Fundamental Analysis of Thermal Overload in Diesel Engines

Fundamental Analysis of Thermal Overload in Diesel Engines
Nanda SK, Jia B, Smallbone AJ, Roskilly AP.
Fundamental Analysis of Thermal Overload in Diesel Engines:
Hypothesis and Validation.
Energies 2017, 10(3), 329.
Copyright:
© 2017 by the authors. Licensee MDPI, Basel, Switzerland. This article is an open access article
distributed under the terms and conditions of the Creative Commons Attribution (CC BY) license
( http://creativecommons.org/licenses/by/4.0/).
DOI link to article:
https://doi.org/10.3390/en10030329
Date deposited:
10/03/2017
This work is licensed under a Creative Commons Attribution 4.0 International License
Newcastle University ePrints - eprint.ncl.ac.uk
energies
Article
Fundamental Analysis of Thermal Overload in Diesel
Engines: Hypothesis and Validation
Sangram Kishore Nanda 1,2 , Boru Jia 2, *, Andrew Smallbone 2 and Anthony Paul Roskilly 2
1
2
*
Wärtsilä Services Switzerland Ltd., CH-8401 Winterthur, Switzerland; sknanda20@yahoo.co.uk
Sir Joseph Swan Centre for Energy Research, Newcastle University, Newcastle upon Tyne NE1 7RU, UK;
Andrew.Smallbone@newcastle.ac.uk (A.S.); tony.roskilly@newcastle.ac.uk (A.P.R.)
Correspondence: boru.jia@newcastle.ac.uk; Tel.: +44-0754-7839-154
Academic Editor: Wenming Yang
Received: 3 January 2017; Accepted: 2 March 2017; Published: 8 March 2017
Abstract: ‘Thermal Overload’ can be defined as a condition under which design threshold values
such as the surface temperature of combustion chamber components is exceeded. In this paper, a low
λ value is identified as the most probable cause of voluminous flame production, resulting in high
surface temperatures of engine components, i.e., engine thermal overload. Test results indicated that
the flame became voluminous when the excess air ratio, λ was low, and the exhaust temperature
increased from 775 to 1000 ◦ C with λ changing from 1.12 to 0.71. Temperature indicating paints
were applied on two piston crowns to investigate the effect of the voluminous flame on component
surface temperature. The piston crown with high rates of hot corrosion was very close to matt
glaze (much in excess of the design temperature), which proved that high surface temperature and
salt deposition on the crown in the heavily burned away regions could have been caused by flame
and fuel impingement, respectively. A numerical calculation was presented to estimate the flame
temperature for various air excess ratio, which provides a guidance for the operation conditions of
diesel engines to avoid engine thermal fatigue due to thermal overload.
Keywords: diesel engine; thermal overload; flame visualisation; validation
1. Introduction
In the last couple of decades, the power output from slow speed diesel engines has increased
steadily to meet the high propulsion power demand of large container vessels. This has motivated
engine makers to increase the power density (output per cylinder) up to 5800 kW (Wärtsilä–Sulzer
RTA96-C Common Rail marine compression-ignition engine) and reduce the specific weight from 55 to
30 kg/kW with the use of new materials and modern design tools [1]. This has resulted in more power
from the same bore size and number of cylinders for an engine family. Other key drivers for engine
development are low specific fuel oil consumption, which influences the direct operating costs of
a ship, and environmental legislation limiting the level of harmful pollutants from diesel engines [2–6].
Rising fuel prices and enforcement of the legislation limiting the NOx and SOx from marine diesel
engines have made these drivers more important [7–10]. A reduction in the fuel consumption can be
achieved by running the cylinder hotter, operating closer to stoichiometric conditions where cycle
temperatures are higher, and increasing the volume expansion ratio to extract more work for a fixed
heat input. This is achieved during the engine optimisation with near uniform air to fuel ratio across
the combustion chamber but at the expense of a smaller safety margin between normal operation and
thermal overload for continuous service rating.
‘Thermal Overload’ can be defined as a condition under which design threshold values such as
surface temperature of combustion chamber components is exceeded. The consequence is a reduction
in operating life of the component or catastrophic failure depending on the extent of overload.
Energies 2017, 10, 329; doi:10.3390/en10030329
www.mdpi.com/journal/energies
Energies 2017, 10, 329
2 of 12
Combustion chamber components such as the cylinder head, exhaust valve, cylinder liner and pistons
are made of materials that meet the necessary design objectives such as good tribological properties
for the cylinder liner and high hot corrosion resistance of exhaust valves on heavy fuel oil burning
engines. A predictive trend establishing the deterioration of engine health due to normal wear and tear
in components is acceptable provided corrective action is undertaken with preventive maintenance or
condition monitoring to minimise loss of performance. If the necessary preventive maintenance or
health monitoring of components taking part in the combustion or gas exchange process is not carried
out, it could lead to a thermal overload condition.
The type of material used and the rate of heat removal are two of the many other factors that limit
the operating temperature of a component. For a cylinder liner made of pearlitic grey cast iron with
good tribological properties, the surface temperature should be limited to 370 ◦ C to prevent ‘heat cracks’
from allotropic expansion of the material [11]. Therefore, to increase engine power, the design constraint
is taken care of in present day diesel engines with a deep cylinder cover, which means only a negligible
proportion of the cylinder liner is exposed to the combustion process. The cylinder cover is made
of cast steel which can withstand higher temperatures. The piston crown is made of forged chrome
molybdenum steel without any form of cladding and can sustain relatively high surface temperatures
but has very poor corrosion resistance to vanadate salts deposited on it as part of the combustion
process. It is one of two combustion chamber components that is cooled intermittently, in addition to
constant underside cooling, by cold scavenge air during the gas exchange process. If it is expected that
the surface temperature will be above the design threshold due to high heal flux then a cladding with
high corrosion resistance and low thermal conductivity needs to be applied. Another component that
is subjected to the most arduous conditions is the exhaust valve spindle that is cooled intermittently
by the scavenge air and its contact with the valve seat. It is made of Nimonic, which offers greater
resistance against hot corrosion.
To date, there has been research into the solutions to the thermal overload in the cylinder head of
a heavy duty 6-cylinder diesel engine, in which thermal cracks were found in the valve-bridge [12].
The temperature of the cylinder head bottom was measured, the flow distribution of coolant through
the upper nozzles of the cylinder head bottom was tested, and the water jacket of the cylinder head was
inquired. A 3D model was developed to simulate the water jacket performance in the worst cooling
conditions. Four designs of water jacket were proposed and simulated, and all these schemes showed
improvement in the flow field of the water jacket. From the test results, the maximum temperature in
the value bridge of the cylinder head was reduced by 9.2 ◦ C, and the temperature gradient reduction
was 19.55%, indicating that the proposed designs reduced the thermal stress of the diesel engine
cylinder head [12].
Engine pistons were regarded as an important part of an engine, and new geometries, materials,
and manufacturing techniques are always being proposed for pistons [13]. Notwithstanding these
studies, a huge number of damaged pistons were reported, among which thermal fatigue and
mechanical fatigue played an important role and was analysed by Silva. Stresses at the piston
crown and pin holes, as well as stresses at the grooves and skirt as a function of land clearances
were presented. Thermal fatigue cracks were easy to identify as some visible fatigue cracks on the
areas with cyclic thermal gradients could be seen on brake disks and other components. Two reasons
were considered through which thermal gradients act on stressed, i.e., vertical distribution of the
temperature along the piston, and temperature difference at the head of the piston due to the flow of
the hot air or fuel impingement [13].
An experimental and analytical approach was undertaken by Lee et al. to study stress distributions
and causes of failure in diesel cylinder heads [14]. The influence of thermal shock loading under rapid
transients was analysed, and the steady state temperature gradients and the level of temperatures
were reported as the primary causes of thermal fatigue in cast-iron cylinder heads [14]. They also
provided a finite element analysis to predict the detailed temperature and stress distributions within
the cylinder head, and validated the model using measurements. These works attributed thermal
Energies 2017, 10, 329
3 of 12
shock loading as playing a role in thermal fatigue, along with steady-state temperature gradients and
the level of temperatures [14].
In order to protect the structural engineering materials of engines from corrosion, erosion,
and wear, thermal barrier coating technologies have been proposed and investigated [15–19].
Different types of coating were used to provide engine lubrication and thermal insulation, as they could
insulate engine components from the hot gas stream, which improved the engine durability [16,17].
However, very little evidence has been published that considers the probable causes of the thermal
overload that will cause thermal fatigue from the engine side. None of them addressed the suggestions
of diesel engine operation conditions in order to avoid the engine thermal overload.
In this paper, the hypothesis for the most probable causes of thermal overload of diesel engines is
discussed and validated. Low λ value is identified as the most probable cause of voluminous flame
production, resulting in a high surface temperature of engine components, i.e., engine thermal overload.
After the probable cause of the engine thermal overload is identified, a numerical calculation will be
presented to estimate the flame temperature for various air excess ratio, which will provide guidance
for the operation conditions of diesel engines to avoid engine thermal fatigue due to thermal overload.
2. Fundamental Analysis
2.1. Diesel Engine Combustion
In a conventional diesel combustion process, fuel is injected into high-temperature air at the end of
the compression process followed by auto ignition and heat release. After an injection event, the process
can be identified on a pressure versus crank angle diagram as four distinct phases: ignition delay period,
rapid combustion phase, mixing controlled combustion phase and late combustion phase [20–22].
The duration and pressure profile for each of these phases depends on the quality of fuel, fuel injection
timing (fixed or variable), spray pattern, spatial distribution of fuel in the combustion space and
local air-to-fuel ratio. As the fuel is injected, its lower molecular weight fractions will be dragged by
the swirling air towards the periphery of the chamber, while the heavier fractions stay at the core.
The spray may be divided into four regions: lean flame region (LFR), lean flame out region (LFOR),
spray core and spray tail [23].
LFR is the concentration of vapour between the core and downstream edge of the spray and is
not homogeneous in 3D space. In this region, combustion is complete and nitrogen oxides (NO, NO2 )
are formed in higher local concentrations. Locally, LFOR is near the far downstream edge of the spray,
and the mixture is too lean to ignite or support continuous combustion. Following the ignition and
combustion in the LFR, the flame propagates towards the core of the spray. In this region, the fuel
droplets are larger. They gain heat by radiation from the already established flames and evaporate
at a higher rate. The spray tail is the last part of the spray to be injected and usually forms large
droplets due to relatively small pressure differential acting on the fuel near the end of the injection
process. This region is not responsible for NOx formation but is responsible for formation of unburned
hydrocarbons (UHC) and soot [23].
2.2. Hypothesis for the Most Probable Cause of Thermal Overload
The surface temperature of each component will be governed by conservation of energy i.e.,
the net heat input and output from the material. If the heat input is equal to the output, a steady state
surface temperature will ultimately result; however, an increase or decrease in either will result in
a change of the surface temperature of the material. It has been observed that when combustion takes
place in diffusion flames that there is an increase in flame size when the air has diluted concentrations
of oxygen. A blue flame (gas–gas phase reaction) is observed in the area close to the injection nozzle,
which, correspondingly, becomes larger as the oxygen concentrations are decreased. As a result,
a larger flame volume could result in greater flame/wall interaction and thus an increase in the wall
Energies 2017, 10, 329
4 of 12
temperatures. Then, the hypothesis for the most probable cause of thermal overload condition on
diesel engines is summarised as:
1.
2.
Low λ value is the most probable cause of voluminous flame production.
High2017,
surface
Energies
10, 329temperature of engine components due to a more voluminous flame.
4 of 12
2.To validate
High surface
temperature
components
due toexperiments
a more voluminous
flame.
the two
elementsofofengine
the hypothesis
above,
were undertaken
to investigate
the effectToofvalidate
λ value the
on flame
size and of
its the
temperature
validate
the hypothesis
and the to
effect of
two elements
hypothesistoabove,
experiments
were (1),
undertaken
voluminous
flame
on
component
surface
temperature
to
validate
the
hypothesis
(2).
investigate the effect of λ value on flame size and its temperature to validate the hypothesis (1), and
the effect of voluminous flame on component surface temperature to validate the hypothesis (2).
3. Hypothesis Validation and Discussion
3. Hypothesis Validation and Discussion
3.1. Effect of λ on Flame Size and Temperature
3.1. Effect of λ on Flame Size and Temperature
Combustion in this test rig was therefore used to simulate uniflow scavenged two-stroke diesel
Combustion inwith
this the
test rig
was therefore
used
to simulate
uniflow
scavenged
diesel
engine combustion,
trapped
air to fuel
ratio
indicating
conditions
in two-stroke
the primary
chamber
engine
combustion,
with
the
trapped
air
to
fuel
ratio
indicating
conditions
in
the
primary
chamber
and
and overall air to fuel ratio indicating conditions across the combustor. A flame visualisation
air to fuel ratio indicating conditions across the combustor. A flame visualisation test rig was
test overall
rig was
designed and manufactured using a pair of inter-connected combustion chambers of
designed and manufactured using a pair of inter-connected combustion chambers of 456 × 76 × 76 mm
456 × 76 × 76 mm and 228 × 170 × 76 mm (shown in Figure 1). The primary combustion chamber
and 228 × 170 × 76 mm (shown in Figure 1). The primary combustion chamber operated with a rich
operated
with
rich therefore,
fuel/air ratio
and,
therefore,
yielded a mixture
ofwith
combustion
products
with excess
fuel/air
ratioa and,
yielded
a mixture
of combustion
products
excess carbon
monoxide
carbon
monoxide
and
unburned
hydrocarbon.
The
secondary
combustion
chamber
was
designed
to
and unburned hydrocarbon. The secondary combustion chamber was designed to use the products
use of
thecombustion
products of
combustion
from
the primary
supplied
with
additional
from
the primary
chamber
and waschamber
supplied and
with was
additional
air to
burn
all of the air to
burncarbon
all ofmonoxide
the carbon
to carbon dioxide.
to monoxide
carbon dioxide.
Figure 1. Schematic diagram of the flame visualisation test rig.
Figure 1. Schematic diagram of the flame visualisation test rig.
Prior to a test, the combustion chambers were preheated with 200 °C air for a minimum period
ofPrior
30 min
until
thecombustion
exhaust gaschambers
temperature
reached
a steady
60200
°C. ◦A
is
to aortest,
the
were
preheated
with
C K-type
air for thermocouple
a minimum period
of
◦
usedortountil
measure
the exhaust
gas temperature.
Twoa exhaust
gas C.
sampling
points
were selected,
30 min
the exhaust
gas temperature
reached
steady 60
A K-type
thermocouple
is used to
between
primary
secondary combustion
chambers
and after
the secondary
combustion
measure
thethe
exhaust
gasand
temperature.
Two exhaust
gas sampling
points
were selected,
between the
chamber (as shown in Figure 1). A section of the exhaust sampling line was kept heated with a
primary and secondary combustion chambers and after the secondary combustion chamber (as shown
line heater. The sample line was divided into three branches, one branch leading to a UHC
in Figure 1). A section of the exhaust sampling line was kept heated with a line heater. The sample line
analyser, a second branch to a NOx analyser and the third to CO2, CO and O2 analysers via a gas
wasconditioning
divided intounit.
three
branches,
onethen
branch
leading
to athe
UHC
analyser,
a second
branch
to a NOx
The
ignitor was
switched
on and
master
fuel valve,
upstream
of the
analyser
and
the
third
to
CO
,
CO
and
O
analysers
via
a
gas
conditioning
unit.
The
ignitor
2 firing. On initiation
2
burner, opened to commence
of combustion, the ignitor was switched off and was
after a short overlap, the flame was allowed to stabilise for a period of 20 min. At the same time, all
other parameters were kept constant except for the fuel flow rate, which was reduced gradually to
Energies 2017, 10, 329
5 of 12
Overall λ
Exhaust temperature (deg C)
then switched on and the master fuel valve, upstream of the burner, opened to commence firing.
On initiation of combustion, the ignitor was switched off and after a short overlap, the flame was
allowed to stabilise for a period of 20 min. At the same time, all other parameters were kept constant
except for the fuel flow rate, which was reduced gradually to prevent soot formation. The test rig was
allowed to stabilise for an additional 15 min at each test point before any reading were recorded.
Energies 2017, 10, 329
5 of 12
Fuel rich conditions in the primary combustion chamber were achieved either by keeping the
prevent
soot formation.
The test rig the
was fuel
allowed
to rate
stabilise
for anversa.
additional
min at each
test
air flow rate
constant
while increasing
flow
or vice
The15former
provided
the best
point before any reading were recorded.
approach as the flow pattern through the combustion chamber remained constant while the fuel/air
Fuel rich conditions in the primary combustion chamber were achieved either by keeping the
ratio valueairwas
the course
of initial
testing
airprovided
flow rate
flowchanging.
rate constantDuring
while increasing
the fuel
flow rate
or vicewith
versa.constant
The former
the and
best increasing
fuel concentrations,
it
was
observed
that
the
flame
lift
off
distance
from
the
burner
increased
and the
approach as the flow pattern through the combustion chamber remained constant while the fuel/air
ratio
value
was
changing.
During
the
course
of
initial
testing
with
constant
air
flow
rate
and
observable flame became more voluminous. From the experimentation, carbon monoxide was found
increasing fuel concentrations, it was observed that the flame lift off distance from the burner
to form even
at slightly lean conditions (λ = 1.2). Tests were carried out to achieve fuel rich levels
increased and the observable flame became more voluminous. From the experimentation, carbon
down to λmonoxide
= 0.65, but,
at λ =to0.7,
critical
point,lean
theconditions
flame became
unstable
due
to a out
longer
was found
forma even
at slightly
(λ = 1.2).
Tests were
carried
to ignition
delay andachieve
lift offfuel
distance
from
the
This
in thepoint,
termination
of combustion
rich levels
down
to burner.
λ = 0.65, but,
at resulted
λ = 0.7, a critical
the flame became
unstable with any
due to a longer
andfinding,
lift off distance
from the
burner.out
Thisat
resulted
the termination
further reduction
in λ.ignition
Based delay
on this
tests were
carried
λ = 0.7inand
above.
of combustion with any further reduction in λ. Based on this finding, tests were carried out at λ = 0.7
Figure
1 can be used to explain the structure of the flame at different λ. The primary combustion
and above.
chamber was Figure
considered
represent
element
a dieselThe
engine
and
the location of the
1 can beto
used
to explainthe
the in-cylinder
structure of the
flame at of
differentλ.
primary
combustion
exhaust temperature
probe
at
the
outlet
as
the
position
of
any
critical
combustion
chamber
components,
chamber was considered to represent the in-cylinder element of a diesel engine and the location of
the
exhaust
temperature
probe
at
the
outlet
as
the
position
of
any
critical
combustion
chamber
such as the piston crown or cylinder liner. If the fuel rate is increased at constant air mass flow rate or
components, such as the piston crown or cylinder liner. If the fuel rate is increased at constant air
vice versa,
λ will decrease and the adiabatic flame temperature is supposed to decrease, as the total
mass flow rate or vice versa, λ will decrease and the adiabatic flame temperature is supposed to
energy from
the fuel
is not
due the
to only
oxidation
of only
the hydrocarbons.
as the
decrease,
as the
totalreleased
energy from
fuel ispartial
not released
due to
partial oxidation However,
of the
combustion
reaction slows
down,
flame becomes
voluminous.
The becomes
value ofvoluminous.
λ is calculated with
hydrocarbons.
However,
as thethe
combustion
reaction slows
down, the flame
The value
λ is fuel
calculated
air mass flow
andthis
fuelhas
rate. occurred
An indication
thatbe
this
occurred in exhaust
the air mass
flow ofand
rate.with
Anthe
indication
that
will
anhas
increase
will be an increase in exhaust temperature at the outlet probe as a result of closer proximity of the
temperature at the outlet probe as a result of closer proximity of the flame. As shown in Figure 2,
flame. As shown in Figure 2, the temperature at the probe
increased from 775 to 1000 °C with λ
◦ C with
the temperature
the probe
fromtemperature
775 to 1000
changing
1.12with
to 0.71.
changingatfrom
1.12 to increased
0.71. This high
flame
whenλclose
to or infrom
contact
any This high
temperature
flame
when
close
to
or
in
contact
with
any
combustion
chamber
component
will result in
combustion chamber component will result in an increased rate of heat transfer through the
component
and
its
consequent
failure.
an increased rate of heat transfer through the component and its consequent failure.
Figure
Flame position
indication.
Figure
2.2.Flame
position
indication.
3.2. Effect of Voluminous Flame on Component Surface Temperature
3.2. Effect of Voluminous Flame on Component Surface Temperature
To validate the hypothesis that a voluminous flame is the most probable cause of high surface
temperatures
on the pistonthat
crowns,
an experimentflame
was set
up on
two probable
piston crowns
with
To validate
the hypothesis
a voluminous
is the
most
cause
ofthe
high surface
application of temperature indicating paints. The thermal paint, named code MC277-7, with an
temperatures on the piston crowns, an experiment was set up on two piston crowns with the application
initial colour of yellow giving a quantitative estimate of peak piston crown surface temperatures.
of temperature indicating paints. The thermal paint, named code MC277-7, with an initial colour of
yellow giving a quantitative estimate of peak piston crown surface temperatures. Colour change takes
place when the material surface temperature is sustained for more than 10 min, which was suitable
Energies 2017, 10, 329
6 of 12
Energies 2017, 10, 329
6 of 12
for this
application since it was not required to measure short-term flame impingement.
Table 1
◦
shows the temperature in C for each colour change as a function of time. There is no benchmark for
Colour change takes place when the material surface temperature is sustained for more than 10 min,
this technique
slow speed
engines
as itit was
used
the first
time forflame
this type
which wason
suitable
for thisdiesel
application
since
was probably
not required
to for
measure
short-term
of application.
impingement. Table 1 shows the temperature in ℃ for each colour change as a function of time.
There is no benchmark for this technique on slow speed diesel engines as it was probably used for
Table
Thermal
paint temperature (◦ C) indication as function of time.
the first time for
this 1.
type
of application.
Colour
Change
5 mintemperature
10 min (℃)15indication
min
30asmin
60 min
Table
1. Thermal paint
function
of time. 120 min
Yellow to dusty grey
313
277
Colour Change
5 min
10 min
15 min
30 min
60 min
120 min
Dusty grey to yellow
525
482
450
435
Yellow to dusty grey
313
277
Yellow to orange
601
590
584
573
562
552
Dusty grey to yellow
525
482
450
435
Orange to green
699
682
672
656
641
625
Yellow to orange
601
590
584
573
562
552
Green to brown
806
792
784
770
756
742
Orange to green
699
682
672
656
641
625
Brown to grey green 806 981
948
929
899 770 869
840
Green to brown
792
784
756
742
Matt glaze
1235
-929
- 899
Brown to grey green
981
948
869
840
Matt glaze
1235
-
-
-
-
-
Piston crown temperature estimation was carried out on a Sulzer RTA 84 T engine (Wärtsilä Sulzer,
Piston crown temperature estimation was carried out on a Sulzer RTA 84 T engine (Wärtsilä
Cranford, NJ, USA), where high rates of hot corrosion were evident on a unit and also on a unit where
Sulzer, Cranford, NJ, USA), where high rates of hot corrosion were evident on a unit and also on a
piston
crown wear rates were normal (Figure 3a). The specifications of the selected Sulzer engine is
unit where piston crown wear rates were normal (Figure 3a). The specifications of the selected
listedSulzer
in Table
2. The
ambient
pressure
1.025 bar,
and is
the
pressure
ratio
the compressor
engine
is listed
in Table
2. The is
ambient
pressure
1.025
bar, and
the of
pressure
ratio of theis 1.7.
For the
thermal
paint
crown
trials,
the
engine
load
was
increased
over
a
period
of
2
h
to
followed
compressor is 1.7. For the thermal paint crown trials, the engine load was increased over a 75%
period
of
by 302min
operation
at
95%
with
a
new
piston
crown
fitted
in
each
case.
This
was
to
make
sure
that the
h to 75% followed by 30 min operation at 95% with a new piston crown fitted in each case. This was
to cooling
make sure
the were
pistonclean
cooling
oil any
boresexcessive
were clean
and any excessive
could not
piston
oilthat
bores
and
temperature
couldtemperature
not be attributed
tobe
piston
attributed
to piston crown
inefficiency.
The trapped
valueson
during
experiments
on
crown
cooling inefficiency.
Thecooling
trapped
lambda values
during lambda
experiments
the engine
were between
the2.0
engine
between
1.6 andvalue
2.0 with
an overall
lambda value
of would
3.2. Theco-relate
measuredtovalues
1.6 and
withwere
an overall
lambda
of 3.2.
The measured
values
0.9–1.2 on
would co-relate to 0.9–1.2 on a continuous combustion rig as diesel engine combustion is
a continuous combustion rig as diesel engine combustion is heterogeneous, heavily dependent on air
heterogeneous, heavily dependent on air fuel ratio distribution in a relatively large combustion
fuel ratio distribution in a relatively large combustion chamber. Design trapped lambda value is in
chamber. Design trapped lambda value is in the range of 2–2.2. The compression pressure was
the range
of 2–2.2. The compression pressure was around 110 bar at the start of fuel injection and
around 110 bar at the start of fuel injection and maximum pressure was around 130–135 bar. The
maximum
pressure
was the
around
pressure
makes
the flame
compact,
but the
high pressure
makes
flame130–135
compact,bar.
butThe
the high
relative
expansion
of flame
with lower
lambda
relative
expansion
of
flame
with
lower
lambda
values
is
consistent.
values is consistent.
(a)
(b)
(c)
Figure 3. View of thermal paint on the top of the piston crowns; (a) before temperature estimation
Figure 3. View of thermal paint on the top of the piston crowns; (a) before temperature estimation trial;
trial; (b) after trial with high wear rate; (c) after trial with normal wear rate.
(b) after trial with high wear rate; (c) after trial with normal wear rate.
Table 2. Specifications of the selected Sulzer engine.
Table 2. Specifications of the selected Sulzer engine.
Engine Type
Maker/Model
Engine Type
Operating cycle
Maker/Model
Rated
power
Operating cycle
Rated speed (rpm)
Rated power
Bore (mm)
Rated
speed (rpm)
Stroke
(mm)
Bore
(mm)
Compression
ratio
Stroke (mm)
Compression ratio
Turbocharging system
Slow Speed
Sulzer RTA 84T-D
Slow Speed
Two stroke
Sulzer4100
RTAkW/cylinder
84T-D
Two stroke
74
4100 kW/cylinder
840
74
840 3150
3150 17:1
17:1
Constant pressure
Energies 2017, 10, 329
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Figure 3b shows that the colour shades on completion of the trial on the piston crown with the
highest wear rate, which are very close to matt glaze and the surface temperatures around the hot
zone, are probably much in excess of the design temperature. A comparison of the thermal paint
results for the best and worst crown (Figure 3a,b) shows that the latter with normal wear rate has
a dusty grey shade with yellowish tinge at places. If we apply a 2 h correction to the change from
dusty grey to yellow, the temperature would be considered to be in the vicinity of 435 ◦ C. The test
results demonstrate that a high surface temperature and salt deposition on the crown in the heavily
burned away regions could have been caused by flame and fuel impingement, respectively. As in
an ideal scenario, the flame should be compact and should not come in contact with the piston crown
or exhaust valve.
4. Estimation of Flame Temperature with Different λ Values
In the previous section, it has been identified that a voluminous flame in the vicinity of combustion
chamber components will result in high thermal stress as a function of the flame temperature.
Meanwhile, voluminous flame production at low oxygen concentration. It is necessary to estimate the
flame temperature for various λ values to identify the scale of the problem.
Calculations were made for a steady flow constant pressure combustion chamber operating on
gas oil to estimate the adiabatic flame temperature and residual oxygen volume in the exhaust gas
products. The fuel combustion is assumed to be complete, and the products will be CO2 , H2 O and N2 .
The classic combustion equation, where a is the excess air coefficient is given by:
C12 H26 + 18.5a [O2 + 3.76N2 ] = bCO2 + cH2 O + dO2 + 69.56aN2 .
(1)
An elemental balance of the reactants and products returns the following molar volumes of the
products of combustion, i.e., b = 12, c = 13, d = 18.5 × (a − 1).
From classical thermodynamics and the first law of a differential form of the general energy
equation for a control volume (CV) can be obtained in molar terms:
−
.
.
.
.
dE
= Q − W + ∑ in N i ei − ∑ out N j e j ,
dt
.
(2)
.
where E is the total CV system energy; Q is the rate of heat transfer to/from the CV; W is rate of work
.
.
transfer to/from the CV; N i is the molar flow rate of material into the CV; N j is the molar flow rate of
material out of the CV; ei is the intensive total energy per unit mole of material into the CV; and e j is
the intensive total energy per unit mole of material out of the CV.
If all the properties within the CV, as well as those transferred across the CV boundary, do not
change with time (steady state and steady flow), then the rate of change of energy must remain
constant. Therefore,
dE
=0
(3)
dt
and
.
.
Q−W =
.
.
∑ in N i ei − ∑ out N j e j .
(4)
Considering the adiabatic combustion (at constant pressure), the enthalpy of reactants will be
equal to the enthalpy of products which results in an increase of temperature of products:
.
.
∑ in N i ei = ∑ out N j e j .
(5)
The intensive total energy ei into the CV is the sum of potential, kinetic, internal and chemical
potential energy. In the combustion process, the kinetic and potential energy will be negligible
compared to the internal and chemical potential energy; therefore, the total energy Ei , into the CV
system can be written as:
Energies 2017, 10, 329
8 of 12
Ei =
∑ Ni
ui + h0f
= ∑ Nj hi h T i − RT + h0f i ,
(6)
where h0f is the heat of formation and hi h T i is the enthalpy at T (in kelvin).
As the general energy equation equates the heat and work transfer rates to changes in energy
levels of materials that undergo state changes, it is the difference between the energy terms that is of
interest. Returning to the assumption that combustion takes place at constant pressure, an enthalpy
balance of the products and reactants can be used to calculate the temperature of the flame in the
cylinder for different values of λ. Therefore, equation above becomes:
∑ jprod Nj [h0f +
hhTj i − hhT0 i ] = ∑ jreact Ni [h0f + {hh Ti i − hh T0 i}] ,
j
i
(7)
where T0 is the thermodynamic reference temperature of 298 K.
Re-writing the above equation in terms of the products gives:
∑ jprod Nj h j = 12[h0f + ∆hhT4 i]CO2 + 13[h0f + ∆hhT4 i] H2O + 18.5(a − 1)[h0f +
∆hh T4 i]O2 + 69.56a[h0f + ∆hh T4 i]
N2
.
(8)
Re-writing the above equation in terms of the reactants gives:
∑ ireact Ni hi = 1[h0f + ∆hhTf i]C12 H26 + 18.5[h0f + ∆hhT3 i]O2 + 69.56a13[h0f + ∆hhT3 i] N2 ,
(9)
where Tf is the temperature of the fuel; T3 is the temperature at the start of constant pressure combustion
process; and T4 is the temperature at the end of constant pressure combustion process.
The use of enthalpy balance to calculate the flame temperature at fixed λ can be best understood
with the following example in which combustion is assumed to take place under stoichiometric
conditions, i.e., a = 1, and inlet air temperature is 400 K with a corresponding pressure of 1.46 bar.
The general combustion equation can be rewritten as:
C12 H26 + 18.5[O2 + 3.76N2 ] = 12CO2 + 13H2 O + 69.56N2 .
(10)
Using the enthalpy balance equation and the JANAF (Joint Army-Navy-Air Force)
thermo-chemical tables, gives:
∑ jprod Nj h j = 12[94054 + ∆hhT4 i]CO2 + 13[57798 + ∆hhT4 i] H2O + 69.56[0 + ∆hhT4 i] N2 ,
∑ ireact Ni hi = 1[85370 + 0]C12 H26 + 18.5[0 + 724]O2 + 69.5613[0 + 710] N2 .
(11)
(12)
The above two equations gives:
12[∆hh T4 i]CO2 + 13[∆hh T4 i] H2 O + 69.56[∆hh T4 i] N2 = 1857433 kcal .
(13)
The temperature of products from the equation above can be obtained by iteration is equal to
2530 K. Calculation code was written in Matlab software (The MathWorks, Inc., Natick, MA, USA) and
used to derive temperatures for a range of λ values at different amount of air preheat.
The data in Figure 4 shows the calculation results (using the energy balance equation, where the
enthalpy of products is always equal to the enthalpy of reactants) starting from stoichiometric condition.
This is a more complex method of obtaining the flame temperature compared to the simpler heat
release approach. However, the advantage of this method is that it will provide a more accurate
temperature estimate because the variation in gas properties due to temperature and composition
changes that are taken into account. This theoretical analysis does not include dissociation effects,
which will result in a reduction in flame temperatures when λ values approach unity.
Energies
2017,
10,10,
329
Energies
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Energies
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329
Adiabatic
Adiabatic flame
flame temperature
temperature (K)
(K)
9 of
12
of12
12
99of
Figure4.4.Adiabatic
Adiabaticflame
flametemperature
temperaturefor
fordifferent
differentλλvalues
valueswith
withdifferent
differentpreheats.
preheats.
Figure
Figure 4. Adiabatic flame temperature for different λ values with different preheats.
Adiabatic
Adiabatic flame
flame temperature
temperature (K)
(K)
Theresults
resultsin
inFigure
Figure55demonstrate
demonstratethe
therelationship
relationshipbetween
betweenflame
flametemperature
temperatureand
andλλvalues,
values,
The
The results
in Figure
5 demonstrate
the relationship
between flame
temperature
and λ values,
indicating
rapid
increase
in flame
flame temperature
temperature
as stoichiometric
stoichiometric
combustion
approached.
The
indicating
aa rapid
increase
in
as
combustion
isis approached.
The
indicating
a
rapid
increase
in
flame
temperature
as
stoichiometric
combustion
is
approached.
The
flame
flame temperature
temperature isis greatest
greatest at
at stoichiometric
stoichiometric condition
condition and
and reduced
reduced with
with increase
increase in
in λ.
λ. For
For fuel
fuel
flame
temperature
is greatest
at stoichiometric
condition
and reduced
with increase
in λ. For fuel
rich
rich conditions,
conditions,
the data
data
in Figure
Figure 55 shows
shows
the theoretical
theoretical
adiabatic
flame temperature
temperature
without
rich
the
in
the
adiabatic
flame
without
conditions,
thefor
data
in Figure
5values.
shows the theoretical adiabatic flame temperature without dissociation
dissociation
for
different
dissociation
different
λλvalues.
for different λ values.
Figure
Adiabatic
flame
temperature
under
fuel
rich
conditions
with
different
preheats.
Figure
Adiabatic
flame
temperature
under
fuel
rich
conditions
with
different
preheats.
Figure
5.5.5.
Adiabatic
flame
temperature
under
fuel
rich
conditions
with
different
preheats.
Calculations made
made so
so far
far were
were based
based on
on complete
complete combustion
combustion of
of hydrocarbon
hydrocarbon fuel
fuel in
inair
air with
with
Calculations
Calculations made so far were based on complete combustion of hydrocarbon fuel in air with the
the resulting
resulting in
in CO
CO22,, H
H22O,
O, N
N22 and
and O
O22 products.
products. This
This method
method for
for flame
flame temperature
temperature calculation
calculation isis
the
resulting in CO2 , H2 O, N2 and O2 products. This method for flame temperature calculation is accurate
accurate for
for temperatures
temperatures below
below 1500
1500 KK as
as dissociation,
dissociation, and
and breakdown
breakdown of
of products
products backing
backing their
their
accurate
for temperatures below 1500 K as dissociation, and breakdown of products backing their original
original elements
elements does
does not
not take
take place.
place. The
The effect
effect of
of dissociation
dissociation results
results in
in aa reduction
reduction of
of the
the flame
flame
original
elements does not take place. The effect of dissociation results in a reduction of the flame temperature
temperature and
and this
this becomes
becomes significant
significant when
when temperatures
temperatures are
are in
in excess
excess of
of 2000
2000 K.
K. In
In the
the
temperature
and this becomes significant when temperatures are in excess of 2000 K. In the dissociation mechanism,
dissociation mechanism,
mechanism, some
some of
of the
the CO
CO22 will
will split
split back
back into
into CO
CO and
and O
O22,, and
and H
H22O
O into
into H
H22 and
and O
O22..
dissociation
some of the CO2 will split back into CO and O2 , and H2 O into H2 and O2 . The results show that
The results
results show
show that
that there
there isis always
always some
some oxidant
oxidant left
left in
in the
the products.
products. Dissociation
Dissociation isis an
an
The
there is always some oxidant left in the products. Dissociation is an endothermic process and will
endothermicprocess
processand
andwill
willreduce
reducethe
theamount
amountof
ofenergy
energyreleased.
released.The
Theeffect
effectof
ofdissociation
dissociationon
onthe
the
endothermic
reduce the amount of energy released. The effect of dissociation on the peak flame temperature is more
peak flame
flame temperature
temperature isis more
more pronounced
pronounced when
when operating
operating close
close to
to stoichiometric
stoichiometric conditions,
conditions,
peak
pronounced when operating close to stoichiometric conditions, which is shown in Figure 6, with no
whichisisshown
shownin
inFigure
Figure6,6,with
withno
nodissociation
dissociationbeing
beingcompared
comparedin
inthe
thesame
samefigure.
figure.
which
dissociation being compared in the same figure.
10 10
of 12
of 12
Adiabatic flame temperature (K)
Energies
2017,
10, 10,
329329
Energies
2017,
Figure
Adiabatic
flame
temperature
with
and
without
dissociation
with
different
preheats.
Figure
6. 6.
Adiabatic
flame
temperature
with
and
without
dissociation
with
different
preheats.
5. Discussion
5. Discussion
For most diesel engines, other than the homogeneous charge compression ignition engines, the
For most diesel engines, other than the homogeneous charge compression ignition engines, the air
air to fuel ratio is stratified across the combustion space. This results in non-uniform temperatures
to fuel ratio is stratified across the combustion space. This results in non-uniform temperatures
across the piston crown in the radial and circumferential direction In theory, as discussed above,
across the piston crown in the radial and circumferential direction In theory, as discussed above,
maximum flame temperature is observed at λ = 1. However, in practice, due to non-uniformity in the
maximum flame temperature is observed at λ = 1. However, in practice, due to non-uniformity in
air-fuel mixtures, this occurs at λ ≅ 1.2 (for continuous combustion systems). For diesel engines,
the air-fuel mixtures, this occurs at λ ∼
= 1.2 (for continuous combustion systems). For diesel engines,
with intermittent combustion, the value of λ at which maximum flame temperature is observed
with intermittent combustion, the value of λ at which maximum flame temperature is observed could
could be higher, possibly in the region of λ = 1.6 to λ = 1.8. An engine operating with λ = 1.6 will have
be higher, possibly in the region of λ = 1.6 to λ = 1.8. An engine operating with λ = 1.6 will have its
its combustion chamber components running significantly hotter than those operating at greater
combustion chamber components running significantly hotter than those operating at greater values
values of λ. This indicates that engine components responsible for maintaining the pre-requisite air
of λ. This indicates that engine components responsible for maintaining the pre-requisite air to fuel
to fuel ratio within the combustion space are critical since malfunction will result in thermal
ratio within the combustion space are critical since malfunction will result in thermal overloading of
overloading of combustion chamber components.
combustion chamber components.
It could be argued that, in fuel rich conditions, the maximum adiabatic flame temperature is less
It could be argued that, in fuel rich conditions, the maximum adiabatic flame temperature is
than stoichiometric conditions; therefore, combustion chamber components should not fail due to
less than stoichiometric conditions; therefore, combustion chamber components should not fail due
temperatures exceeding design threshold values. However, there are significant problems with
to temperatures exceeding design threshold values. However, there are significant problems with
combustion chamber components failing through excessive temperatures in fuel rich conditions. This
combustion chamber components failing through excessive temperatures in fuel rich conditions.
does not occur under lean burn conditions, where the flame is more compact and a boundary of air
This does not occur under lean burn conditions, where the flame is more compact and a boundary of
exists between the flame and the component surface. This voluminous flame comes into contact with
air exists between the flame and the component surface. This voluminous flame comes into contact
the combustion chamber components, increasing the rate of heat input, and also inducing the engine
with the combustion chamber components, increasing the rate of heat input, and also inducing the
to thermal overload.
engine to thermal overload.
Conclusions
6. 6.
Conclusions
this
research,
most
probable
cause
diesel
engines’
thermal
overload
is discussed
and
In In
this
research,
thethe
most
probable
cause
of of
diesel
engines’
thermal
overload
is discussed
and
validated.
Test
results
indicated
that
when
the
fuel
rate
of
the
flame
visualisation
test
rig
was
validated. Test results indicated that when the fuel rate of the flame visualisation test rig was increased
constant
air mass
flow
ratethen
or vice
thenThe
λ will
decrease.
The flame
also became
at increased
constant airatmass
flow rate
or vice
versa,
λ willversa,
decrease.
flame
also became
voluminous
due
voluminous
due
to
the
lack
of
active
radicals.
The
exhaust
temperature
increased
from
775
to
1000
◦
to the lack of active radicals. The exhaust temperature increased from 775 to 1000 C with λ changing
°C
with
λ
changing
from
1.12
to
0.71.
This
high
temperature
flame
when
close
to
or
in
contact
with
from 1.12 to 0.71. This high temperature flame when close to or in contact with any combustion
any combustion
component
would result
an increased
rate of heat
transfer through
chanter
componentchanter
would result
in an increased
rate in
of heat
transfer through
the component
and itsthe
component
and
its
consequent
failure.
Temperature
indicating
paints
were
applied
on
two
piston
consequent failure. Temperature indicating paints were applied on two piston crowns to investigate
crowns
to
investigate
the
effect
of
voluminous
flame
on
component
surface
temperature.
the effect of voluminous flame on component surface temperature. A comparison of the thermal paint A
comparison
of that
the thermal
paint
results
indicated
thatshowed
the piston
withgrey
normal
ware
rate
showed a
results
indicated
the piston
with
normal
ware rate
a dusty
shade
with
yellowish
dusty
grey
shade
with
yellowish
tinge
at
places,
while
the
piston
with
high
rates
of
hot
corrosion
tinge at places, while the piston with high rates of hot corrosion is very close to matt glaze (much in is
very close to matt glaze (much in excess of the design temperature). The test results proved that high
Energies 2017, 10, 329
11 of 12
excess of the design temperature). The test results proved that high surface temperatures and salt
deposition on the crown in the heavily burned regions could have been caused by the flame and fuel
impingement, respectively.
Acknowledgments: This work was funded using the EPSRC (Engineering and Physical Sciences Research Council)
Impact Acceleration Account EP/K503885/1. Data supporting this publication is openly available under an Open
Data Commons Open Database License. Additional metadata are available at: http://dx.doi.org/10.17634/
160151-2. Please contact Newcastle Research Data Service at rdm@ncl.ac.uk for access instructions.
Author Contributions: Sangram Kishore Nanda and Anthony Paul Roskilly conceived and designed the
experiments; Sangram Kishore Nanda performed the experiments; Sangram Kishore Nanda and Boru Jia analyzed
the data and wrote the paper; and Andrew Smallbone revised the manuscript.
Conflicts of Interest: The authors declare no conflict of interest.
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