See discussions, stats, and author profiles for this publication at: https://www.researchgate.net/publication/228855636
Numerical Evaluation by Models of Load and Spark Timing Effects
on the In-Cylinder Heat Transfer of a SI Engine
Article in Numerical Heat Transfer Applications · October 2009
DOI: 10.1080/10407780903244312
CITATIONS
READS
9
1,890
5 authors, including:
Ali Sanli
Cenk Sayin
Piri Reis Vocational and Technical Anatolian High School
Marmara University
9 PUBLICATIONS 80 CITATIONS
58 PUBLICATIONS 3,126 CITATIONS
SEE PROFILE
SEE PROFILE
Metin Gumus
Ibrahim Kilicaslan
Marmara University
Kocaeli University
44 PUBLICATIONS 1,683 CITATIONS
15 PUBLICATIONS 499 CITATIONS
SEE PROFILE
SEE PROFILE
Some of the authors of this publication are also working on these related projects:
Variable Valve Timing and Its Impacts on Air Parameters of an SI Engine View project
EFFECT of USING DUAL FUEL (BIOGAS-DIESEL) on the PERFORMANCE and EMISSIONS of a LOW HEAT REJECTION DIESEL ENGINE View project
All content following this page was uploaded by Metin Gumus on 15 May 2014.
The user has requested enhancement of the downloaded file.
Numerical Heat Transfer, Part A, 56: 444–458, 2009
Copyright # Taylor & Francis Group, LLC
ISSN: 1040-7782 print=1521-0634 online
DOI: 10.1080/10407780903244312
NUMERICAL EVALUATION BY MODELS OF LOAD
AND SPARK TIMING EFFECTS ON THE IN-CYLINDER
HEAT TRANSFER OF A SI ENGINE
A. Sanli1, C. Sayin2, M. Gumus2, I. Kilicaslan1,
and M. Canakci 1,3
1
Department of Mechanical Education, Kocaeli University, Izmit, Turkey
Department of Mechanical Education, Marmara University, Istanbul, Turkey
3
Alternative Fuels R&D Center, Kocaeli University, Izmit, Turkey
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
2
The aim of this study is to examine numerically the effects of spark timing and load
parameters on the in-cylinder heat transfer of a SI engine by using experimental engine
test data. For the investigation, a four-stroke, air-cooled, single-cylinder SI engine was
tested at different spark timings and loads at a single engine speed of 2000 rpm. Woschni,
Hohenberg, and Han models were employed to estimate the in-cylinder heat transfer
coefficient in the case of different test conditions because of being favorable models on
the SI engine operations. The evaluations show that the in-cylinder heat transfer characteristics of the air-cooled SI engine strongly depend on the load while they slightly depend on
the spark timing.
1. INTRODUCTION
In design of the internal combustion engines (ICEs), thermal behavior of the
engine is an important topic because of its effect on indicated efficiency, power
output, emissions, lubricant, and cooling capacity of the engine. Heat transfer to
the combustion chamber walls influences the indicated efficiency because it reduces
the cylinder pressure and temperature and thus the work done on the piston
decreases per cycle. Heat loss to the cooling system of an ICE is approximately
30% of the total fuel energy supplied to the engine during the one working cycle.
About half of this loss is due to the in-cylinder heat transfer and the rest is due to
the cylinder head and to the exhaust port [1, 2].
From the in-cylinder points of view, the heat transfer changes locally
instantaneous temperatures that have exponential dependence on controlling the
formation rate of nitric oxide emissions (NOx). High temperatures lead to thermal
Received 4 March 2009; accepted 7 July 2009.
Experimental data used in this study were obtained from the project supported by The Scientific
Research Council of Marmara University, Project No. BSE-075=131102. The authors are grateful to
the institutes and the individuals involved in making this work possible.
Address correspondence to Mustafa Canakci, Department of Mechanical Education, Kocaeli
University, Umuttepe Yerleskesi, Umuttepe, Izmit 41380, Turkey. E-mail: mustafacanakci@hotmail.com
444
LOAD AND SPARK TIMING EFFECTS ON HEAT TRANSFER OF A SI ENGINE
445
NOMENCLATURE
a, C1, C2, C3
D
h
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
k
l
m
n
Nu
P
Pr
q
R
Re
S
T
Up
V
W
x
ABDC
AFR
ATDC
BBDC
constants
cylinder bore, m
heat transfer coefficient,
W=m2 K
thermal conductivity of the gas,
W=m K
connection rod length, m
exponent or air mass, kg
exponent or engine speed, rpm
Nusselt number
in-cylinder pressure, bar
Prandtl number
heat flux, MW=m2
gas constant, kJ=kg K
Reynolds number
cylinder stroke, m
temperature, K
(2Sn=60) piston mean velocity,
m=s
volume, m3
effective gas velocity, m=s
distance from top dead
center, m
after bottom dead center
air fuel ratio
after top dead center
before bottom dead center
BTDC
CA
CVH
q
h
k
before top dead center
rank angle
compound valve angle
hemispherical combustion
chamber
double over head camshaft
revolutions per minute
spark ignition
top dead center
dynamic viscosity of cylinder
gases, Pa.s
density of cylinder gases, kg=m3
crank angle, o
the ratio of specific heats
Subscripts
a
c
d
g
in
mg
p
r
sp
w
air
combustion
displacement
gas
inlet
mean gas
piston
reference
spark plug
wall
DOHC
rpm
SI
TDC
m
stresses of material, and impact on fatigue failure limits of various engine
components, thereby causing the cracks in cylinder head or deforming cylinder bore
dimension and valve stems. Keeping below the wall temperature of the combustion
chamber from the certain limits is necessary to prevent the oil film from deterioration
or oxidizing and its viscosity from diminishing, so the oil film temperature should
not exceed about 200 C. In order to avoid from pre-ignition and knock risks resulting from overheated spark-plug electrodes and exhaust valves in the spark ignition
(SI) engines, the spark-plug and valves must be kept cool. In addition, the heat
transfer to the inflowing air or charge reduces volumetric efficiency on account of
reducing the inflowing fluid density [3].
A number of experimental and theoretical studies concerning the effects of
various engine operating parameters on the heat transfer to the combustion chamber
walls have been published since the last decades. Alkidas and Myers [4] investigated
the effect of the air-fuel ratio (18, 16, 14, and 11.5 AFR) and volumetric efficiency
(40, 50, 60%) on the heat flux and showed that the highest heat flux to the combustion chamber walls occurred at the near stoichiometric ratio (16 AFR) and the
increasing volumetric efficiency resulted in an increasing in the peak heat fluxes.
Alkidas [5] measured the heat flux at different engine speeds and loads and showed
that the increasing engine speed and load increased the heat flux to the combustion
chamber walls. He also compared the heat flux values obtained by the Woschni
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
446
A. SANLI ET AL.
model with the experimentally measured ones and observed that the one obtained by
the model was in good agreement with the experimental results. Karamangil et al [6].
investigated parametrically how the convective heat transfer coefficients (HTCs) for
the gas and coolant side vary with different engine operating parameters, such as the
engine speed, compression ratio, excess air coefficient, combustion duration, inlet
pressure, and temperature. Shojaefard et al [7]. studied transient thermal analysis
of an exhaust valve by using the finite-element method. The temperature distribution
and resultant thermal stresses at each opening and closing time were obtained.
Rakopoulos and Mavropoulos [8] disclosed to be the significant differences between
the overall and local HTC variations of an air-cooled direct injection (DI) diesel
engine, and they also presented the effects of engine speed and load on the incylinder heat transfer.
In a previous study [9], the effects of engine speed and load on the in-cylinder
heat transfer of a water-cooled, four-stroke diesel engine have been investigated. In
this study, in-cylinder heat transfer characteristics of an air-cooled SI engine were
investigated. However, this study is not presenting something new, but rather, it is
intended to raise the awareness and importance of the in-cylinder heat transfer in
the ICEs, and to present how the in-cylinder heat transfer characteristics of a SI
engine are varying under different operating parameters by using some models in
the light of engine test data.
2. CALCULATION OF THE IN-CYLINDER HEAT TRANSFER
In the development of ICEs, the heat transfer from the burning gases to the
in-cylinder walls is the main subject of numerous investigations. Historically, Nusselt
made pioneer attempts on the in-cylinder heat transfer based on the combustion
experiments in a spherical bomb in 1923. It was intended to predict the steady-state
or time-average heat flux, but it has often been used for the prediction of instantaneous heat flux because it was expressed in terms of instantaneous P and T values.
Then, Eichelberg [10] developed a model correlated in large-scale two-stroke and
four-stroke diesel engines. It is preferred by most investigators owing to being simple
relatively easy to use. However, when the unit analysis is made, it is shown that it
does not rely on the theoretical bases and yield dimensionally consistent results
[11]. It is believed that the fault is stemmed from the experiments with poor
technology in those years [12]. Oguri [13] used the Eichelberg model to estimate
the in-cylinder HTC of a water-cooled, single cylinder SI engine and observed that
the model’s results coincided well with the experimental ones at the expansion stroke,
but for the compression stroke it was not good. Therefore, he developed a new
in-cylinder heat transfer correlation by making some variations in the Elser formula,
including entropy change, specific heat under constant pressure change, and Nusselt
relation. In the following years, an approach, including Nusselt, Reynolds, and
Prandtl number relationships, has been suggested for the turbulent flow inside the
circular tubes, as follows.
Nu ¼ aRem Prn
ð1Þ
LOAD AND SPARK TIMING EFFECTS ON HEAT TRANSFER OF A SI ENGINE
447
Pr can be omitted if the fluid is a gas. In this case, its effect might be included in
the a coefficient. Nu includes the HTC. So, by expanding Nu and Re numbers the
equation becomes
qUp l m
hl
¼a
k
m
ð2Þ
The value of exponent m generally ranges from 0.7 to 0.8 for forced convection
with turbulent flow [14]. By emphasizing that the exponent m is equal to 0.8,
Woschni [15] developed the in-cylinder HTC below by using Eq. (2).
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
hðhÞ ¼ aPðhÞ0:8 Tg ðhÞ0:55 D0:2 W ðhÞ0:8
ð3Þ
He assumed that the characteristic length l, piston mean velocity Up, and fluid
density q can be replaced, respectively, with the cylinder bore diameter D, gas velocity W, and P=RT. Thermal conductivity k and dynamic gas viscosity m are proportional to T 0.75 and T 0.62, respectively [3]. Effective gas velocity W was
expressed by Woschni as follows.
Vd Tr
W ðhÞ ¼ C1 Up þ C2
ðPðhÞ Pm ðhÞÞ
Pr Vr
ð4Þ
where Vd is the displacement volume, Tr, Pr, and Vr are evaluated, respectively, as
temperature, pressure, and, volume at any reference state, such as inlet valve closing
or combustion start. (P(h) Pm(h)) is the pressure rise resulted from combustion.
While P(h) is the firing pressure, Pm(h), calculated from the Watson-Janota model
[16], is the motoring pressure at the same crank angle with P(h). Values suggested
for C1 and C2 are
C1 ¼ 6:18
C2 ¼ 0 for gas exchange period
C1 ¼ 2:28
C2 ¼ 0 for compression period
C1 ¼ 2:28
C2 ¼ 3:24 103
for combustion and expansion period
Shayler et al. [17] investigated two methods of determining the rate of heat
transfer from the combustion chamber of three SI engines, which are Ford’s 1.1 L
Valencia, 1.6 L CVH, and 2.0 L DOHC at the different speed, load, spark timing,
and AFR conditions. In the first method that is an analysis based on the application
of the first law of thermodynamics. The instantaneous cylinder heat transfer rate was
found to be subject to large errors stemmed from uncertainties in the gas properties.
In the second method, the instantaneous cylinder heat transfer rates were calculated
using the Woschni, Eichelberg, and Annand models and were integrated over the
engine cycle to generate cycle-averaged heat transfer rates. The rates were compared
with a cylinder heat transfer rate deduced from the difference between heat transfer
to the coolant and heat transfer from the exhaust gases to the exhaust port surface.
448
A. SANLI ET AL.
The comparisons showed that the best agreement was obtained with the Woschni
correlation. The researchers note that this is important information about the
models’ comparison. From the results presented, it appears that the original
coefficient values (for a, C1, and C2) proposed by Woschni can be applied without
major adjustment. In addition, it should be noted that the Woschni formula not
only includes the effect of convection but also includes radiation effect in a lumped
form [2].
In the following years, Hohenberg [18] proposed an in-cylinder heat transfer
equation by modifying the Woschni equation in the DI diesel engines with swirl.
Equation (5) describes the in-cylinder HTC proposed by Hohenberg.
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
hðhÞ ¼ aPðhÞ0:8 Tg ðhÞ0:4 V ðhÞ0:06 ðUp þ 1:4Þ0:8
ð5Þ
Zeng et al. [19] improved a method adding the capability of predicting the
effect of engine spark-timing on the cylinder pressure and applied it to a SI engine.
The cylinder pressure was reconstructed including the third input of spark timing
along with speed and load. Comparisons between measured and reconstructed cylinder pressures demonstrated that the method was applicable over a wide range of the
SI engine operating. The reconstructed cylinder pressure was used to achieve the
in-cylinder heat transfer and heat release analyses. In order to study the heat transfer, they used the Hohenberg model.
Han et al. [20] successfully established a new in-cylinder heat transfer model for
the SI engines based on the turbulent flow. They used 0.75 for the value of the
exponent m given in Eq. (2), taking into account the results of experiments carried
out on 19 different ICEs by Taylor and Toong [21]. Consequently, the formula
became as follows.
hðhÞ ¼ C1 PðhÞ0:75 Tg ðhÞ0:465 D0:25 W ðhÞ0:75
ð6Þ
For the effective gas velocity W, Han et al. assumed that an increase in the gas
velocity during the combustion period was caused by the heat release revealed from
the chemical reactions of gas mixture. Using this assumption, the gas velocity W can
be expressed as follows.
W ¼ C2 Up þ C3 ðkPðhÞdV þ V ðhÞdPÞ
ð7Þ
The constants C1, C2, C3 and k values are 687, 0.494, 0.73 106, and 1.35,
respectively. Any comparison has not been coincided between the Han model and
another model in the literature. More comprehensive information dealing with the
in-cylinder heat transfer models can be found in references [1–3, 11, 22].
It is well known that the in-cylinder heat transfer is strongly influenced by the
combustion gas pressure and temperature. The main data to calculate these models
are the pressure, temperature, and volume in the cylinder. Once the pressure is
acquired, the temperature of the gases in the cylinder can be calculated using the
pressure, assuming the first law of thermodynamics. In the above equations, the
instantaneous cylinder volume V(h) as a function of crank angle can be found using
LOAD AND SPARK TIMING EFFECTS ON HEAT TRANSFER OF A SI ENGINE
449
the following well-known formulas.
V ðhÞ ¼ Vc þ
pD2
xðhÞ
4
ð8Þ
x(h) represents the distance from TDC, and thus according to the crank angle,
qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
xðhÞ ¼ l þ R R cosðhÞ þ ðl 2 R2 sin2 ðhÞÞ
ð9Þ
Applying the steady-state heat transfer assumption for the present study, the
heat transfer on the inside surface of the cylinder is found from the Newton’s cooling
law by multiplying the simultaneous HTC with the temperature difference, i.e.,
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
qðhÞ ¼ hðhÞðTg ðhÞ Tw Þ
ð10Þ
Figure 1 illustrates both the convective and radiative components of the heat
transfer between gas and in-cylinder wall, and conduction through the combustion
chamber wall, and convection to the air for air-cooled engines [23]. Mean gas
temperature Tmg is depicted with dotted lines. Heat from the burning gases is also
transferred to the cylinder wall indirectly by conduction through the piston rings
and skirts. However, it is beyond the scope of this study. In the present study, it
is focused on the heat transfer between gas and in-cylinder wall.
As mentioned previously, the temperature of the gases in the cylinder can
be calculated using the cylinder pressure and volume, applying the first law of
thermodynamics.
Tg ðhÞ ¼
PðhÞV ðhÞ
mR
Figure 1. Schematic view of the in-cylinder heat transfer process in an air-cooled engine.
ð11Þ
450
A. SANLI ET AL.
where P is the instantaneous in-cylinder pressure, V is the instantaneous
cylinder volume, m is the air mass in the cylinder, and R is the gas constant.
Tw inside wall temperature of combustion chamber varies with the engine
speed, load, equivalence ratio, start of combustion, charge motion, inlet temperature, wall material, and the coolant and combustion temperatures. It involves apparently too sophisticated to predict the wall temperature. Therefore, Tw was replaced
with spark-plug temperature Tsp in this study, as it was assumed in a heat transfer
study, which was performed by Wu et al. [24] for an air-cooled SI engine. Tsp is
formulated as follows.
Tsp ð CÞ ¼ ac1 n ac2 n2 ac3 Pin ac4 þ P2in ac5 þ n Pin ac6
ð12Þ
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
where n is engine speed (rev=s), Pin is the intake manifold pressure (bar), and aci
is the coefficients obtained from curve fitting by the reference.
3. EXPERIMENTAL APPARATUS AND PROCEDURE
The experiments were performed on a single cylinder, air-cooled, four-stroke
SI engine. Its specifications are given in Table 1. The engine was loaded by an electrical dynamometer rated at 10 kW and 380 volts. The load on the dynamometer was
measured by means of a strain gauge load sensor. An inductive pickup speed sensor
was used to determine the engine speed. The in-cylinder pressure was measured using
a Kistler Model 6052B air-cooled piezo-quartz pressure sensor mounted on the cylinder head. The pressure signals were then passed into a Kistler Model 5644A charge
amplifier. Crankshaft position was obtained by a crankshaft angle sensor in order to
determine cylinder gas pressure as a function of crank angle. The crank angle signal
was attained through an angle-generating device mounted on the main shaft. The
intervals of the cylinder pressure were 0.75 CA for all test cases and pressure average
of the consecutive 100 cycles was received. The start of combustion was accepted
2 CA after the spark instant because of a period of time delay between the spark
instant and start of combustion. The engine was fueled with the commercial grade
gasoline of 95 research octane number.
Table 1. Engine specifications
Brand=model
Engine type
Bore=stroke
Compression ratio
Connection rod length
Max. torque
Max. power
Inlet valve opening
Inlet valve closing
Exhaust valve opening
Exhaust valve closing
Briggs and Stratton=243431
Four stroke, single cylinder, air cooled
77.79 mm=82.55 mm
10:1
110 mm
30 Nm at 2400 rpm
7.35 kW at 3600 rpm
16 CA BTDC
44 CA ABDC
45 CA BBDC
15 CA ATDC
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
LOAD AND SPARK TIMING EFFECTS ON HEAT TRANSFER OF A SI ENGINE
451
Figure 2. Schematic diagram of the engine experimental setup.
The test engine gives the maximum brake torque (MBT) at the optimum spark
timing that is 20 CA BTDC. The electrical spark is provided by means of a magneto.
By changing the magneto position, the spark timings were adjusted to 17.5 and
22.5 CA BTDC. All tests were conducted on a test-bench, as shown in Figure 2.
Each of the tests was repeated three times and the average was taken. Before each
test, the engine was carefully regulated according to the catalogue values and all data
were collected after the engine was stabilized. The accuracy of the measurements and
the uncertainty in the calculated results are given in Table 2.
4. RESULTS AND DISCUSSION
Heat transfer rates generally increase during the compression and the expansion strokes, and the peak heat flux mostly occurs after TDC for a typical engine.
Table 2. Accuracy of the measurements and the uncertainty in
the calculated results
Measurements
Load
Speed
Crankshaft position
Charge amplifier
Temperatures
Calculated results
HTC
Heat flux
Accuracy
2 Nm
25 rpm
0.1 CA
1%
0.1 K
Uncertainty
2%
2.1%
452
A. SANLI ET AL.
On the other hand, heat fluxes might nearly decrease too small negative values
during the intake and exhaust strokes. Therefore, the data, obtained between the
period of intake valve closing and exhaust valve opening, are presented in
Figures 3–9. The figures include HTC and heat flux values of the Woschni,
Hohenberg, and Han models as a function of crank angle at different spark timings
and loads for 2000 rpm engine speed.
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
4.1. Evaluations of the Heat Transfer Models
The HTC and heat flux histories of three models examined are depicted in
Figure 3 as a function of crank angle at the MBT spark timing of the full load for
2000 rpm engine speed. At first it was observed that HTC and heat flux amounts
of the Han model were somewhat higher than those from the Woschni and
Hohenberg models. The peak value of the Han model is 30% higher than the
Woschni model, 37.6% higher than the Hohenberg model about the HTC; and the
heat flux is 31% higher than the Woschni model, 37.1% higher than the Hohenberg
model. A reason as to why the Han model is higher than the others may be because
the a constant has been correlated to 128 in the Woschni and Hohenberg models;
whereas, it has been correlated to 687 in the Han model, provided that pressure is
in the bar unit. On the other hand, exponent m in the Woschni and Hohenberg
models is higher than that in the Han model, i.e., 0.8 against to 0.75. However, it
has been believed that the exponent m has not been as dominant as the constant a
on the discrepancy. In view of the Hohenberg HTC and heat flux, when they were
compared with those of Woschni, the Hohenberg model overestimates the HTC
and heat flux of certain sections of the compression and expansion periods;
Figure 3. Comparisons of Han, Woschni, and Hohenberg heat transfer correlations for the spark timing
(20 CA BTDC) at full load—2000 rpm.
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
LOAD AND SPARK TIMING EFFECTS ON HEAT TRANSFER OF A SI ENGINE
453
Figure 4. Comparison of the Woschni HTCs and heat fluxes for different spark timings at full load.
however it underestimates at only peak regions. As pointed out by Soyhan et al. [25],
the Woschni model gives higher gas displacement velocity compared to the
Hohenberg model during the combustion phase; whereas, the Hohenberg model
gives a constant gas displacement velocity for the whole cycle. The heat transfer
results are comparatively seen in Figure 3, which are very close to each other.
Accordingly, in order to evaluate the spark timing and load effects on the in-cylinder
heat transfer, the results of each model were presented separately as a function of the
crank angle.
Figure 5. Comparison of the Woschni HTCs and heat fluxes for different spark timings at 50% load.
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
454
A. SANLI ET AL.
Figure 6. Comparison of the Hohenberg HTCs and heat fluxes for different spark timings at full load.
4.2. Heat Transfer Behaviors at Different Spark Timings and Loads
Figures 4 and 5 show the histories of the HTCs and heat fluxes obtained from
the Woschni model at different spark timings and load conditions. The peak HTC
and heat flux values of Woschni at spark timings of 17.5, 20, and 22.5 CA for full
load are 1577, 1606, and 1650 W=m2 K, and 0.45, 0.47, and 0.49 MW=m2, respectively; for the 50% load condition, they are 1421, 1428, and 1460 W=m2K, and
0.37, 0.38, and 0.40 MW=m2, respectively.
Figure 7. Comparison of the Hohenberg HTCs and heat fluxes for different spark timings at 50% load.
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
LOAD AND SPARK TIMING EFFECTS ON HEAT TRANSFER OF A SI ENGINE
455
Figure 8. Comparison of the Han HTCs and heat fluxes for different spark timings at full load.
The HTC and heat flux results of the Hohenberg model as a function of the
crank angle at different spark timings and loads at 2000 rpm can be seen in Figures
6 and 7. At this speed, the peak HTC and heat flux magnitudes of the 17.5, 20, and
22.5 CA BTDC spark timings of the full load which are 1482, 1521, and
1551 W=m2 K, and 0.43, 0.45, and 0.465 MW=m2, respectively; whereas, at 50% load
they are 1385, 1406, and 1441 W=m2K, and 0.37, 0.38, and 0.40 MW=m2, respectively, which are almost the same as the magnitudes of Woschni at the same load.
Figure 9. Comparison of the Han HTCs and heat fluxes for different spark timings at 50% load.
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
456
A. SANLI ET AL.
Figures 8 and 9 show the histories of the HTCs and heat fluxes obtained from
the Han model for the spark timing variations at the full load and 50% load at the
2000 rpm. At this speed, the peak values of the Han HTCs and heat fluxes for the
spark timings of 17.5, 20, and 22.5 CA BTDC at full load are 2071, 2093, and
2108 W=m2 K, and 0.60, 0.617, and 0.627 MW=m2, respectively; whereas, at 50%
load they are 1935, 1943, and 1985 W=m2K, and 0.522, 0.528, and 0.545 MW=m2,
respectively. In order to be seen easily, all peak values of the heat fluxes and HTCs
are listed in Table 3.
As can be seen in the above figures, in-cylinder HTCs and heat fluxes of the SI
air-cooled engine change with the variations of spark timing at the constant engine
speed and load. Taking spark timing earlier or later with respect to the MBT spark
timing at constant engine operating conditions causes the charge pressure and temperature to change near the TDC. Accordingly, in all models, earlier spark timing
leads to a slight increase in the peak heat fluxes and HTCs, and it leads the peak
values to happen at earlier CA than those of MBT spark timing. On the other hand,
later spark timing leads to a decrease in the peak heat fluxes and HTCs, and it leads
the peak values to happen at later CA than those of MBT spark timing. As for the
load, the larger the load, the more the heat will be rejected to the combustion chamber walls. Since the increasing load means more fuel-rich mixture in the combustion
chamber, the energy of the compressed mixture would increase and thereby increase
the heat transfer to the cylinder walls. Furthermore, the histories of the HTCs and
heat fluxes obtained from the Woschni, Hohenberg, and Han models are illustrated
to be on the same line until the spark instant in the compression period because the
engine operating parameters are constant until that point, but they would vary with
the spark timings.
At the 50% load, the peak heat flux and HTC values nearly remain constant at
the spark timing variations from 20 to 17.5 CA BTDC for each model, but at the full
load the corresponding variations are comparatively noticeable.
In the middle of the expansion stroke, at the same crank angle, it can be said
that the retarded spark timing gives higher HTC and heat flux compared with the
advanced spark timing for all heat transfer models. However, from the spark instant
until the peak HTC and heat flux points, it can be said that the advanced spark timing gives higher HTC and heat flux compared with the retarded spark timing for all
the heat transfer models.
Table 3. All peak heat flux and HTC values obtained in the tests
Full load
50% load
Spark timing ( CA BTDC)
17.5
20
22.5
17.5
20
22.5
h (W=m2K)
2071
1577
1482
0.60
0.45
0.43
2093
1606
1521
0.617
0.47
0.45
2108
1650
1551
0.627
0.49
0.465
1935
1421
1385
0.522
0.37
0.37
1943
1428
1406
0.528
0.38
0.38
1985
1460
1441
0.545
0.40
0.40
q (MW=m2)
Han
Woschni
Hohenberg
Han
Woschni
Hohenberg
LOAD AND SPARK TIMING EFFECTS ON HEAT TRANSFER OF A SI ENGINE
457
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
5. CONCLUSION
In this study, the effect of spark timing and load variations on the heat transfer
between gas and in-cylinder wall of an air-cooled, four-stroke, single cylinder SI
engine run at 2000 rpm was investigated by using the Woschni, Hohenberg, and
Han models. The conclusions given below are summarized from the present work.
When the spark timing is advanced from its original timing (spark timing for
MBT), the in-cylinder HTC and heat flux increase slightly and their peak values
occur at earlier crank angle than that of the original timing at constant load and
speed conditions. In contrast to advanced timing, when the spark timing is retarded
from its original timing, the in-cylinder HTC and heat flux decreases slightly and
their peak values occur at later crank angle than that of the original timing.
The in-cylinder HTCs and heat fluxes at constant spark timing vary with the
changing of engine load. With increasing the load, both the HTC and heat flux
magnitudes increase more than magnitudes occuring with variations of spark timing.
The Han model gives higher HTC and heat flux compared with the Woschni
and Hohenberg models under the same engine operating conditions.
Hohenberg model overestimates the HTC and heat flux at certain sections of
compression and exhaust period, while it underestimates peak HTC and heat flux
at only peak regions compared with the Woschni model.
From the beginning of spark timing toward the peak point, each model gives
higher HTC and heat flux relative to the original timing at the advanced spark
timing, but at the retarded spark timing corresponding HTC and heat fluxes have lower
values. Toward the end of the compression stroke, each model gives lower HTC and
heat flux values relative to the original timing at advanced spark timing, while at the
retarding spark timing corresponding HTC and heat fluxes have higher values.
REFERENCES
1. C. R. Ferguson and A. T. Kirkpatrick, Internal Combustion Engines—Applied Thermosciences, 2nd ed., pp. 221–252, John Wiley & Sons, New York, 2001.
2. G. Borman and K. Nishiwaki, Internal-Combustion Engine Heat Transfer, Prog. in
Energy Comb. Sci., vol. 13, pp. 1–46, 1987.
3. J. B. Heywood, Internal Combustion Engine Fundamentals, pp. 668–711, Mc-Graw Hill,
New York, 1988.
4. A. C. Alkidas and J. P. Myers, Transient Heat-Flux Measurements in the Combustion
Chamber of a Spark-Ignition Engine, Trans. of ASME—J. of Heat Transfer, vol. 104,
pp. 62–67, 1982.
5. A. C. Alkidas, Heat Transfer Characteristics of a Spark-Ignition Engine, Trans. of
ASME––J. of Heat Transfer, vol. 102, pp. 189–193, 1980.
6. M. I. Karamangil, O. Kaynakli, and A. Surmen, Parametric Investigation of Cylinder
and Jacket Side Convective Heat Transfer Coefficients of Gasoline Engines, Energy
Conversion and Management, vol. 47, no. 6, pp. 800–816, 2006.
7. M. H. Shojaefard, A. R. Noorpoor, D. A. Bozchaloe, and M. Ghaffarpour,
Transient Thermal Analysis of Engine Exhaust Valve, Numer. Heat Transfer A, vol. 48,
pp. 627–644, 2005.
8. C. D. Rakopoulos and G. C. Mavropoulos, Experimental Evaluation of Local
Instantaneous Heat Transfer Characteristics in the Combustion Chamber of Air-Cooled
Direct Injection Diesel Engine, Energy, vol. 33, pp. 1084–1099, 2008.
Downloaded By: [TÜBTAK EKUAL] At: 13:43 28 September 2009
458
A. SANLI ET AL.
9. A. Sanli, A. N. Ozsezen, I. Kilicaslan, and M. Canakci, The Influence of Engine Speed
and Load on the Heat Transfer between Gases and In-Cylinder Walls at Fired
and Motored Conditions of an IDI Diesel Engine, Appl. Therm. Eng., vol. 28, pp.
1395–1404, 2007.
10. G. Eichelberg, Some New Investigations on Old Combustion Engine Problems, Eng.,
vol. 148, pp. 463–446, pp. 547–560, 1939.
11. R. Stone, Introduction to Internal Combustion Engines, 3rd ed., pp. 471–499, Macmillan
Press Limited, Basingstoke, Hampshire, 1999.
12. K. Aydin, Investigation of the Sensitivity of Heat Transfer Calculation on the Results of
Combustion Analysis in Spark Ignition Engines,. 4th Int. Combustion Sym., July 19–21,
pp. 427–436, 1995 (text in Turkish).
13. T. Oguri, On the Coefficient of Heat Transfer between Gases and Cylinder Walls of the
Spark-Ignition Engine, Bulletin of JSME, vol. 3, no. 11, pp. 363–369, 1960.
14. G. Woschni and W. Spindler, Heat Transfer with Insulated Combustion Chamber Walls
and Its Influence on the Performance of Diesel Engines, Trans. of ASME––J. of Eng. for
Gas Turbines and Power, vol. 110, pp. 482–488, 1988.
15. G. Woschni, A Universally Applicable Equation for the Instantaneous Heat Transfer
Coefficient in the Internal Combustion Engine, SAE Paper 670931, 1967.
16. N. Watson and M. S. Janota, Turbocharging the Internal Combustion Engine, The
Macmillan Press, London, 1982.
17. P. J. Shayler, S. A. May, and T. Ma, The Determination of Heat Transfer from the
Combustion Chambers of SI Engines, SAE Paper 931131, 1993.
18. G. F. Hohenberg, Advanced Approaches for Heat Transfer Calculations, SAE Paper
790825, 1979.
19. P. Zeng, R. G. Prucka, Z. S. Filipi, and D. N. Assanis, Reconstructing Cylinder Pressure
of a Spark Ignition Engine for Heat Transfer and Heat Release Analyses, Proc. ASME,
ICE Division Fall Tech. Conf., Long Beach, CA, USA, pp. 1–11, 2004.
20. S. B. Han, Y. J. Chung, Y. J. Kwon, and S. Lee, Empirical Formula for Instantaneous
Heat Transfer Coefficient in Spark Ignition Engine, SAE Paper 972995, 1997.
21. C. F. Taylor and T. Y. Toong, Heat Transfer in Internal-Combustion Engines, ASME
Paper, 57-HT-17, 1957.
22. C. A. Finol and K. Robinson, Thermal Modeling of Modern Engines: A Review of
Empirical Correlations to Estimate the In-Cylinder Heat Transfer Coefficient, Proc.
IMechE, vol. 220, pp. 1765–1781, 2006.
23. W. W. Pulkrabek, Engineering Fundamentals of the Internal Combustion Engine, 2nd ed.,
pp. 360–409, Pearson Prentice-Hall, Upper Saddle River, NJ, 2004.
24. Y. Y. Wu, B. C. Chen, and F. C. Hsieh, Heat Transfer Model for Small-Scale Air-Cooled
Spark-Ignition Four Stroke Engines, Int. J. Heat Mass Transfer, vol. 49, pp. 3895–3905,
2006.
25. H. S. Soyhan, H. Yasar, H. Walmsley, B. Head, G. T. Kalghatgi, and C. Sorusbay,
Evaluation of Heat Transfer Correlations for HCCI Engine Modeling, Appl. Therm.
Eng., vol. 29, pp. 541–549, 2009.
View publication stats