Application Manual for System Design and Fan Installation

Application Manual for System Design and Fan Installation
Vane Axial
Application and Design
May
®
2010
Table of Contents
Introduction . . . . . . . . . . . . . . . . . . .
Definitions . . . . . . . . . . . . . . . . . . .
Vane Axial Fan Arrangements
• Arrangement 4 Direct Drive . . . . . . . . . . . . .
• Arrangement 9 Belt Drive . . . . . . . . . . . . .
Variations of Vane Axial Construction
• Hub-to-Tip Ratio . . . . . . . . . . . . . . . .
• Half-Blade Fans . . . . . . . . . . . . . . . .
• Two-Stage Fans . . . . . . . . . . . . . . . .
• Fans in Parallel . . . . . . . . . . . . . . . . .
Factors Affecting Air Performance
• System Effect . . . . . . . . . . . . . . . . .
• Air Density . . . . . . . . . . . . . . . . . .
Vane Axial Accessories Affecting Performance
• Inlet Bell . . . . . . . . . . . . . . . . . . .
• Inlet Cone . . . . . . . . . . . . . . . . . .
• Outlet Cone . . . . . . . . . . . . . . . . . .
Understanding Direct Drive Performance Charts . . . . . .
• The Total Pressure Concept. . . . . . . . . . . . .
How Outlet Conditions Affect Total, Static and Velocity Pressure .
• Diagram of Pressure Variations for Various Outlet Conditions . .
Making Fan Selections
• Operating Stability . . . . . . . . . . . . . . . .
• Avoiding Vane Axial Stall . . . . . . . . . . . . . .
• Avoiding Motor Overload . . . . . . . . . . . . .
• Vane Axial Efficiency . . . . . . . . . . . . . . .
Vane Axial Fans in Variable Air Volume Systems . . . . . . .
Methods of Providing Variable Air Volume
• Two-Speed Motors . . . . . . . . . . . . . . .
• Variable Pitch Sheaves . . . . . . . . . . . . . .
• Inlet Vane Dampers . . . . . . . . . . . . . . .
• Outlet Volume Dampers . . . . . . . . . . . . . .
• Variable Frequency Drives . . . . . . . . . . . . .
Vane Axial Sound and Methods of Attenuation
• Greenheck’s Sound Trap Vane Axial . . . . . . . . . .
• Inlet and Outlet Sound Attenuators . . . . . . . . . .
• Acoustical Diffuser Cones . . . . . . . . . . . . .
• Sound Absorbing Materials . . . . . . . . . . . . .
• Fan Speed and Vane Axial Sound . . . . . . . . . . .
• Vibration Isolators . . . . . . . . . . . . . . . .
• Flexible Duct Connections . . . . . . . . . . . . .
• Thrust Restraints . . . . . . . . . . . . . . . .
Economic Considerations of Vane Axial Selection and Application.
Maintenance Costs . . . . . . . . . . . . . . . .
Specifications . . . . . . . . . . . . . . . . . .
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Backcover
2
®
INTRODUCTION
This manual provides information on the application of vane axial fans in variable or constant air volume systems.
Many problems encountered with air moving devices such as the vane axial fan, are a result of misapplication
due to lack of easy to read, comprehensive and understandable information. Greenheck makes every effort to
provide the customer with extensive product information. Due to the relatively high volumes, pressures and
velocities generated by vane axial fans and the potential for significant performance variations, this application
manual offers information for proper selection, installation and use.
DEFINITIONS
Adjustable Pitch - Vane axial rotor blades may be manually adjusted to various pitches. Fan must be off,
electrical power locked-out, blade retaining nuts loosened, and blades manually set to desired pitch (within
horsepower limitations).
Hub - The center of the rotor. Hubs contain a provision for attachment to the driven shaft and machined sockets
or holes for attaching the blades. The hub is usually covered by a nose-cone (a spun aluminum cover for
streamlining the hub).
Rotor - A term used to describe the vane axial propeller. The rotor consists of a hub and blades.
Static Regain - Conversion of the energy of motion (kinetic energy) or velocity pressure to potential energy or
usable static pressure. An example is the increase in static pressure as velocity is reduced across an outlet cone.
Swirl (Vortex) - Airflow rotating perpendicular to the intended axis of airflow. It is a swirling movement of air
generated by the vane axial rotor.
System Effect - A pressure loss resulting from fan inlet or outlet restrictions or other condition within the system
affecting fan performance. System effect is difficult to quantify and results in poor efficiency, noise and vibration.
Vane Axial Fan - An air moving device with axial airflow and straightening vanes to reduce swirl created by the
rotor.
Variable Frequency Drive (VFD) - A system for controlling the rotational speed of an AC motor. Traditionally
used on direct drive fans for changing the rotor speed and performance of the fan (may also be used on belt
drive fans).
VANE AXIAL FAN ARRANGEMENTS
Arrangement 4 Direct Drive
Arrangement 4 direct drive vane axial fans have the rotor attached directly to the motor. This arrangement has
several advantages over a belt drive unit in that it is more compact, has no drive losses reducing efficiency, and
requires relatively little maintenance. The disadvantages include fan speeds limited to the motor speed (if used
without a variable frequency drive (VFD)), poor motor accessibility, and maximum airstream temperature of 105°F
using standard motor insulation. Arrangement 4 direct drive fans are available with adjustable pitch rotors and
the sound trap option.
Arrangement 9 Belt Drive
Arrangement 9 belt drive fans are constructed with the motor mounted on the fan housing, out of the airstream.
The rotor is attached to a fan shaft supported by grease lubricated bearings. A belt tube provides passage of the
belts from the motor to the driven pulley. Belt drive advantages include the wide range of fan speeds available,
tolerance of airstream temperatures up to 200°F, and easy motor accessibility. Also, motors for belt drive units
are generally lower cost and more readily available than those in direct drive vane axials. Arrangement 9 belt
drive fans are available with adjustable pitch rotors and the sound trap option.
3
®
VARIATIONS OF VANE AXIAL FAN CONSTRUCTION
Hub-to-Tip Ratio
Two-Stage
Static Pressure (Ps)
The hub-to-tip ratio
of a fan is the ratio
of the hub diameter
to the blade tip
diameter. Varying this
ratio will change the
fan's performance
capabilities. Rotors
with higher hub to tip
ratios will generate
higher static pressures.
Rotors with lower
hub-to-tip ratios will
generate less static
pressure. Selecting the
correct hub-to-tip ratio
for a given size fan can
optimize fan efficiency
and reduce chances
of the fan stalling in
the field if the system
resistance increases.
See page 12 for further
details on making fan
selections.
High Hub-Tip Ratio
Mid Hub-Tip Ratio
Low Hub-Tip Ratio
Half-Blade
Hub-to-Tip Ratio
Curves
Large H/T Æ High Ps
CFM
Small H/T Æ Low Ps
Half-Blade Vane Axial Fans - Direct Drive
Removing every other blade from the rotor has some definite advantages in low pressure selections. Vane axial
fans with a half-blade rotor will require approximately 65 percent of the horsepower required for a full blade rotor,
yet will deliver the same volume (cfm). The downside of a half-blade rotor is that it will generate approximately
65 percent of the pressure of a full blade rotor. Therefore, when the operating point falls low on the vane axial fan
curve and the application is for relatively low static pressures, a half-blade rotor should be considered to reduce
brake horsepower and increase efficiency.
Half-bladed fan selections are available for the smallest hub size for each direct drive fan size in order to extend
the useful pressure range as shown above.
4
®
Two-Stage Vane Axial Fans - Direct Drive
Where very high static pressures are required, the two-stage vane axial should be considered. Two-stage vane
axial fans have two rotors, one at each end of
the motor. These fans will generate twice the
amount of pressure, require twice the brake
horsepower and will deliver the same volume
as a single stage unit. A second vane section
is used on the exhaust end of the fan to
reduce the swirl from the second stage rotor.
Diagrams showing single- and two-stage
Two-Stage
Single-Stage
vane axial fans are shown.
Fans in Parallel
There are times when one fan may be too large and not fit into a desired space or the required operating range
of a system may necessitate multiple fans instead of one large fan. For these applications it is common to use
multiple fans in parallel. Multiple fans for capacity control may be more economical if cost of operation is critical,
especially at very low flow rates for long time intervals.
Static Pressure (Ps)
For multiple fans in parallel, each
fan will be selected for the same
static or total pressure with the
flow rate being the total flow
Single fan
divided by the number of fans.
performance curve
Use care when selecting fans in
Do not make selections
parallel to ensure that the system
above this line.
resistance remains on a stable
portion of the fan curve at all times.
Standard surge
Parallel fan
line for single
performance curve
This is particularly true when the
fan operation.
Single fan surge
fans have a pronounced surge
line for parallel fan
applications.
area or a dip in the fan curve and
some form of control is applied.
Parallel surge line.
The operating point with all fans
running must be lower than
CFM
the lowest pressure in the dip.
This minimizes the possibility that the fan will hunt back and forth across the peak of the curve looking for an
operating point. This policy also minimizes the likelihood that the fans will experience unequal loading causing
differences in motor load or creating unequal velocity profiles if used within a plenum, which may result in a
system effect.
For fans in parallel be sure to keep adequate distance
between fans and walls to ensure proper intake conditions.
See diagram for general spacing guidelines.
1.5 D
Airflow
2D
Airflow
1.5 D
5
®
FACTORS AFFECTING AIR PERFORMANCE
System Effect
Imagine a vane axial fan selected with great care to provide exactly the performance required in the
specifications. Once installed, the air balancer reports that air performance is considerably lower than required.
What went wrong?
The answer is probably system effect. The Air Movement and Control Association International Inc. (AMCA)
defines system effect as "a pressure loss which recognized the effect of fan inlet restrictions, fan outlet
restrictions, or other conditions influencing fan performance when installed in the system."
Fan manufacturers go to great lengths to test fans and provide reliable air performance data in their literature.
These fans are tested under very specific conditions as specified on the performance pages. Statements such
as, "Performance shown is for model 'xyz' with inlet and outlet ducts," indicate how the fan was tested. An
installation where elbows, transitions, dampers and other disruptions to airflow are located before or after the fan
can create a condition different from the manufacturer's test methods. Therefore, a performance loss or system
effect is created.
System effect is very difficult to quantify and correct. Frequently, the only means to correct the resulting poor
performance is to
increase fan speed
or increase the
Three Fan Diameters
blade pitch. Both of
these situations may
increase horsepower
Turning
Vanes
requirements that
exceed the capability
Good
Poor
of the motors. Also, the
system effect may be so
D
great that the fan is not
capable of generating
enough static pressure
Good
Poor
even at maximum fan
Ducted Inlet Conditions
speed. This could mean
replacing the fan with
one of greater capacity.
Finally, system effect
One
Fan
will rob an air moving
Diameter
device of efficiency.
Inlet Bell
Higher fan speeds and
Good
Poor
greater horsepower
used to overcome a
Non-Ducted Inlet Conditions
design deficiency result
in wasted energy.
The diagrams show
some of the more
common causes of
system effect. Nonuniform airflow created
by duct elbows,
transitions, dampers
or other obstacles
in the airstream may
dramatically reduce
fan performance. Refer
to AMCA Publication
201 for a quantitative
discussion of system
effects.
Length of Straight Duct
Minimum of three fan diameters
Good
Poor
Ducted Outlet Conditions
Two Fan
Diameter
Outlet Cone
Good
Poor
Non-Ducted Outlet Conditions
6
®
Air Density
Air density is a function of elevation and temperature, and both variables affect fan air performance. Air density
will affect the total pressure that a fan can generate and the horsepower required to move the air.
Most fan performance is published at a density based on air at 70°F and at sea level. This is referred to as
“standard air.” The resulting density is 0.075 Ibs. per cubic foot.
A fan operating at a higher elevation or temperature will move the SAME VOLUME of air as it will at standard
conditions, however, it will generate LESS TOTAL PRESSURE and will require LESS HORSEPOWER.
When selecting a vane axial fan to operate at a non-standard density using standard air density tables and
curves, corrections must be made to the parameters affected by air density. These parameters are static
pressure and brake horsepower.
At higher than standard elevations and temperatures, air density will be lower than standard. Therefore, we
must determine the static pressure at standard air density that will equate to the specified static pressure at our
operating density. Since standard air density is greater than operating air density in this case, we would expect
the corrected static pressure to be greater than the operating static pressure.
The following table provides air density correction factors for non-standard temperatures and elevations.
The example shows
how to select belt drive
model VAB-30F14 with a
ducted outlet, elevation
of 8,000 feet and
temperature of 100°F.
Performance required
is 22,000 cfm at 2.11
inches static pressure.
DRY AIR DENSITY CORRECTION FACTOR (I-P)
Multiply Standard Air Density, 0.075 lbm / ft3 by the Factor to obtain Density at Condition pb
Altitude, (Z)
Barometric Pressure
( pb )
1. Air volume delivered
by the fan is not
affected by density
and remains at
22,000 cfm.
2. Using the Dry Air
Density Correction
Factor (I-P) table, the
correction factor for
100°F and 8,000 ft.
is 0.703.
Temperature °F, (t)
ft.
-1000
Sea
Level
1000
2000
3000
4000
5000
6000
7000
8000
9000
10000
in. Hg
31.02
29.92
28.85
27.82
26.82
25.84
24.89
23.98
23.09
22.22
21.39
20.58
in. wg 421.71 406.75 392.21 378.20 364.61 351.29 338.37 326.00 313.90 302.07 290.79 279.78
-40
1.309
1.262
1.217
1.174
1.131
1.090
1.050
1.012
0.974
0.937
0.902
0.868
0
1.195
1.152
1.111
1.071
1.033
0.995
0.959
0.924
0.889
0.856
0.824
0.793
40
1.099
1.060
1.022
0.986
0.950
0.915
0.882
0.850
0.818
0.787
0.758
0.729
70
1.037
1.000
0.964
0.930
0.896
0.864
0.832
0.801
0.772
0.743
0.715
0.688
100
0.981
0.946
0.913
0.880
0.848
0.817
0.787
0.759
0.730
0.703
0.677
0.651
150
0.901
0.869
0.838
0.808
0.779
0.750
0.723
0.696
0.670
0.645
0.621
0.598
200
0.832
0.803
0.774
0.747
0.720
0.693
0.668
0.644
0.620
0.596
0.574
0.552
250
0.774
0.746
0.720
0.694
0.669
0.645
0.621
0.598
0.576
0.554
0.534
0.513
300
0.723
0.697
0.672
0.648
0.625
0.602
0.580
0.559
0.538
0.518
0.498
0.480
350
0.678
0.654
0.631
0.608
0.586
0.565
0.544
0.524
0.505
0.486
0.468
0.450
400
0.639
0.616
0.594
0.573
0.552
0.532
0.513
0.494
0.475
0.458
0.440
0.424
450
0.604
0.582
0.561
0.541
0.522
0.503
0.484
0.467
0.449
0.432
0.416
0.401
500
0.572
0.552
0.532
0.513
0.495
0.477
0.459
0.442
0.426
0.410
0.395
0.380
550
0.544
0.525
0.506
0.488
0.470
0.453
0.436
0.420
0.405
0.390
0.375
0.361
600
0.518
0.500
0.482
0.465
0.448
0.432
0.416
0.401
0.386
0.371
0.357
0.344
700
0.474
0.457
0.440
0.425
0.409
0.394
0.380
0.366
0.352
0.339
0.327
0.314
3. Divide the static
800 0.436 0.420 0.405 0.391 0.377 0.363 0.350 0.337 0.324 0.312 0.301 0.289
900 0.404 0.390 0.376 0.362 0.349 0.336 0.324 0.312 0.301 0.289 0.278 0.268
pressure (2.11) by the
1000
0.376 0.363 0.350 0.337 0.325 0.313 0.302 0.291 0.280 0.269 0.259 0.250
altitude/temperature
correction factor
Adapted from AMCA Standard 99-09, section 0200, Charts and Tables, with written permission from Air Movement and
Control Association International, Inc.
(0.703) to find the
standard air density
equivalent static pressure. 2.11 inches ÷ 0.703 = 3.0 inches.
4. Enter the performance chart for VAB-30F14 in Greenheck's Vane Axial Fan Performance Supplement using
22,000 cfm and 3.0 inches corrected static pressure. The fan RPM at the operating point is 2156 and the Bhp
is 20.9.
5. Since the horsepower selected is based on standard air density, it must be corrected to reflect Bhp at the less
dense conditions. Therefore, multiply the Bhp (20.9) by the altitude/temperature correction factor (0.703). The
new Bhp is 14.7 and a 15 horsepower motor can be selected.
An important point to remember is if a fan is selected to operate at high temperatures, the motor must be of
sufficient horsepower to handle the increased load at any lower operating temperature where the air is more
dense. For example, if the start-up temperature for the VAB-30F14 was 0°F, the correction factor would be 0.856
and the Bhp would be 17.9 (20.9 x 0.856 = 17.9). A 20 horsepower motor is now required.
7
®
VANE AXIAL ACCESSORIES AFFECTING AIR PERFORMANCE
Inlet Bell
Uniform airflow into the vane axial rotor is a prerequisite of cataloged
performance. A non-ducted vane axial inlet, without the aid of an inlet
bell, does not provide for the smooth airflow required. High velocity
airflow drawn over an abrupt edge of the vane axial housing creates a
phenomenon known as “vena-contracta.” In other words, the airflow is
diverted away from the walls of the fan housing, slightly reducing the
effective inlet area and creating little or no loading of the rotor blade
tips. This results in poor performance, vibration and excessive noise. An
inlet bell streamlines the housing of a non-ducted vane axial fan assuring
cataloged performance.
Example of “vena-contracta”
with no inlet bell
Inlet Cone
Inlet cones accomplish the same purpose as inlet bells, but for ducted inlet applications. The inlet cone is an
excellent transition from a larger duct diameter to a smaller vane axial housing. It allows for a smooth gradual
reduction in duct size and uniform airflow into the vane axial rotor. Converging angle should be limited to
30° per side.
Outlet Cone
Outlet cones, also known as diffuser sections, improve vane axial efficiency by providing for static regain.
Vane axial fans are typically high volume air moving devices with relatively high outlet velocities. These high
outlet velocities and their resultant velocity pressure losses, rob a vane axial of a significant portion of its total
efficiency. An outlet cone installed on the discharge end of a vane axial fan allows high velocity airflow to expand
gradually, converting much of the velocity pressure to usable static pressure. This savings of energy is apparent
when two vane axials with high outlet velocities are selected, one without an outlet cone and one with an outlet
cone. The fan selected with the outlet cone will require considerably less horsepower to deliver the same CFM
and static pressure. An important point to remember is that an outlet cone must discharge into a duct diameter
the same size as the large end of the outlet cone and this duct must not be reduced in size immediately after the
cone. Diverging angle should be limited to 15° per side to prevent expansion losses.
Pages 10 through 13 offer a detailed discussion of total, static and velocity pressure effects on vane axial
performance.
Best - Ducted Outlet with Outlet Cone
Best - Free Discharge with Outlet Cone
Good - Ducted Outlet
Poor - Free Discharge
Vane Axial Discharge Configurations
8
®
UNDERSTANDING DIRECT DRIVE VANE AXIAL PERFORMANCE CHARTS
Because of overlapping performance, numerous variables affecting fan performance, and a requirement to
select vane axial fans using total pressure instead of static pressure, traditional fan selection charts using static
pressure and CFM are unsuitable. Instead, direct drive vane axial fans are selected from charts containing fan
curves for several blade pitch settings.
In addition to fan curves, each chart contains variables for CFM, velocity pressure, total pressure, brake
horsepower, outlet velocity and efficiency.
The Total Pressure Concept
Total pressure is the sum of velocity pressure and static pressure (Pt = Pv + Ps ).
This fundamental equation is true at any location in a fan and duct system.
Velocity pressure is created by movement of the air through a fan, duct or similar device. The higher the velocity
of airflow, the greater the velocity pressure. Velocity pressure cannot be measured directly. It is the difference
between total and static pressure (Pv = Pt - Ps ).
Velocity pressure is always a positive value whereas static pressure is measured in relation to the surrounding
atmosphere and may be a positive or negative value. The actual values of Pt , Pv , and Ps change throughout a
system, so it is important to define the location of the pressure reading. See diagram below.
Pt
Ps
Pv
Flow
Basis for measurement of Total Pressure (Pt ), Static Pressure (Ps ) and Velocity Pressure (Pv).
The following formula is used to calculate velocity pressure:
Pv =
U(
V
1096
)
2
Pv = Velocity Pressure (inches H2O)
U = Air Density (lb./feet3)
V = Air Velocity (feet/minute)
Air velocity and the corresponding velocity pressure, changes with the fan or duct area.
9
®
HOW OUTLET CONDITIONS AFFECT TOTAL, STATIC, AND VELOCITY PRESSURE
The purpose of a fan in an air movement system is to increase the total pressure of the system. This total
pressure rise takes place primarily in the fan propeller, but additionally in the straightening vane section as
swirling air velocity is converted to static pressure. The total pressure rise in the fan is used to overcome system
resistance losses as well as losses due to outlet conditions.
The outlet losses change with different outlet configurations and are illustrated on the opposing page. When
looking at the fluctuations in pressures through the various outlet conditions, it is important to remember that
the total pressure is always equal to the sum of the static and velocity pressures. The curves show typical
performance only and are used to show changes in the pressure values. If a free inlet is required, the inlet duct
can be replaced with an inlet bell which provides a smooth transition to the fan velocity with no loss in total
pressure.
The Ducted Outlet configuration results in the highest value of total pressure rise. The constant area of the duct
connection eliminates any expansion losses. Expansion losses occur whenever air is forced to expand and slow
down into a larger area. Some expansion losses are present within the fan, as the air expands around the back
side of the motor to fill the entire fan area. This loss is ignored in these examples for simplicity, but is included in
published performance values.
The slight drop in total pressure through the outlet cone in the Ducted with Cone situation is an expansion loss
due to the increasing area of the cone. This loss is kept to a minimum by using a cone with a diverging angle of
15 degrees. The advantage of using an outlet cone, however, can be seen in the conversion of velocity pressure
to usable static pressure. This is referred to as static regain and the result is lower fan Bhp for a given static
pressure. An additional benefit is reduced duct resistance due to the lower duct velocity.
When the air undergoes a large expansion, as in a Free Discharge into a plenum or the atmosphere, the
expansion loss becomes significant. In this case, velocity pressure is reduced to approximately zero, since
the area is very large. However, none of this velocity pressure is converted to static pressure, in fact, the static
pressure also drops since the expansion is so sudden. Total pressure is equal to static pressure at the discharge,
since velocity pressure drops to zero.
The addition of an outlet cone in a free discharge, Free with Cone, is beneficial for two reasons. First, some
static regain takes place in the outlet cone, similar to the ducted with cone situation. Secondly, since the air has
already expanded and slowed down through the outlet cone, the expansion loss at the cone’s discharge is not as
high as it would be at the fan outlet. Therefore, due to its dramatic effect on outlet static pressure, an outlet cone
should be used whenever possible with a free discharge.
10
®
Outlet
Duct
Straightening
Vanes
Propeller
Inlet
Duct
Outlet Cone
PC
Pt
2
Ps
1
Ducted
Outlet
Pv
0
-1
PC
Pt
Ps
2
1
Ducted
with Cone
Pv
0
-1
PC
2
Ps = Pt
1
Free
Discharge
0
Pv
-1
PC
2
Ps = Pt
1
Free
with Cone
0
Pv
-1
Pressure variations for various outlet conditions
11
®
MAKING FAN SELECTIONS
Greenheck vane axial fans were tested using an inlet bell to simulate a ducted inlet and an outlet duct with a
diameter equal to the fan diameter. The curves are plotted in total pressure for various blade pitches and RPM’s.
In order to convert static pressure at any of the four outlet conditions to the published total pressure with ducted
outlet, a pressure correction (Pc) is used. Pc is defined as the difference between ducted total pressure and static
pressure in the duct or plenum. The pressure correction includes any static regain through outlet cones, as well
as expansion losses in outlet cones and in free discharges. Pc should not be confused with velocity pressure,
since they are equal only in the case of a ducted outlet. Values of Pc can be determined from Outlet Condition
Corrections charts on each page of the Vane Axial Fan Performance Supplement.
It is important to consider outlet conditions when selecting a Greenheck vane axial fan to ensure the required
performance will be met when installed. It is also important to consider outlet conditions when comparing the
performance of different manufacturers. Some manufacturers publish a “stage” total pressure, or a total pressure
increase from the inlet of the fan to a point in the straightening vanes just prior to the expansion into the center
portion of the fan. By considering this expansion loss as part of their outlet condition correction, they are able to
publish a higher total pressure (and total efficiency) than is physically possible, even with a ducted outlet. Since
stage total pressure is not the same as fan total pressure, static pressure in the duct or plenum should be used
as a means of fan selection when making comparisons between manufacturers.
Operating Stability
A vane axial fan operating at a stable portion of the fan curve will be efficient, relatively quiet and produce
practically no vibration. System resistance may increase or decrease within reasonable limits and the fan will
continue to operate smoothly. This is an example of a system well designed and a fan selection made well below
the surge or stall point of the fan curve. Conversely, a fan selection which does not consider the possibility of an
increase from design system resistance may result in vane axial stall and/or motor overload.
1.75
250
200
150
1.50
1.25
Stall Area
1.0
RV
E
300
Static Pressure (in. wg)
350
Static Pressure (Pa)
Vane axial stall is a result of operation
beyond the fan’s capacity to generate
enough pressure for the volume of air
required. As total pressure increases,
air volume decreases to a point where
the rotor blades “stall out.” The result
is greatly reduced airflow due to flow
separation at the trailing edge of the
blades. Immediately after the stall,
system pressure drops, airflow begins
to increase, and the entire cycle starts
again. This is a very undesirable
condition known as surge. During
surge the fan is continually “hunting” or
cycling, trying to overcome excessive
system resistance. The result is
abnormal vibration, noise and stress on
fan components. If left unchecked, vane
axial stall may result in catastrophic
failure of one or more fan components.
.75
100
.50
50
.25
0
0.0
0
TEM
SYS
20
CE
AN
T
SIS
RE
40
60
CU
80
100
120
140
160
180
CFM X 1000
0
50
100
150
200
250
300
350
3
m /hr X 1000
Vane axial fan curve with stall area indicated
Avoiding Vane Axial Stall
Both system design and fan selection play an equally important role in prevention of stall. Systems designed with
little variation between design and actual resistance, and with minimal system effect, are most likely to produce
stable fan operation. The importance of accurate system resistance calculations cannot be stressed enough.
Stable vane axial performance is a result of selecting an operating point below the maximum pressure the fan
is capable of delivering. On a fan curve, that means avoiding the top area of the curve. A selection near the top
or stall area of the fan curve will not allow an unscheduled increase in total pressure and should reflect how
confident you are with the system resistance calculations. A selection near the center of the volume/pressure
curve is best and will usually provide the highest efficiency.
12
®
Avoiding Motor Overload
Specifying the correct motor horsepower is a very important part of selecting a vane axial fan. A motor selected
with no margin for error in design system static pressure may be short lived. Higher than anticipated system
static pressure may overload the motor and lead to eventual failure. Additionally, the correct horsepower must be
selected for a specific fan RPM. This is usually not a problem with direct drive vane axial fans but with belt drive
units an increase in fan RPM can easily result in motor overload. It is a good policy to consult the factory before
attempting to increase fan speed on any vane axial fan.
On direct drive vane axial fans, brake horsepower can usually be increased into the motor service factor. Airflow
over the motor creates a cooling effect and operation into the service factor, typically a factor of 1.15, is allowed.
For example, with adequate cooling a 10 hp motor is capable of delivering 11.5 Bhp with no reduction in service
life. However, a note of caution here – electrical supply wiring, circuit protection and switches must be capable
of handling the increased electrical load.
If a fan RPM increase is required, refer to the appropriate performance chart and select brake horsepower
required at the new operating point. After the speed change is made, always check motor load amperage and
compare it to the motor nameplate rating.
Vane Axial Efficiency
One reason for specifying a vane axial fan is its efficient operation. The straight-through airflow, minimal
discharge vortex and static regain available from a vane axial fan make it a very desirable air moving device.
Total efficiencies from 70 to 80 percent are not uncommon for fans with a good operating point.
However, oversights in the selection process can have a detrimental affect on fan efficiency. First, fan size has
an affect on efficiency. Selecting too small a fan will increase air velocity and create excessive pressure losses.
To overcome these losses a larger motor is necessary. Here we have two detrimental effects: high pressure
losses and inefficient unit size. Secondly, choosing the wrong hub-to-tip ratio will reduce efficiency. For relatively
large air volumes at low total pressures, try to select a small hub in relation to the fin tip diameter. Smaller hubs
mean longer rotor blades, which will move a greater volume of air at relatively low fan speeds. For performance
requirements with relatively low volumes at high total pressures, try to select a large hub with relation to the fin
tip diameter. A large hub with short rotor blades, turning at high fan RPM, will generate the high total pressures
required.
VANE AXIAL FANS IN VARIABLE AIR VOLUME SYSTEMS
Variable air volume (VAV) systems are quite common in the HVAC industry. Ventilation, heating and cooling
demands in a building will vary significantly in 24 hours because of occupancy and outside air temperature. It
makes economical sense to reduce the ventilation, heating or cooling at night when the building is not occupied.
Conversely, during peak business hours maximum HVAC usage can be expected. A variable air volume system
makes this all possible. Vane axial fans with variable frequency drives are highly suited to variable air volume
systems.
13
®
METHODS OF PROVIDING VARIABLE AIR VOLUME
Two-Speed Motors - Motors used in direct or belt drive vane axial fans are typically available with one-third or
one-half speed reduction. This in turn will provide a corresponding air volume reduction of one-third or one-half.
Variable Pitch Sheaves - Belt drive vane axial fans with variable pitch motor sheaves provide changeable air
volume, but only after stopping the fan and mechanically adjusting the sheave. In addition to being inconvenient,
adjustable motor sheaves should not be used on motors over 25 horsepower because of excessive weight and
possible imbalance.
Inlet Vane Dampers - Variable inlet vanes provide airflow modulation, but should not be used for vane axial
fans because of high pressures and velocities. Inlet vanes create a swirl into the fan inlet, which reduces airflow
and power requirements. A pressure drop results from airflow resistance through the inlet vane assembly and
reduces performance.
Outlet Volume Dampers - Outlet volume dampers are seldom used for variable air volume control with vane
axial fans. High pressures and velocities require use of heavy-duty industrial-type dampers which are costly
and inefficient. Using an outlet volume damper on a propeller type fan—such as the vane axial—can increase
horsepower requirements, overload the motor, and result in eventual motor burnout.
Variable Frequency Drives - Varying motor speed with a variable frequency drive provides a very efficient
means of modulating airflow. The motor speed is varied electrically and airflow is varied proportional to fan
speed. Advantages of using a variable frequency drive are reliable, precise microprocessor control of motor
speed and long-term energy savings resulting from fan operation at reduced horsepower levels. A preset pitch
fan with a variable frequency drive has no additional moving parts and theoretically should be very reliable.
As fan speed is lowered, brake horsepower is reduced by the cube of the speed ratio and results in significant
energy savings.
One disadvantage of varying fan speed is that fan performance follows a system resistance line on the fan
curve, meaning that pressure varies with the square of the fan speed ratio. In other words, if you want to keep
the system pressure and only vary the air volume, it cannot be done. While this is an efficient means of varying
air volume, it is not suited for all applications. Varying fan speed in order to vary air volume is not suitable for a
system which requires constant pressure.
One special note regarding variable frequency drives for use with explosion-proof motors: The variable frequency
drive and explosion-proof motor must be UL listed together as a complete system when they are used in a
hazardous environment.
VANE AXIAL SOUND AND METHODS OF ATTENUATION
Vane axial sound power levels in the upper octave bands are typically higher than other axial and centrifugal
fans. The following example (40,000 cfm and 2.0 inches static pressure) shows how sound power levels and
frequencies compare between a 60 inch vane axial and a 60 inch single width centrifugal fan.
Octave Band
1
2
3
4
5
6
7
8
Center Frequency (Hz)
63
125
250
500
1000
2000
4000
8000
Vane Axial (dB)
91
97
99
98
95
91
86
81
Centrifugal (dB)
96
91
86
83
78
74
70
65
Sound power comparison of a VAB-60F26 vane axial fan and a 60-BISW centrifugal fan
As shown above, it is apparent that a vane axial fan is louder in the higher frequency bands. Consequently, one
of the first steps in controlling vane axial sound is consideration of these high frequencies in system design. A
vane axial fan located in an equipment room directly above the chief executive officer’s suite would not be an
example of good system design.
Although vane axial fans generate high frequency sound, it is the easiest to attenuate. Reducing the high
frequencies of a vane axial fan is relatively simple compared to the low frequency rumble generated by large
centrifugal fans. A choice of many different methods of reducing vane axial sound is available to the system
designer. The following section discusses various methods of attenuation.
14
®
Greenheck’s Sound Trap Vane Axial
Mechanical sound attenuators have been used in the HVAC industry for many
years, and although very effective, are bulky and costly. Greenheck has taken the
mechanical sound attenuator one step further and incorporated it into the vane
axial fan housing.
Greenheck’s Sound Trap housing is unique in the air movement industry.
With floor space at premium prices in the building industry, eliminating costly
add-on silencers has a two-fold benefit. First, conventional acoustical silencers
add considerable length to the unit and are ineffective for reducing sound
radiated from the fan housing. Secondly, add-on silencers
reduce air performance by adding additional static pressure
to the system. In critical applications, this may increase brake
horsepower and result in a larger motor requirement.
Greenheck’s Sound Trap housing increases the fan length only
slightly to reduce sound transmitted from the inlet and outlet.
No additional pressure drop is experienced using the Sound
Trap housing. The Sound Trap housing is available on all
Greenheck vane axial fans, direct or belt drive.
Typical Sound Power Attenuation - Greenheck Sound Trap
construction effectively reduces inlet and outlet sound power
levels in each of the eight octave bands as shown below.
Octave Band
1
2
3
dB Reduction
0
-3
-7
4
5
-10 -10
6
7
8
-7
-7
-1
Inlet and Outlet Sound Attenuators
Sound attenuators placed at the vane axial inlet and outlet
provide reduction of sound emitted from each end of the fan, but are ineffective for sound radiated through the
fan housing. Additional sound attenuating material surrounding the fan would be necessary for adequate sound
control. Most inlet and outlet attenuators add considerable length to the fan plus create an additional pressure
drop which in turn increases the brake horsepower requirements.
One advantage of add-on type attenuators is the center cone found in many models. For the outlet end silencer,
the center cone fills the area of backwash turbulence directly behind the motor on direct drive fans or the bearing
tube on belt drive fans. In addition to reducing turbulence, these center cones also provide additional sound
absorption.
Acoustical Diffuser Cones
Sound attenuating diffuser cones are constructed similar to inlet/outlet sound attenuators described above, but
serve a two-fold purpose. First, they provide attenuation for sound transmitted out the ends of the fan housing.
Secondly, acoustical diffuser cones act as an outlet cone by reducing velocity pressure and allowing static regain
when installed at the exhaust end of a vane axial fan. The inner surface of the acoustical diffuser cone is conical
in shape, allowing high velocity airflow to expand before entering ductwork, plenums or free atmosphere.
When installed on the inlet of a vane axial fan, the acoustical diffuser cone serves as an acoustical transition,
allowing larger ductwork to be connected to a smaller fan. Some acoustical diffuser cones are also supplied with
center cones.
15
®
Sound Absorbing Materials
A variety of sound absorbing materials such as fiberglass, foam, perforated metal panels, quilted fiberglass, and
asphaltic mastic are available from manufacturers of noise control products. Some of these products can be
wrapped around the vane axial fan housing and associated ductwork to reduce radiated sound. Other methods
of application would be to line an equipment room or adjacent wall with acoustical foam tiles. Since the high
frequencies emitted by vane axial fans are easily attenuated, these materials can be very effective.
Fan Speed and Vane Axial Sound
When sound is an important consideration in vane axial application, fan speed must be analyzed carefully. For
a given duty it is much wiser to select a larger fan at lower RPM and low sound power levels than a small fan
at a high RPM. Small vane axial fans operating at high fan RPM’s present two problems. First, high rotor blade
tip speeds create excessive high frequency sound power levels. Secondly, high velocities created by a small
fan housing generate an abundance of air noise. Of course, high total pressures and operating stability may not
allow selection of a larger vane axial with lower fan RPM.
Vibration Isolators
Vane axial fans typically do not create a significant amount of vibration due to very demanding balance
tolerances. However, even small amounts of vibration transmitted through the building framework or ductwork
can be a problem. Left unchecked, low levels of fan vibration can be amplified by walls, floors and ducts. This
creates not only a noise problem but may also cause physical damage to the structure. Therefore, vibration
isolators are recommended for most vane axial installations. This includes horizontal or vertical floor mounting
and horizontal or vertical ceiling hung. All fans are subject to thrust loads resulting from differential pressure,
but vertical upblast installations present the unique problem of additional weight due to thrust. This must be
considered when sizing the vibration isolators. Use the formula below for calculating the force (F) in pounds due
to thrust.
F = 5.2 x Pt x OA
F = Thrust Force (lbs.)
Pt = Total Pressure Differential (inches H2O)
OA = Fan Outlet Area (ft²)
Formula for calculating thrust force (F) resulting from a differential pressure across fan.
Vibration isolators are available in several varieties; the most common are shown in the chart below.
Isolator Type
Free Standing Spring
Housed Spring
Restrained Spring
Application
Permits radial and axial vibration dampening.
Permits radial and axial vibration dampening where less motion
can be tolerated.
Used where large weight changes or high wind loads occur.
Upward vertical movement is prevented by mechanical restraints.
Seismic Control
Restricts movement of supported equipment during seismic
events (earthquakes) while providing isolation.
Spring Hanging
Provides vibration isolation of suspended equipment. Threaded
suspension rods typically supplied by installer.
Rubber-in-Shear
Neoprene isolators highly effective for relatively small fans with
RPMs of 1800 and over.
16
®
Flexible Duct Connections
Flexible duct connections should be considered for all vane axial fans with ducted inlets or outlets. Flexible
connectors prevent any residual fan vibration from being transmitted along the ductwork and being amplified
in the process. Metal duct is an excellent conductor of vibration and amplifies small vibrations into loud noises.
Flexible duct connectors absorb any vibration before it gets to the ductwork. Although flexible connectors are
very effective, two areas of caution must be addressed.
First, flexible duct connectors should be taut between the vane axial inlet and the connecting duct. High
velocities and pressures created immediately prior to the vane axial rotor tend to draw the flexible connector into
the airstream. A loose flexible connector will cause “necking” and will starve the rotor blade tips of air. Uneven
loading of the rotor blades creates poor fan performance and increases noise and vibration. Therefore, the
flexible connector should not be slack and should be just long enough for mechanical isolation.
Poor
Good
Thrust Restraints
The second recommendation for installing flexible duct connectors is use of thrust restraints if air thrust exceeds
10% of the fan weight (see page 16 for calculating thrust force). Pressure differentials create thrust in a direction
opposite to airflow and a force on the flexible duct connector if the fan is allowed to move. Thrust restraints limit
fan or duct movement while providing a spring loaded snubbing action. Thrust restraints are commonly used in
pairs, on opposite sides of the fan and flexible duct connection. See the diagram below.
Ductwork
Flex Duct
Connection
Fan
Flex Duct
Connection
Ductwork
17
®
ECONOMIC CONSIDERATIONS OF VANE AXIAL SELECTION AND APPLICATION
With limited construction budgets and ever increasing energy costs, selecting the most cost-effective fan can
be a dilemma. One is faced with either low initial cost for a fan with high energy cost or with higher initial cost
for a more energy-efficient fan. Selecting the most energy-efficient fan will generally provide a payback of the
additional fan price in a relative short time. How then, does one select the most energy-efficient vane axial fan?
Fan size plays an important role in vane axial efficiency. For a given performance using lowest brake horsepower
as a selection criteria on Greenheck’s Computer Aided Product Selection (CAPS) software, the most efficient
fan will usually be a larger size. Intuitively, we realize it takes less energy to move a given volume of air through
a 60 inch diameter tube than it does an 18 inch diameter tube. Velocities are lower, resistance is less and brake
horsepower required is less. The following CAPS selection for 80,000 cfm and 6 inches static pressure shows
how fan size affects brake horsepower.
Model
TP
OV
FRPM
Bhp
Mtr HP
TE
Pitch
$Yr*
1) VAD-54F30-14-A125
6.66
3184
1770
115
125
72
14
$45,091
2) VAD-48F30-27-A150
6.95
4095
1770
126
150
70
27
$49,404
*Operating costs estimated at $.09/kW-hr and 16 hours per day
CAPS selections showing how fan size affects brake horsepower
Hub-to-tip diameter ratio has an effect on vane axial efficiency. For large volumes at relatively low static pressure
the smallest diameter hub, providing longest blade length, will generally be most efficient. Conversely, for high
static pressures a large hub diameter with short blade lengths would be more efficient.
Fan speed is another factor affecting efficiency. Although direct drive vane axial fans require more costly motors
for low fan speeds, the long-term energy savings may compensate for the higher initial cost. The following CAPS
selections show how lower fan speeds translate into lower brake horsepowers and corresponding lower energy
cost.
Model
TP
OV
FRPM
Bhp
Mtr HP
TE
Pitch
$Yr*
1) VAD-54F26-21-B60
4.37
2388
1170
58.7
60
71
21
$23,016
2) VAD-54F26-5-A75
4.37
2388
1770
67.4
75
59
5
$26,427
*Operating costs estimated at $.09/kW-hr and 16 hours per day
Comparison of brake horsepower for model VAD-54F26 at 1170 and 1770 RPM
18
®
The chart below shows initial cost comparisons for the various direct drive motor speeds.
Horsepower
3500 RPM
1770 RPM
1170 RPM
870 RPM
25
.94
1.0
1.6
2.4
50
.91
1.0
1.9
3.2
75
.99
1.0
1.7
2.9
100
1.01
1.0
1.7
3.1
Cost comparison of direct drive motor speeds
The above cost factors were taken from 460 volt, 3 phase, open motors for a Model VAD vane axial.
The vane axial outlet cone is a true energy saver. Use of an outlet cone on a ducted or non-ducted vane axial
outlet will reduce velocity pressure, allow static regain and lower brake horsepower requirements. The long-term
savings from the reduced electrical load will provide payback for the outlet cone in minimal time.
Comparison of outdated
Fan Type
Centrifugal
Vane Axial
Vane Axial
response control fans
and adjustable pitch fans
Inlet Vane
Response Control
Adjustable Pitch with
Configuration
Control
(no longer available)
Variable Frequency Drive
with a variable frequency
drive deserves special
Percent of Full
70%
70%
70%
attention. The longLoad Airflow
term cost savings of
Percent of Power
variable frequency drives
65%
45%
30%
Consumption
is substantial because
modulating airflow by
Relative power consumption of fans in a VAV system
varying fan speed is very
efficient. A small reduction in fan speed results in a significant reduction in energy consumption. For example, an
adjustable pitch vane axial fan with a variable frequency drive operating at 70% of full load airflow will consume
only about 30% of full load Bhp. The chart shows relative power consumption with various means of providing
variable air volume.
MAINTENANCE COSTS
Scheduled preventive maintenance costs for the various vane axial models are relatively low. Belt drive vane
axial fans require periodic belt tensioning, bearing lubrication and general cleaning. Direct drive models require
only periodic motor bearing lubrication and cleaning.
Vane axial fans with a variable frequency drive require virtually no additional maintenance. In fact, using
variable frequency drives may reduce maintenance costs by eliminating power surges and reducing mechanical
stress during start-up. Electrical surges are all but eliminated because of reduced inrush current and longer
acceleration times. Wear and tear on belts, pulleys and bearings is greatly reduced by slow acceleration during
start-up.
19
®
Specifications
Belt Drive
Vane axial fans shall be belt driven,
Arrangement 9, with the motor attached to the
exterior of the fan housing on an adjustable
base. Turned, precision ground and polished
steel shafts shall be sized so the first critical
speed is at least 25% over the maximum
operating speed. Bearings shall be grease
lubricated, air handling quality ball or roller
type selected for a minimum average (L50)
life in excess of 200,000 hours at maximum
operating speed. Rotor blades and hub shall
be heat-treated cast aluminum alloy A356-T6
with blade bases and hub sockets precision
machined. Blades shall be attached to the hub
with steel studs and self-locking nuts. Hub
shall be positively secured with a steel taper
lock bushing keyed to the fan shaft. Rotor
blade pitch shall be manually adjustable within
horsepower limitations. Rotor shall be statically
and dynamically balanced to within 0.15 in./
sec. peak vibration velocity as measured on the
bearings. Fan housing shall be fabricated from
heavy-gauge steel with prepunched flanges at
both ends. A minimum of seven heavy-gauge
straightening vanes shall be welded to the
fan housing downstream from the rotor. (For
optional Sound Trap construction insert the
last paragraph). Vane axial fans shall be model
VAB or VABS (select one) as manufactured
by Greenheck Fan Corporation of Schofield,
Wisconsin, and shall be supplied as shown on
the plans and in the fan schedule.
Direct Drive
Vane axial fans shall be direct driven,
Arrangement 4, with the fan rotor secured to the
motor shaft. Motors shall be located downstream
from the rotor for maximum cooling. Rotor
blades and hub shall be heat-treated cast
aluminum alloy A356-T6 with blade bases and
hub sockets precision machined. Blades
shall be attached to the hub with steel
studs and self-locking nuts. Hub shall be
positively secured with a steel taper lock
bushing keyed to the motor shaft. (Add
paragraph for appropriate rotor type
here). Rotor shall be statically and
dynamically balanced to within 0.08 in./
sec. peak vibration velocity as measured
on the fan housing. Fan housing shall be
fabricated from heavy-gauge steel with
prepunched flanges at both ends. A
minimum of seven heavy-gauge
straightening vanes shall be welded to
the fan housing downstream from the
rotor. (For optional Sound Trap
construction insert paragraph below
here). Vane axial fans shall be model VAD or
VADS (select one) as manufactured by
Greenheck Fan Corporation of Schofield,
Wisconsin, and shall be supplied as shown on
the plans and in the fan schedule.
Adjustable Pitch Rotor
Blades shall be manually adjustable within
horsepower limitations. A blade tip angle scale
shall be machined into the base of the master
blade and indexed to the hub. All blades shall
be adjustable to align with the master blade
pitch setting.
Sound Trap Construction
When specifying optional Sound Trap
construction, add the following in the locations
noted above: Fan construction shall be doublewalled with two inches of sound absorbing
material between the walls. The inner wall
shall be constructed of perforated steel.
Air performance ratings shall be equal to
equivalent size fans with a single wall housing.
Our Commitment
As a result of our commitment to continuous improvement, Greenheck reserves the right to change specifications
without notice.
Specific Greenheck product warranties are located on greenheck.com within the product area
tabs and in the Library under Warranties.
Prepared to Support
Green Building Efforts
P.O. Box 410 • Schofield, WI 54476-0410 • Phone (715) 359-6171 • greenheck.com
Copyright © 2010 Greenheck Fan Corp. • 00.CVI.1025 R3 5-2010 RG
®
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