned nose had
the highest drag coefficient CD of 0.24 whereas
there was very little difference between the central
and downturned nose profiles which gave drag
coefficients CD of 0.223 and 0.224 respectively.
However the front end lift coefficient CL for the
three shapes showed a marked difference, here the
upturned nose profile gave a positive lift coefficient
CL of ‡0.2, the central nose profile provided an
almost neutral lift coefficient CL of ‡0.02, whereas
the downturned nose profile generated a negative
lift coefficient CL of 0.1.
Direction of
air flow
Direction of
air flow
Drag force
(a) Reaction force on an inclined plate
(b) Lift and drag components on an inclined plate
Large lift
Angle of
(angle of attack)
(d) Large angle of inclination
(c) Small angle of inclination
Fig. 14.23 (a±d)
Lift and drag on a plate inclined at a small angle to the direction of air flow
14.4 Car body drag reduction
rounding the edges; however, beyond 40 mm radius
there was no further advantage in increasing the edge
radius or chamfer.
14.4.1 Profile edge rounding or chamfering
(Fig. 14.27(a and b))
There is a general tendency for aerodynamic lift
and drag coefficients to decrease with increased
edge radius or chamfer: experiments carried out
showed for a particular car shape (see Fig. 14.27(a))
how the drag coefficient was reduced from 0.43 to
0.40 with an edge radius/chamfer increasing from
zero to 40 mm (see Fig. 14.27(b)), and there was a
slightly greater reduction with chamfering than
Direction of
air flow
14.4.2 Bonnet slope and windscreen rake
(Fig. 14.28(a±c))
Increasing the bonnet (hood) slope angle a from
zero to roughly 10 reduces the drag coefficient,
but beyond 10 the drag reduction is insignificant,
see Fig. 14.28(b). Likewise, increasing the rake angle
g reduces the drag coefficient (see Fig. 14.28(c))
particularly when the rake angle becomes large;
Direction of
air flow
Turbulent flow
over upper surface
Smooth flow over
upper surface
Increased air speed
lower pressure
Up wash
(a) Inclined plate
Fig. 14.24 (a and b)
air speed
higher pressure
(b) Inclined aerofoil
Air flow over a flat plate and aerofoil inclined at a small angle
Down wash
Drag coefficient (CD)
Lift coefficient (CL)
Angle of attack (Θ) deg
Fig. 14.25 Lift and drag coefficient versus angle of inclination (attack)
however, very large rake angles may conflict with
the body styling.
(a and b) which shows a marked reduction in the
drag coefficient with both a 50 mm and then a
125 mm rear end contraction on either side of the
car; however, there was no further reduction in the
drag coefficient when the rear end contraction was
increased to 200 mm.
14.4.3 Roof and side panel cambering
(Figs 14.29(a and b) and 14.30(a and b))
Cambering the roof (see Fig. 14.29(a and b)) and the
side panels (see Fig. 14.30(a and b)) reduces the drag
coefficient. However, if the roof camber curvature
becomes excessive the drag coefficient commences
to rise again (see Fig. 14.29(b)), whereas the reduction in drag coefficient with small amounts of sidepanel cambering is marked (see Fig. 14.30(b)), but
with excessive camber the reduction in the drag
coefficient becomes only marginal. Both roof and
side panel camber should not be increased at the
expense of enlarging the frontal area of the car as
this would in itself be counter-productive and would
increase the drag coefficient.
14.4.5 Underbody rear end upward taper
(Fig. 14.32(a and b))
Tilting upwards the underfloor rear end produces
a diffuser effect which shows a promising way to
reduce the drag coefficient, see Fig. 14.32(a and b).
However, it is important to select the optimum
ratio of length of taper to overall car length and
the angle b of upward inclination for best results.
14.4.6 Rear end tail extension
(Fig. 14.33(a and b))
Windtunnel investigation with different shaped tail
models have shown that the minimal drag coefficients were produced with extended tails, see Fig.
14.33(a and b), but this shape would be impractical
for design reasons. Conversely if the rear end tail is
14.4.4 Rear side panel taper
(Fig. 14.31(a and b))
Tapering inwards the rear side panel reduces the
drag coefficient. This can be seen in Fig. 14.31
Small trailing
Large mass
air flow overhead
Small mass
air flow below
Small diagonal
side flow
(a) Downturned nose profile
Medium mass
air flow overhead
Medium trailing
Medium mass
air flow below
Medium side
air flow
(b) Central nose profile
Small mass
air flow overhead
Large mass
air flow below
Large side
air flow
Large trailing
Viscous air
(c) Upturned nose profile
Fig. 14.26 (a±c) A greatly exaggerated air mass distribution around a car body for various nose profiles
cropped at various lengths and curved downwards
there is an increase in the drag coefficient with each
reduction in tail length beyond the rear wheels.
edge rounding and general shape dictates the drag
resistance. Moulding in individual compartments in
the underfloor pan to house the various components
and if possible enclosing parts of the underside with
plastic panels helps considerably to reduce the drag
resistance. The underside of a body has built into it
many cavities and protrusions to cater with the
following structural requirements and operating
14.4.7 Underbody roughness
(Fig. 14.34(a and b))
The underbody surface finish influences the drag
coefficient just as the overbody curvature, tapering,
Drag coefficient (CD)
Radius or chamfer (mm)
Influence of forebody bonnet (hood) edge shape on drag coefficient
Fig. 14.27 (a and b)
rake angle
Fig. 14.28 (a±c)
Bonnet slope angle (α) deg
Drag coefficient (CD)
Drag coefficient (CD)
Drag coefficient (CD)
Windscreen rake angle (γ) deg
Bonnet slope and windscreen rake angle versus drag coefficient
Roof camber
Change in drag coefficient (CD)
Roof camber (h/l)
Effect of roof camber on drag coefficient
Body side panel camber
Change in drag coefficient (CD)
Fig. 14.29 (a and b)
Side panel camber (x/L)
Fig. 14.30 (a and b)
Effect of side panel camber on drag coefficient
Fig. 14.31 (a and b)
Effect of rear side panel taper on drag coefficient
Change in drag coefficient (CD)
– 0.01
– 0.02
– 0.03
t/ = 0.5
t/ = 0.2
– 0.04
– 0.05
Fig. 14.32 (a and b)
Diffuser angle (β) deg
Effect of rear end upward taper on drag coefficient
components: front and rear wheel and suspension
arch cavities, engine, transmission and steering compartment, side and cross member channelling, floor
pan straightening ribs, jacking point straightening
channel sections, structural central tunnel and rear
wheel drive propeller shaft, exhaust system catalytic
converter, silencer and piping, hand brake cable,
and spare wheel compartment etc. A rough underbody produces excessive turbulence and friction
losses and consequently raises the drag coefficient,
whereas trapped air in the underside region slows
down the air flow and tends to raise the underfloor
pressure and therefore positive lift force. Vehicles
with high drag coefficients gain least by smoothing
the underside. The underfloor roughness or depth of
irregularity defined as the centre line average peak
to valley height for an average car is around
‡150 mm. A predictable relationship between the
centre line average roughness and the drag coefficient for a given ground clearance and vehicle length
is shown in Fig. 14.34(a and b).
underfloor air dam the underfloor air flow pressure
increase raises the aerodynamic upthrust, that is, it
produces positive lift, see Fig. 14.35(a). Conversely
a front end air dam reduces the underfloor air
flow pressure thereby generating an aerodynamic
downthrust, that is, it produces negative lift (see
Fig. 14.35(b)). Experimental results show with a front
end dam there is a decrease in front lift (negative
lift) whereas there is a slight rise in rear end lift
(positive lift) as the dam height is increased, and
as would be expected, there is also a rise in drag as
the frontal area of the dam is enlarged, see
Fig. 14.35(c).
14.5.2 Exposed wheel air flow pattern
(Fig. 14.36(a±c))
When a wheel rotates some distance from the
ground air due to its viscosity attaches itself to the
tread and in turn induces some of the surrounding
air to be dragged around with it. Thus this concentric movement of air establishes in effect a weak
vortex, see Fig. 14.36(a). If the rotating wheel is in
contact with the ground it will roll forwards which
makes windtunnel testing under these conditions
difficult; this problem is overcome by using a supportive wheel and floor rig. The wheel is slightly
submerged in a well opening equal to the tyre width
and contact patch length for a normal loaded wheel
and a steady flow of air is blown towards the
frontal view of the wheel. With the wheel rig simulating a rotating wheel in contact with the ground,
the wheel vortex air movement interacts and distorts the parallel main airstream.
14.5 Aerodynamic lift control
14.5.1 Underbody dams (Fig. 14.35(a±c))
Damming the underbody to ground clearance at
the extreme rear blocks the underfloor airstream
and causes a partial pressure build-up in this
region, see Fig. 14.25(a), whereas locating the
underbody dam in the front end of the car joins
the rear low pressure wake region with the underfloor space, see Fig. 14.35(b). Thus with a rear end
5650 mm
5720 mm
Drag coefficient (CD)
Fig. 14.33 (a and b)
Tail extension (mm)
Effect of rear end tail extension on drag coefficient
The air flow pattern for an exposed wheel can be
visualized and described in the following way. The
air flow meeting the lower region of the wheel will
be stagnant but the majority of the airstream will
flow against the wheel rotation following the contour of the wheel until it reaches the top; it then
separates from the vortex rim and continues to flow
towards the rear but leaving underneath and in the
wake of the wheel a series of turbulent vortices, see
Fig. 14.36(b). The actual point of separation will
creep forwards with increased rotational wheel
speed. Air pressure distribution around the wheel
will show a positive pressure build-up in the stagnant air flow front region of the wheel, but this
changes rapidly to a high negative pressure where
the main air flow breaks away from the wheel rim,
see Fig. 14.36(c). It then declines to some extent
beyond the highest point of the wheel, and then
remains approximately constant around the rear
wake region of the wheel. Under these described
conditions, the exposed rotating wheel produces
a resultant positive upward lift force which tends
to reduce the adhesion between the tyre tread and
Drag coefficient (CD)
surface profile
Roughness (centre line average) ± mm
Effect of underbody roughness on drag coefficient
lift (+ve)
Front end
Semi high
Air dam
(rear & partial sides)
Air dam
(front & partial sides)
Wake negative
pressure (–ve)
lift (–ve)
Fig. 14.34 (a and b)
High speed
low pressure
High speed
low pressure
Rear end
(b) Front end underbody air dam
Change in drag and lift coefficients (CD and CL)
(a) Rear end underbody air dam
Rear lift
– 0.1
– 0.2
t li
– 0.3
Dam height (mm)
Fig. 14.35 (a±c)
Effects of underbody front and rear end air dams relative to the lift and drag coefficient
lift (+ve)
Point of
Direction of
wheel movement
(a) Wheel rotation in still air away from the ground
Low pressure
High pressure
(b) Air flow pattern with wheel rolling on the ground
lift (+ve)
pressure (–ve)
of motion
pressure (+ve)
(c) Air pressure distribution with wheel rolling on the ground
Fig. 14.36 (a±c) Exposed wheel air flow pattern and pressure distribution
14.5.3 Partial enclosed wheel air flow pattern
(Figs 14.37(a and b) and 14.38(a±c))
The air flow passing beneath the front of the
car initially moves faster than the main airstream,
this therefore causes a reduction in the local air
pressure. At the rear of the rotating wheel due to
viscous drag air will be scooped into the upper
space formed between the wheel tyre and the
wheel mudguard arch (see Fig. 14.37(a and b)). The
air entrapped in the wheel arch cavity circulates
towards the upper front of the wheel due to a
slight pressure build-up and is then expelled
through the front end wheel to the mudguard gap
which is at a lower pressure in both a downward
and sideward direction. Decreasing the clearance
between the underside and the ground and shielding more of the wheel with the mudguard tends to
produce a loss of momentum to the air so that both
(a) Side view
Flow separation
(b) Plan view
Fig. 14.37 (a and b)
Wheel arch air flow pattern
aerodynamic lift CLW and drag CDW coefficients,
and therefore forces, are considerably reduced
Fig. 14.38(a±c).
height. However, this is at the expense of a slight
rise in the front end lift coefficient, whereas the
drag coefficient initially decreases and then marginally rises again with increased spoiler lip height.
It should be appreciated that the break-up of the
smooth streamline air flow and the increase in rear
downward pressure should if possible be achieved
without incurring too much, if any, increase in
front end positive lift and aerodynamic drag.
14.5.4 Rear end spoiler (Fig. 14.39(a±c))
Generally when there is a gentle rear end body
profile curvature change, it will be accompanied
with a relatively fast but smooth streamline air flow
over this region which does not separate from the
upper surface. However, this results in lower local
pressures which tend to exert a lift force (upward
suction) at the rear end of the car. A lip, see Fig.
14.39(a), or small aerofoil spoiler, see Fig. 14.39(b),
attached to the rear end of the car boot (trunk)
interrupts the smooth streamline air flow thereby
slowing down the air flow and correspondingly
raising the upper surface local air pressure which
effectively increase the downward force known as
negative lift. A typical relationship between rear
lift, front lift and drag coefficients relative to the
spoiler lip height is shown in Fig. 14.39(c). The
graph shows a general increase in negative lift
(downward force) by increasing the spoiler lip
14.5.5 Negative lift aerofoil wings
(Fig. 14.40(a±c))
A negative lift wing is designed when attached to
the rear end of the car to produce a downward
thrust thereby enabling the traction generated by
the rear driving wheels to be increased, or if a
forward negative lift wing is fitted to improve the
grip of both front steering wheels.
With the negative lift wing the aerofoil profile is
tilted downward towards the front end with the
negative and positive aerofoil section camber at
the top and bottom respectively, see Fig. 14.40(a).
The airstream therefore moving underneath the
(a) Side view
(b) Front view
Wheel lift coefficient (CLW)
Wheel drag coefficient(CDW)
Fig. 14.38 (a±c) Effect of underside ground clearance on both lift and drag coefficients
aerofoil wing has to move further and faster than
the airstream flowing over the upper surface; the
pressure produced below the aerofoil wing is therefore lower than above. Consequently there will be a
resultant downthrust perpendicular to the cord of
the aerofoil (see Fig. 14.40(b)) which can be resolved
into both a vertical downforce (negative lift) and a
horizontal drag force. Enlarging the tilt angle of the
wing promotes more negative lift (downthrust) but
this is at the expense of increasing the drag force
opposing the forward movement of the wing, thus
a compromise must always be made between
improving the downward wheel grip and the extra
drag force opposing the motion of the car. Racing
cars have the aerofoil wing over the rear wheel
axles or just in front or behind them, see Fig.
14.40(c). However, the drag force produces a clockwise tilt which tends to lift the front wheels of the
ground, therefore the front aerofoil wings (see Fig.
14.40(c)) are sometimes attached low down and
slightly ahead of the front wheels to counteract
the front end lift tendency.
14.6 Afterbody drag
14.6.1 Squareback drag (Figs 14.41 and 14.42)
Any car with a rear end (base) slope surface angle
ranging from 90 to 50 is generally described as a
squareback style (see Fig. 14.42). Between this
angular surface inclination range for a squareback
car there is very little change in the air flow pattern
High speed
low pressure
Reduced speed
increase in pressure
Lip spoiler
Change in lift and drag coefficients (CL and CD)
lift tendency
High speed
low pressure
Front lift
speed increase
in pressure
lift tendency
Lip height (mm)
Fig. 14.39 (a±c)
Effect of rear end spoiler on both lift and drag coefficients
and therefore there is virtually no variation in the
afterbody drag (see Fig. 14.41). With a parallel
sided squareback rear end configuration, the
whole rear surface area (base area) becomes an
almost constant low negative pressure wake region.
Tapering the rear quarter side and roof of the body
and rounding the rear end tends to lower the base
pressure. In addition to the base drag, the afterbody drag will also include the negative drag due to
the surrounding inclined surfaces.
ing attached to the body from the rear of the roof
to the rear vertical light-plate and at the same time
the condition which helps to generate attached and
trailing vortices with the large sloping rear end is
no longer there. Consequently the only rearward
suction comes from the vertical rear end projected
base area wake, thus as the rear end inclined angle
diminishes, the drag coefficient decreases, see Fig.
14.41. However, as the angle approaches zero there
is a slight rise in the drag coefficient again as the
rear body profile virtually reverts to a squareback
style car.
14.6.2 Fastback drag (Figs 14.41 and 14.43)
When the rear slope angle is reduced to 25 or less
the body profile style is known as a fastback, see
Fig. 14.43. Within this much reduced rear end inclination the airstream flows over the roof and rear
downward sloping surface, the airstream remain-
14.6.3 Hatchback drag (Figs 14.41, 14.44 and
Cars with a rear sloping surface angle ranging from
50 to 25 are normally referred to as hatchback
Slower airstream
higher pressure
air flow
Faster airstream
lower pressure
(a) Air streamlining for an inclined
negative lift aerofoil wing
(b) Lift and drag components on
an inclined negative lift wing
Rear negative
lift wing
FD h
Front negative
lift wing
wheel load
Wheel base (L)
(c) Racing car incorporating negative lift wings
Fig. 14.40 (a±c) Negative lift aerofoil wing considerations
style, see Fig. 14.44. Within this rear end inclination range air flows over the rear edge of the roof
and commences to follow the contour of the rear
inclined surface; however, due to the steepness of
the slope the air flow breaks away from the surface.
At the same time some of the air flows from the
higher pressure underfloor region to the lower pressure roof and rear sloping surface, then moves
slightly inboard and rearward along the upper
downward sloping surface. The intensity and direction of this air movement along both sides of the
rear upper body edging causes the air to spiral into
a pair of trailing vortices which are then pushed
downward by the downwash of the airstream
flowing over the rear edge of the roof, see
Fig. 14.45. Subsequently these vortices re-attach
themselves on each side of the body, and due to
the air's momentum these vortices extend and trail
well beyond the rear of the car. Hence not only
does the rear negative wake base area include the
vertical area and part of the rearward projected
slope area where the airstream separates from the
body profile, but it also includes the trailing conical
vortices which also apply a strong suction pull
against the forward motion of the car. As can be
seen in Fig. 14.41 there is a critical slope angle
range (20 to 35 ) in which the drag coefficient
rises steeply and should be avoided.
Drag coefficient (CD)
Critical angle
Slope angle (deg)
Fig. 14.41 Effect of rear panel slope angle on the afterbody drag
Rear screen
attachment and
90° –50°
Fig. 14.42 Squareback configuration
Rear screen
Flow separation
Fig. 14.43 Fastback configuration
and separation
Rear screen
Fig. 14.44 Hatchback configuration
Fig. 14.45 Hatchback transverse and trailing vortices
14.6.4 Notchback drag (Figs 14.46, 14.47(a and b)
and 14.48(a and b))
A notchback car style has a stepped rear end body
profile in which the passenger compartment rear
window is inclined downward to meet the horizontal
rearward extending boot (trunk) lid (see Fig. 14.46).
With this design, the air flows over the rear roof
edge and follows the contour of the downward
sloping rear screen for a short distance before
separating from it; however, the downwash of the
airstream causes it to re-attach itself to the body
near the rear end extended boot lid. Thus the basewake area will virtually be the vertical rear boot
and light panel; however, standing vortices will be
generated on each side of the body just inboard on
the top surface of the rear window screen and boot
lid, and will then be projected in the form of trailing
conical vortices well beyond the rear end of the
boot, see Fig. 14.19(b). Vortices will also be created
along transverse rear screen to boot lid junction
and across the rear of the panel light.
Experiments have shown (see Fig. 14.47(a)) that
the angle made between the horizontal and the
inclined line touching both the rear edges of the
roof and the boot is an important factor in determining the afterbody drag. Fig. 14.47(b) illustrates
the effect of the roof to boot line inclination;
when this angle is increased from the horizontal
the drag coefficient commences to rise until reaching a peak at an inclination of roughly 25 , after
which the drag coefficient begins to decrease. From
this it can be seen that raising the boot height or
extending the boot length decreases the effective
inclination angle e and therefore tends to reduce
the drag coefficient. Conversely a very large effective
inclination angle e will also cause a reduction in the
attachment and
attachment and separation
Change in drag coefficient (∆CD)
Fig. 14.46 Notchback configuration
Critical angle
– 0.4
Fig. 14.47 (a and b)
Φe = rear effective slope angle
Effective slope angle (Φe) deg
Influence of the effective slope angle on the drag coefficient
drag coefficient but at the expense of reducing the
volume capacity of the boot. The drag coefficient
relative to the rear boot profile can be clearly illustrated in a slightly different way, see Fig. 14.48(a).
Here windtunnel tests show how the drag coefficient can be varied by altering the rear end profile
from a downward sloping boot to a horizontal
boot and then to a squareback estate shape. It
will be observed (see Fig 14.48(b)) that there is a
critical increase in boot height in this case from 50
to 150 mm when the drag coefficient rapidly
decreases from 0.42 to 0.37.
horse. Cabriolet these days describes a car with
a folding roof such as a sports (two or four seater)
or roadster (two seater) car. These cars may be
driven with the folding roof enclosing the cockpit
or with the soft roof lowered and the side screen
windows up or down. Streamlining is such that the
air flow follows closely to the contour of the nose
and bonnet (hood), then moves up the windscreen
before overshooting the screen's upper horizontal
edge (see Fig. 14.49). If the rake angle of the windscreen is small (such as with a high mounted off
road four wheel drive vehicle) the airstream will be
deflected upward and rearward, but with a large
rake angle windscreen the airstream will not rise
much above the windscreen upper leading edge
as the air flows over the open driver/passenger
14.6.5 Cabriolet cars (Fig. 14.49)
A cabriolet is a French noun and originally referred
to a light two wheeled carriage drawn by one
Fig. 14.48 (a and b)
Drag coefficient (CD)
sloping boot
Boot (trunk) height (h) mm
Effect of elevating the boot (trunk) height on the drag coefficient
compartment towards the rear of the car. A separation bubble forms between the airstream and
the exposed and open seating compartment, the
downstream air flow then re-attaches itself to the
upper face of the boot (trunk). However, this bubble is unstable and tends to expand and burst in a
cyclic fashion by the repetition of separation and
re-attachment of the airstream on top of the boot
(trunk), see Fig. 14.49. Thus the turbulent energy
causes the bubble to expand and collapse and the
fluctuating wake area (see Fig. 14.49), changing
between h1 and h2 produces a relatively large drag
resistance. With the side windscreens open air is
drawn into the low pressure bubble region and in
the process strong vortices are generated at the side
entry to the seating compartment; this also therefore contributes to the car drag resistance. Typical
drag coefficients for an open cabriolet car are given
Header rail
14.7 Commercial vehicle aerodynamic
14.7.1 The effects of rounding sharp front cab
body edges (Fig. 14.50(a±d))
A reduction in the drag coefficient of large vehicles
such as buses, coaches and trucks can be made by
rounding the front leading edges of the vehicle.
Roll over
as follows: folding roof raised and side screens up
CD 0.35, folding roof down and side screens up CD
0.38, and folding roof and side screens down CD
0.41. Reductions in the drag coefficient can be
made by attaching a header rail deflector, streamlining the roll over bar and by neatly storing or
covering the folding roof, the most effective device
to reduce drag being the header rail deflector.
Fig. 14.49 Open cabriolet
Side flow
CD = 0.88
(a) Coach with sharp leading edges
Flow almost remains attached
CD = 0.36
(b) Coach with rounded leading edges
Flow remains attached
CD = 0.34
(c) Coach with rounded edges and backsloping front
(d) Effect of rounding vehicle leading edges
upon the aerodynamic drag
Change in drag coefficient (∆CD)
Leading edge radius (R) mm
Fig. 14.50 (a±d)
Forebody coach streamlining
Simulated investigations have shown a marked
decrease in the drag coefficient from having sharp
forebody edges (see Fig. 14.50(a)) to relatively large
round leading edge radii, see Fig. 14.50(b). It can
be seen from Fig. 14.50(d) that the drag coefficient
progressively decreased as the round edge radius
was increased to about 120 mm, but there was only
a very small reduction in the drag coefficient with
further increase in radii. Thus there is an optimum
radius for the leading front edges, beyond this there
is no advantage in increasing the rounding radius.
The reduction in the drag coefficient due to rounding the edges is caused mainly by the change from
flow separation to attached streamline flow for
both cab roof and side panels, see Fig. 14.50(a
and b). However, sloping back the front profile of
the coach to provide further streamlining only
made a marginal reduction in the drag coefficient,
see Fig. 14.50(c).
then flow between the cab and trailer body, but the
majority flows over the trailer roof leading edge
and attaches itself only a short distance from the
front edge of the trailer roof thereby producing a
relatively low drag coefficient, see Fig. 14.51(c).
With the medium height trailer body the air flow
remains attached to the cab roof; some air flow
again impinges on the front of the trailer body
and is deflected between the cab and trailer body,
but most of the air flow hits the trailer body leading
edge and is deflected slightly upwards and only reattaches itself to the upper surface some distance
along the trailer roof. This combination therefore
produces a moderate rise in the drag coefficient, see
Fig. 14.51(c). In the extreme case of having a very
high trailer body the air flow over the cab still
remains attached and air still flows downwards
into the gap made between the cab and trailer;
however, more air impinges onto the vertical front
face of the trailer body and the deflection of the air
flow over the leading edge of the trailer body is
even steeper than in the case of the medium height
trailer body. Thus re-attachment of the air flow
over the roof of the trailer body takes place much
further along its length so that a much larger roof
area is exposed to air turbulence; consequently there
is a relatively high drag coefficient, see Fig. 14.51(c).
14.7.2 The effects of different cab to trailer body
heights with both sharp and rounded upper
windscreen leading edges (Fig. 14.51(a±c))
A generalized understanding of the air flow over
the upper surface of an articulated cab and trailer
can be obtained by studying Fig. 14.51(a and b).
Three different trailer heights are shown relative to
one cab height for both a sharp upper windscreen
leading edge (Fig. 14.51(a)) and for a rounded
upper windscreen edge (Fig. 14.51(b)). It can be
seen in the case of the sharp upper windscreen
leading edge cab examples (Fig. 14.51(a)) that
with the low trailer body the air flow cannot follow
the contour of the cab and therefore overshoots
both the cab roof and the front region of the trailer
body roof thereby producing a relatively high coefficient of drag, see Fig. 14.51(c). With the medium
height trailer body the air flow still overshoots
(separates) the cab but tends to align and attach
itself early to the trailer body roof thereby producing a relatively low coefficient of drag, see Fig.
14.51(c). However, with the high body the air flow
again overshoots the cab roof; some of the air then
hits the front of the trailer body, but the vast
majority deflects off the trailer body leading edge
before re-attaching itself further along the trailer
body roof. Consequently the disrupted air flow
produces a rise in the drag coefficient, see Fig.
In the case of the rounded upper windscreen
leading edge cab (see Fig. 14.51(b)), with a low
trailer body the air flowing over the front windscreen remains attached to the cab roof, a small
proportion will hit the front end of the trailer body
14.7.3 Forebody pressure distribution
(Fig. 14.52(a and b))
With both the conventional cab behind the engine
and the cab over or in front of the engine tractor
unit arrangements there will be a cab to trailer gap
to enable the trailer to be articulated when the
vehicle is being manoeuvred. The cab roof to trailer
body step, if large, will compel some of the air flow
to impinge on the exposed front face of the trailer
thereby producing a high pressure stagnation
region while the majority of air flow will be
deflected upwards. As it brushes against the upper
leading edge of the trailer the air flow then separates from the forward region of the trailer roof
before re-attaching itself further along the flat
roof surface, see Fig. 14.52(a). As can be seen the
pressure distribution shows a positive pressure
(above atmospheric pressure) region air spread
over the exposed front face of the trailer body
with its maximum intensity (stagnant region) just
above the level of the roof; this contrasts the negative pressure (below atmospheric pressure) generated air flow in the forward region of the trailer
roof caused by the air flow separation turbulence.
Note the negative pressure drops off towards the
rear of the roof due to air flow re-attachment.
Highest CD
Medium CD
Lowest CD
(a) Tractor cab with sharp windscreen/roof leading edge (flow separation over cab roof)
Lowest CD
Medium CD
Highest CD
(b) Tractor cab with rounded windscreen/roof leading edge (attached air flow over cab roof)
Drag coefficient (CD)
air flow
over roof
Air flow
over roof
Body height (h) m
(c) Influence of cab to body height and cab shape
upon the drag coefficient
Fig. 14.51 (a±c) Comparison of air flow conditions with both sharp and rounded roof leading edge cab with various trailer
body heights
Trailer roof
pressure distribution
Trailer front
panel pressure
(a) Cab without roof deflector
(b) Cab with roof deflector
Fig. 14.52 (a and b)
Trailer flow body pressure distribution with and without cab roof deflector
By fitting a cab roof deflector the pattern of air
flow is diverted upwards and over the roof of the
trailer body, there being only a slight degree of flow
separation at the front end of the trailer body roof,
see Fig. 14.52(b). Consequently the air flow moves
directly between the cab roof deflector and the roof
of the trailer body; it thus causes the air pressure in
the cab to trailer gap to decrease, this negative pressure being more pronounced on the exposed upper
vertical face of the trailer, hence the front face upper
region of the trailer will actually reduce that portion
of drag produced by the exposed frontal area of the
trailer. Conversely the negative pressure created by
the air flow over the leading edge of the roof falls
rapidly, indicating early air flow re-attachment.
the increase in the trailer body to cab height ratio up
to a t/c ratio of 1.5, which is equivalent to the maximum body height of 4.2 m; this corresponded to a
maximum drag coefficient of 0.86. Hence increasing
the trailer body step height ratio from 1.2 to 1.5
increases the step height from 0.56 m to 1.4 m and
in turn raises the drag coefficient from 0.63 to 0.86.
The rise in drag coefficient of 0.23 is considerable
and therefore streamlining the air flow between the
cab and trailer body roof is of great importance.
14.8 Commercial vehicle drag reducing devices
14.8.1 Cab roof deflectors (Figs 14.54(a and b),
14.55(a and b) and 14.56(a±c))
To partially overcome the large amount of extra
drag experienced with a cab to trailer height mismatch a cab roof deflector is commonly used. This
device prevents the air movement over the cab roof
impinging on the upper front of the trailer body
and then flowing between the cab and trailer gap,
see Fig. 14.54(a). Instead the air flow is diverted by
the uptilted deflector surface to pass directly
between the cab to trailer gap and then to flow
relatively smoothly over the surface of the trailer
roof, see Fig. 14.54(b). These cab roof deflectors
are beneficial in reducing the head on air flow but
they do not perform so well when subjected to side
winds. Slight improvements can be made to prevent
air flowing underneath and across the deflector
plate by enclosing the sides; this is usually achieved
14.7.4 The effects of a cab to trailer body roof
height step (Fig. 14.53(a and b))
Possibly the most important factor which contributes to a vehicle's drag resistance is the exposed
area of the trailer body above the cab roof relative
to the cab's frontal area (Fig. 14.53(a)). Investigation
into the forebody drag of a truck in a windtunnel
has been made where the trailer height is varied
relative to a fixed cab height. The drag coefficient
for different trailer body to cab height ratios (t/c)
were then plotted as shown in Fig. 14.53(b). For
this particular cab to trailer combination dimensions there was no noticeable change in the drag
coefficient C of 0.63 with an increase in trailer body
to cab height ratio until about 1.2, after which the
drag coefficient commenced to rise in proportion to
0.2 m
Drag coefficient (CD)
Trailer to cab height step ratio = ct
1.8 m
t /c
Fig. 14.53 Influence of cab to trailer body height upon the drag coefficient
Flow re-attachment
(a) Cab to trailer height mismatch
Fig. 14.54 (a and b)
(b) Cab to trailer height mismatch
bridged with roof deflector
Air flow between cab and trailer body with and without cab roof deflector
by using a fibre glass or plastic moulded deflector,
see Fig. 14.55(b).
If trailers with different heights are to be coupled
to the tractor unit while in service, then a mismatch
of the deflector inclination may result which will
again raise the aerodynamic drag. There are some
cab deflector designs which can adjust the tilt of the
cab deflector to optimize the cab to trailer air flow
transition (see Fig. 14.55(a)), but in general altering
the angle setting would be impractical. How the
cab roof deflector effectiveness varies with deflector plate inclination is shown in Fig. 14.56(c) for
both a narrow and a wide cab to trailer gap, representing a rigid truck and an articulated vehicle
respectively (see Fig. 14.56(a and b)). These graphs
illustrate the general trend and do not take into
account the different cab to trailer heights, cab to
trailer air gap width and the width to height ratio of
the deflector plate. It can be seen that with a
rigid truck having a small cab to trailer gap the
(a) Section view
Fig. 14.55 (a and b)
(b) Pictorial view
Moulded adjustable cab roof deflector
l = 1.64 m
x = 0.82 m
Drag coefficient (CD)
x/l = 0.5
(a) Rigid truck
l = 2.66 m
x = 0.8 m
x/l = 0.3
Deflector inclination angle() deg
(b) Articulated truck
Fig. 14.56 (a±c)
Optimizing roof deflector effectiveness for both rigid and articulated trucks
reduction in the drag coefficient with increased
deflector plate inclination is gradual, reaching an
optimum minimum at an inclination angle of 80
and then commencing to rise again, see Fig. 14.56(c).
With the articulated vehicle having a large cab to
trailer gap, the deflector plate effectiveness
increases rapidly with an increase in the deflector
inclination angle until the optimum angle of 50 is
reached, see Fig. 14.56(c). Beyond this angle the
drag coefficient begins to rise steadily again with
further increase in the deflector plate angle; this
indicates with the large gap of the articulated vehicle
the change in drag coefficient is much more
sensitive to deflector plate inclination.
14.8.2 Yaw angle (Figs 14.57 and 14.58)
With cars the influence of crosswinds on the drag
coefficient is relatively small; however, with much
larger vehicles a crosswind considerably raises the
drag coefficient therefore not only does the direct
air flow from the front but also the air movement
from the side has to be considered. It is therefore
necessary to study the effects crosswinds have on
the vehicle's drag resistance, taking into account the
velocity and angle of attack of the crosswind relative to the direction of motion of the vehicle and its
road speed. This is achieved by drawing to scale a
velocity vector triangle, see Fig. 14.57. The vehicle
velocity vector line is drawn, then the crosswind
Vehicle velocity
angle relative
to direction of
(yaw angle)
Fig. 14.57 The yaw angle
Relative flow
air velocity
Wind angle
relative to direction
of motion
Wind direction
velocity vector at the crosswind angle to the direction of motion; a third line representing the relative
air velocity then closes the triangle. The resultant
angle made between the direction the vehicle is
travelling and the resultant relative velocity is
known as the yaw angle, and it is this angle which
is used when investigating the effect of a crosswind
on the drag coefficient.
In addition to head and tail winds vehicles are also
subjected to crosswinds; crosswinds nearly always
raise the drag coefficient, this being far more
pronounced as the vehicle size becomes larger and
the yaw angle (relative wind angle) is increased. The
effect crosswinds have on the drag coefficient for
various classes of vehicles expressed in terms of
the yaw angle (relative wind angle) is shown in
Fig. 14.58. Each class of vehicle with its own head
on (zero yaw angle) air flow drag coefficient is given
a drag coefficient of unity. It can be seen using a
drag coefficient of 1.0 with zero yaw angle (no wind)
that the drag coefficient for a car reaches a peak of
1.08 at a yaw angle of 20 , whereas for the van,
coach, articulated vehicle and rigid truck and trailer
the drag coefficient rose to 1.18, 1.35, 1.5 and 1.7
respectively for a similar yaw angle of 20 .
14.8.3 Cab roof deflector effectiveness versus
yaw angle (Fig. 14.59(a and b))
The benefits of reducing the drag coefficient with a
cab roof deflector are to some extent cancelled out
when the vehicle is subjected to crosswinds. This
can be demonstrated by studying data taken from
Rigid truck and
Articulated truck
Drag coefficient (CD)
Yaw angle(ψ) deg
Fig. 14.58 Influence of yaw angle upon aerodynamic drag
curve is now below that of the 10 yaw angle curve.
Note the minimum drag coefficient deflection inclination angle is only relevant for the dimensions of
this particular cab to trailer combination.
one particular vehicle, see Fig. 14.59(a and b), which
utilizes a cab roof deflector; here with zero yaw
angle the drag coefficient reduc