Proceedings of ICEF2005

ASME Internal Combustion Engine Division 2005 Fall Technical Conference

ASME Internal Combustion Engine Division 2005 Fall Technical Conference
September 11-14, 2005, Ottawa, Canada
September 11 – 14, 2005, Ottawa, Canada

ICEF2005-1275
ICEF2005-1275

EFFECT OF COMBUSTION CHAMBER GEOMETRY ON EMISSIONS FROM A

SINGLE CYLINDER DIESEL ENGINE

Boggavarapu V.V.S.U. Prasad Ravikrishna R. V. 1

Indian Institute of Science, Indian Institute of Science,
Department of Mechanical Engineering, Department of Mechanical Engineering,
Bangalore – 560012, India. Bangalore – 560012, India.
Email: bprasad@mecheng.iisc.ernet.in Email: ravikris@mecheng.iisc.ernet.in

ABSTRACT INTRODUCTION

Many of the stationary power generation and agricultural Stringent norms are being imposed by Governments all
pumping applications in India utilize diesel engines. Recently, over the world on emissions from internal combustion engines.
as per Government regulations, these engines are required to The present study is to estimate the effect of combustion
satisfy stringent emissions norms. This forms the motivation chamber geometry on a single-cylinder, 948 CC, 10 HP diesel
for the present study on a stationary, direct-injection, single engine. In particular, the objective is to explore the potential of
cylinder, 10 HP diesel engine. The selected engine was not re-entrancy in the combustion chamber geometry to lower
satisfying the norms. The engine has a hemi- spherical piston emissions in this category of engines.
bowl and an injector with a finite sac volume. The combustion In diesel engines, towards the end of compression, almost
chamber was made re-entrant and the injector was replaced all the air in the cylinder is transferred into the combustion
with a sac-less injector. After these modifications, there is a chamber (generally piston bowl in DI Diesel Engines). This
significant change in emission levels. implies that any change in the geometry of combustion
To understand clearly the effect of the combustion chamber affects the flow field at the end of compression. Flow
chamber geometry on the emission levels, three-dimensional field at the end of compression is important because of events
computational fluid dynamics (CFD) simulations have been like spray and combustion which take place around this time.
performed for the complete suction and closed-valve part of the Swirl, squish and turbulence are identified as important
cycle. Comparisons of turbulent kinetic energy and swirl levels parameters which are to be kept in view while designing a
of old and new geometries were systematically conducted. In diesel engine. Swirl motion is very important as it significantly
contrary to the expected, that the swirl and turbulence levels affects combustion [1]. Proper shape of combustion chamber
are consistently less in the modified geometry than that of also improves turbulence at the end of compression which in
original geometry. A third combustion chamber was proposed turn reduces smoke emission [2].
and tested computationally. It was found that the in the Different types of combustion chambers were studied in
proposed combustion chamber swirl and turbulence levels are earlier experimental and computational studies in literature.
much higher than the baseline engine. Thus, the proposed Saito et al. [3] noted that the use of re-entrant chambers will
combustion chamber geometry shows significant potential for increase NOX but reduce smoke emission. As the reduction in
the engine to meet the prescribed norms. smoke is almost independent of injection timing, NO emissions
can be controlled by adjusting the injection timing. Larger in-
cylinder velocities were observed when re-entrant chambers
Keywords: Diesel engine, Emissions, Re-entrant chamber, were used instead of cylindrical bowls [4, 5]. In cylindrical
CFD. bowls, a large portion of the unburnt hydrocarbons will be
transported into the squish region during reverse squish. In re-
entrant chambers, this spill-out is delayed. Fuel that is
transported into the squish region is generally converted into
1
Corresponding author soot clouds [6].

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 Swirl-squish interaction is quite different for cylindrical
combustion chambers and re-entrant combustion chambers. In % of Fuel Power Emissions (g/hr)
re-entrant chambers, two torroidal vortices are generally Max. cons. (KW) HC CO NOX
observed at TDC of compression stroke, instead of one as in Load (g/hr)
the case of cylindrical bowls [7]. In re-entrant chambers, very 10 620 0.76 9.66 20.00 23.51
high velocities are observed near the bowl entrance. Also, peak 25 800 1.89 9.82 18.68 34.85
swirl will be higher. Turbulent kinetic energy peaks twice 50 1120 3.75 9.90 15.65 48.59
during compression stroke [8, 9, 10]. Hydrocarbon emissions 75 1460 5.56 10.13 13.67 63.34
from DI Diesel Engines are mainly due to fuel injected during 100 1850 7.30 11.36 58.71 75.45
the delay period which mixes at conditions beyond Table 2. Results obtained from performance and
flammability limits, and fuel effusing from nozzle sac at the emission tests.
end of injection with low velocities [11]. For a given intake and
spray configuration, there exists an optimum re-entrancy which Particulate matter for the test cycle = 1.399 g/KW/hr
produces minimum emissions. Thus optimization of geometry
is important to attain minimum emission levels [12]. As per the emission norms, different weighting factors are
Most of the studies in literature concerning re-entrant specified at different loads to calculate the final emission levels
chambers were conducted either on motored engines or on of the engine. The emission levels of the engine calculated
large fired engines. There are very few studies on the effect of using these weighting factors are required to be within the
combustion chamber geometries in medium and small engines. specified values. Table 3 lists the weighting factors specified by
The present study focuses on the effect of introducing re- Central Pollution Control Board (CPCB) as part of the drive
entrancy to the combustion chamber of a single cylinder 10 hp, cycle.
948 cc engine running at 1500 rpm. Experimental and
computational studies on this engine are presented in this paper. % of Max. load 10 25 50 75 100
Weighting factor (Wf) 0.1 0.3 0.3 0.25 0.05
Table 3. Weighting factors specified by CPCB norms.
EXPERIMENTAL INVESTIGATION
The emission levels are then calculated using these
The engine selected for the present study is a stationary weighting factors as given by Eq. (1) below:
direct-injection diesel engine. Specifications of the engine
selected are listed in Table 1.
∑ (wf . p )
Power output 10 hp/7.46 kW pl = Load
--- (1)
Speed 1500 rpm ∑ (wf .Power )
Load
Bore 102 mm
Stroke 112 mm
Stoke volume 948 CC where, Pl – Net Pollutant emission,
Nozzle 0.28 mm X 3 holes p – Pollutant quantity in gm/hr at a particular load, and
wf – Weighting factor.
Table 1. Engine specifications.
Emission levels calculated according to the above formula
The experimental study conducted on this engine includes
are listed in Table 4. In this table, the corresponding CPCB
performance and emission tests in which load and emission
emission norms are also listed. The engine was observed to
data have been measured [13]. Based on the specified drive
significantly exceed the prescribed norms. On inspection of the
cycle, the tests were conducted at 10%, 25%, 50%, 75% and
engine design to identify the reasons for the high emissions,
100% of the maximum load. The static injection timing of the
other than the injection pressure, two factors seemed to be
engine is 26o CA BTDC. Pressure at the engine inlet was
contributing to the high level of emissions. They are:
maintained at 100 kPa to ensure the repeatability of
experimental conditions.
1. Shape of the combustion chamber – hemispherical bowl
In these experiments, cylinder pressure, injection line
2. Finite sac volume in the injector
pressure and injector needle lift were monitored for a particular
cycle at each load. Emission data for unburnt Hydrocarbons
(HC), Carbon Monoxide (CO), Particulate Matter (PM) and
NOX were recorded after the engine had sufficiently warmed
up. Other quantities like fuel consumption, cooling water flow
rate and temperatures were also monitored at each load. Table 2
shows the test results of the selected engine.

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 Pollutant Present engine CPCB norm Margin in reducing the ignition delay. The re-entrant chambers are
(g/KW/hr) (g/KW/hr) (%) prone to increase the NOX emission due to better mixing and
HC 2.83 1.3 -117.94 combustion. Keeping this in view the injection timing is
CO 5.58 3.5 -59.38 delayed by 4o CA to counteract the effect of re-entrancy. The
NOX 13.34 9.2 -45.02 same compression ratio is maintained during the change by
PM 1.399 0.3 -249.75 ensuring that volume of the two bowls is the same.
Table 4. Pollutant levels of current engine and CPCB The suggested modifications were made and tests were
norms. repeated on the modified engine. Table 5 shows the emission
values of the engine after modifications and comparison with
The hemispherical bowl makes some part of the air the emission levels in the older configuration.
unavailable for combustion as shown in Fig. 1. This makes the
combustion inefficient which causes higher HC emissions. Pollutant Un- Modified Govt. Margin
Unavailability of air in spray region may lead to unburnt fuel modified engine norms
which later on gets converted into PM. engine (g/kW/hr) (g/kW/hr) (%)
(g/kW/hr)
HC 2.83 0.94 1.3 27.32
CO 5.58 5.81 3.5 -65.94
NOX 13.34 9.33 9.2 -1.40
PM 1.399 0.504 0.3 -68.00
Table 5. Comparison of emission data of modified
engine with unmodified engine and CPCB norms.

Figure 1. Engine geometry before modifications. From Table 5, it can be seen that there is a significant
reduction in HC emissions. It has come down by 66.8% and
The finite sac volume in the injector leads to injection of has a satisfactory margin when compared to CPCB norms. A
large size drops with very low velocities into the combustion large part of this reduction can be attributed to the replacement
chamber towards the end of injection. This fuel is subjected to of the injector. CO has slightly gone up by 4.12%. The delayed
thermal cracking and leads to high HC and PM formation. In injection could have brought the NOX emission down by 30%.
their study, Lakshminarayana et al. [11] found that about 0.12 Good mixing and combustion led to reduction in PM by 64%.
times the fuel in the sac gets converted into HC emissions. In an overall sense, the results after modification are favorable.
Based on these observations, some modifications were However, the emission levels are still above the norms. After
suggested to this engine. Figure 2 shows the modifications satisfying the minimum requirements, a margin of 15% to 20%
suggested. The combustion chamber was made a re-entrant is required to ensure that production engines satisfy the
chamber, and the injector replaced with a sac-less injector. The emission tests.
new injector has 0.26 mm holes, whereas the old one had 0.28
mm holes. It was expected that better mixing would take place To summarize the experimental study,
because of re-entrancy, which may help in reducing both HC • Use of sac-less injectors has helped in reducing HC
and NOX (by reducing the premixed phase of combustion) emissions.
emissions. Also, the use of sac-less injector helps greatly in • Re-entrant chambers are useful to increase air-fuel
reduction of HC emissions. Smaller injector holes also cause mixing which further reduces HC and PM emissions.
faster break up of spray and evaporation of fuel. This can help • Reduction in NOX can be attributed to delayed injection
and reduction in ignition delay.
• Reasons for increase in CO emission may be delayed
injection.
• Even though emission levels have decreased, there is
still room for improvement, especially in the case of
HC emissions.

For achieving further reduction in emissions, a trial and

error approach with respect to experimentation is obviously
very expensive and time-consuming. Hence, it was decided to
conduct numerical simulations to understand the reduction in
emissions which would then provide future guidelines in
Figure 2. Modifications suggested to the engine.
further modifying the engine to achieve the desired emission

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 levels. Thus, it was decided to conduct 3-D Computational estimating the turbulence during some parts of the cycle in
Fluid Dynamic (CFD) simulations of the in-cylinder processes. swirling engine flows. But because of its simplicity and
problems like convergence and computational time with regard
COMPUTATIONAL INVESTIGATION to other models, k-ε model has been chosen for the present
study. Also, in the present case, the objective is to compare
As mentioned above, numerical simulations are useful in various chamber geometries.
obtaining an estimate of the complete 3-D flow structure, swirl- For suction stroke simulation, the inlet port and cylinder
squish interaction, and turbulence distribution inside the were considered. However, for the closed valve part of the
cylinder. CFD also provides information on how fuel and air simulation, the inlet port has been removed to reduce the
are interacting and regions of pollutant formation. This computational effort. Figure 3 shows the computational grid
understanding can be used to further modify the engine to used for suction and closed valve portion of the cycle. Constant
further reduce the emission levels. temperature boundary conditions were specified for the walls.
CFD simulations are performed for suction, compression Inlet pressure during suction was taken from experiments.
and expansion strokes of both engine configurations (before
and after modifications). For this purpose, a specialized IC COMPUTATIONAL RESULTS AND DISCUSSION
Engine simulation software AVL-FIRE has been used. This
software has been validated by Tatschl et al. [14, 15, 18], The computational study is mainly focused on studying the
Wieser et al. [16], and Winklhofer et al. [17]. It has an inbuilt air motion in detail in order to understand in-cylinder flow
automatic hybrid mesh generator which generates a hexahedral- dynamics. Simulations were started at the beginning of suction
dominated hybrid mesh. After generating the mesh, the moving stroke (-360o CA). A valve overlap of 8o was neglected as
mesh generator prepares the mesh with moving boundaries maximum exhaust valve lift during this period is only 0.3 mm.
which is required by the solver for the analysis. The solver The inlet valve closes at -148o CA.
starts the solution after specifying different boundary
conditions, initial conditions, models and model constants. The 2.5
different models used in the current study are listed below: Old engine
2 New engine
1. Flow – SIMPLE algorithm was used to solve the
1.5
Swirl Number

compressible Navier-Stokes equations.

2. Turbulence – k-ε model with wall function method.
1

0.5

-0.5
-360 -330 -300 -270 -240 -210 -180 -150
Crank Angle
Figure 4. Variation of swirl number with crank angle
during suction stroke.

Variations of swirl number and Turbulent Kinetic Energy

(TKE) during suction are shown in Figs. 4 and 5 respectively.
From these results, it is observed that the bowl has a negligible
effect on the in-cylinder flow fields and TKE at the end of
suction, as observed by other researchers also [10, 19]. The
volumetric efficiency obtained from the computations was
86.9% and the corresponding experimental value is 82.72%.
Figure 3. Computational grid used for suction, and This indicates that the suction process has been simulated
closed valve part of cycle. sufficiently accurate.
In-cylinder flow variables at the end of suction are
To make a note about the k-ε model, it is a simple model transferred automatically to the computational mesh of
which assumes isotropic turbulence. For this reason, even compression stroke (excluding inlet port). Compression and
though it has given reasonable agreement with experiments, expansion are simulated from -148o CA up to 148o CA.
some researchers [10, 19, 20] have found it to be under Variation of mass-averaged swirl during compression and

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 expansion strokes is shown in Fig. 6. It can be observed that The advantage due to the re-entrancy in the engine combustion
the swirl number in the new geometry is consistently less till chamber may have been offset by the large central projection
30o ATDC. which has actually caused a reduction in swirl and TKE.
The velocity contours in a plane passing through the axis
250
of the cylinder at different crank angles around TDC is shown
Old engine
in Fig. 8. It is observed that the flow field is asymmetric in the
New engine
200 section. This asymmetry of flow was carried forward from the
suction stroke. Comparing the old and new engines, except at
-20o CA, the maximum velocity is higher for the old geometry.
TKE (m2/s2)

150
The entry diameter of the new bowl is less by about 5% when
compared to old geometry and a projection was introduced in
100
the middle of the bowl. Observing the flow field at TDC (0o
CA), it is seen that the high velocity region due to squish is
50 occurring almost at the same place in both the geometries. In
the new bowl, as the entry diameter is less, more penetration of
0 the squish flow is expected. But, the central projection seems to
-360 -330 -300 -270 -240 -210 -180 -150 obstruct the squish flow and reduce the penetration. Also, the
Crank Angle surface area of the new bowl is more by about 13% than the old
design leading to increased skin friction. Because of these two
Figure 5. Variation of TKE with crank angle during
factors, the advantages of introducing the re-entrancy may be
suction stroke.
negated.
The peak swirl in the modified engine is less by around 35
8.5%. Variation of Turbulent Kinetic Energy (TKE) during the
Old engine
closed valve part of the cycle is shown in Fig. 7. As observed in 30
New engine
literature [9, 10], a peak in the predicted TKE before TDC is
25
observed in this case also. However, a second peak after TDC
TKE (m2/s2)

as reported in some studies is not observed in this case. This 20

may be due to the shallow piston bowl which leads to weak
squish velocities. Also, it can be seen that the average TKE 15
values in the new engine are less than old engine throughout 10
the closed valve part of the cycle. Peak valve of TKE in the
5
3.5
0
Old engine
-150 -120 -90 -60 -30 0 30 60 90 120 150
3 New engine
Crank angle
Swirl number

2.5 Figure 7. Variation of TKE during closed valve part of

the cycle.
2
Figure 9 shows the TKE distribution in the central plane at
crank angles on either side of TDC. It is observed that the peak
1.5 the values of TKE are slightly higher up to TDC in case of new
geometry. Even though the peak TKE value is higher in new
1 geometry, it exists only in a small region. Since the high TKE
-150 -120 -90 -60 -30 0 30 60 90 120 150 value is restricted to only a small region, average values of
Crank angle TKE are still less than that of old geometry. The peak values
themselves are lesser in new engine than the old engine after
Figure 6. Variation of swirl number with crank angle TDC. In the old geometry, the sharp edge of the combustion
during the closed valve part of the cycle. chamber also contributes to higher TKE during reverse squish
period. The rounded edges in the new geometry may also be
new bowl is less by about 7% than that of the old engine. The one of the reasons for the lower TKE.
trends of swirl number and TKE are opposite to what was
expected due to change of geometry. From the simulation
results, it can be concluded that the reduction in the emissions
is caused mainly due to change in the injector characteristics.

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 comparison of swirl numbers in the the proposed geometry
with those of the previous two geometries. The swirl
intensification in this engine is much higher than that of the
previous two cases. Peak value is higher by around 20%
compared to the old engine. Figure 12 shows the variation of
average TKE in all the three geometries. It is seen that near
TDC the intensification of turbulence is quite substantial.

Figure 10. Proposed geometry to improve swirl and

TKE levels.

Figure 8. Velocity distribution in old and new 4

geometries. Old engine
3.5 New engine
From these simulations, it is clear that there is still a Proposed engine
substantial scope for reducing emissions by optimising the Swirl number 3
combustion chamber geometry. Keeping this in view and
evaluating designs from literature [10], a third geometry is 2.5
proposed with more re-entrancy and without the central
projection. The shape of the bowl is shown in Fig. 10. The 2
volume of the bowl was again chosen to be same as that of the
1.5
baseline engine to maintain the same compression ratio. Based
on the earlier discussion which indicated the lack of
1
dependence of the end-of-suction flow field and TKE
-150 -120 -90 -60 -30 0 30 60 90 120 150
distribution on bowl geometry, the calculation for the proposed
Crank angle
geometry is started at the end of suction and continued up to
Figure 11. Variation of swirl number with crank angle
during the closed valve part of the cycle.

35
Old engine
30
New engine
25 Proposed engine
TKE (m2/s2)

20

15

10

0
-150 -120 -90 -60 -30 0 30 60 90 120 150
Figure 9. TKE distribution in old and new geometries. Crank angle

the end of expansion stroke. The swirl variation with crank Figure 12. Variation of TKE during closed valve part
angle during compression and expansion strokes of proposed of the cycle.
engine is shown in Fig. 11. This figure also shows a

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 The effect of squish and reverse squish are also seen [3] Saito, T., Daisho, Y., Uchida N., and Ikeya, N., 1986,
clearly in this case because of more re-entrancy. The peak TKE “Effects of combustion chamber geometry on diesel
value is higher by about 35% when compared to that of the old combustion,” SAE Paper 861186.
geometry. Higher swirl and turbulence levels around TDC of [4] Zhang, L., Ueda T., Takatsuki T. and Yokota K., 1995,
compression help in mixing and efficient combustion of fuel. “A study of the effects of chamber geometries on flame
This may cause an increase in NOX emission which can be behavior in a DI Diesel Engine”, SAE Paper 952515.
reduced by appropriately delaying the injection. Thus, the [5] Bassoli, C., Biaggini, G., Bodritti, G. and Cornetti, G.
proposed combustion chamber geometry shows significant M., 1984, “Two dimensional combustion chamber analysis of
potential for the engine to meet the CPCB norms. Future work direct injection diesel engine”, SAE Paper 840228.
will focus on fabricating the piston bowl with the proposed [6] Ikegamio, M., Fukeda, M., Yoshihara, Y. and kaneko,
geometry and assessing the engine emissions. J., 1990, “Combustion chamber shape and presurised injection
in high speed direct-injection diesed engines,” SAE Paper
CONCLUSIONS 900440.
[7] Arcoumannis, C., Bicen, A. F. and Whitelaw, J. H.,
Computational investigations of the in-cylinder fluid 1993, “Squish and Swirl-Squish Interation in Motored Model
dynamics on a 10 HP, single-cylinder DI Diesel Engine have Engines”, ASME J. of Fluid Mechanics, 105, pp. 105-112.
been reported along with experimental data on emissions. In [8] Williams, T. J. and Tindal, M. J., 1980, “Gas flow
particular, emissions of NOX, PM, HC and CO have been studies in direct inection diesel entines with re-entrant
reported at various loads for the original engine configuration, combustion chambers,” SAE Paper 800027.
and also a modified version of the engine with as re-entrant [9] Kondoh, T., Fukumoto, A., Ohsawa, K. and Ohkubo,
chamber and sac-less injector. The emission levels in the Y., 1985, “An assessment of a multi-dimensional numerical
modified engine are observed to be significantly lower than method to predict the flow in Internal Combustion Engines”,
those of the original engine. SAE Paper 850500.
In order to understand the effect of re-entrancy on the [10] Bread, P., Mokaddem, K. and Baritaud, T., 1998,
reduction in emissions, detailed three-dimensional CFD “Measurement and Modeling of Flow-Field in DI Diesel
simulations of the in-cylinder processes were conducted. It is Engine: Effects of Piston Bowl Shape and Engine Speed”, SAE
observed that the swirl and TKE levels of the new engine are Paper 982587.
consistently less than that of the old engine. It appears that in [11] Lakshminarayanan, P. A., Nayak, N., Dingare, S. V.
the new design of the combustion chamber, the effect of re- and Dani, A. D., 2002, “Predicting Hydrocarbon Emissions
entrancy is negated by the large central projection in terms of From Direct Injection Diesel Engines”, Transactions of the
increased skin friction and reduced squish-induced TKE. In ASME, 124, pp. 708-716.
other words, the reduction in emissions is mainly due to the [12] Risi, A. D., Donareo, T. and Laforgia, D., 2003,
change in injector characteristics. A third combustion chamber “Optimisation of the combustion chamber of direct injection
is proposed and detailed numerical simulations were diesel engines”, SAE Paper 2003-01-1064.
performed. This design indicated much higher swirl and TKE [13] Narasimhan, M. V., 2003, Project Report “Project
values as compared to the baseline configuration and shows sponsored by IDEMA at MICO/IISc to meet CPCB Norms 2004
significant potential in further reducing emissions. in genset engines ≤ 19kW.”
[14] Tatschl, R., Pachler, K. and Winklhofer, E., 1998, “A
ACKNOWLEDGMENTS Comprehensive DI Diesel Combustion Model for
Multidimensional Engine Simulation”, Combustion in Internal
The authors would like to acknowledge the information Combustion Engines, COMODIA 98, Kyoto, Japan.
and data provided by Prof. M. V. Narasimhan, Mr. J. Kalvani [15] Tatschl, R., Wiesler, B., Alajbegovic A. and Kunsberg
and IDEMA. Sarre, C. v., 2000, “Advanced 3D Fluid Dynamic Simulation
for Diesel Engines”, Thiesel '2000 - Conference on:
REFERENCES Thermofluiddynamic Processes in Diesel Engines, Valencia,
Spain.
[1] Monaghan, M. L., and Perrifer, H. F., 1981, “Air [16] Wieser, K. J. and Ennemose, A., 2002, “3D-CFD
motion and its effects on diesel performance and emissions,” Diesel Combustion and accurate Heat Transfer Modeling for
SAE Paper 810255. Diesel Engines”, JSAE Annual Congress, Yokohama, Japan.
[2] Brandl, F., Reverencic, Cartellieri, W. and Dent, J. C., [17] Tatschl, R., Kunsberg Sarre, C. v., Alajbegovic, A. and
1979, “Turbulent air flow in the combustion bowl of a D I Winklhofer, E., 2000, “Cavitation and spray formation in
Diesel engine and its effect on engine performance,” SAE Diesel flow situations”, ILASS Europe 2000, Darmstadt.
Paper 790040.

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 [18] Tatschl, R., Bogensperger M. and Riediger, H., 1998,
“Multi-dimensional simulation of spray combustion and
pollutant formation in a medium speed marine diesel engine”,
FISITA World Automotive Congress, Paris.
[19] Payri, F., Benjas, J., Margot, X., Gil, A., 2004, “CFD
modeling of the in-cylinder flow in direct-injection Diesel
engines”, Computers & Fluids, 33, pp. 995-1021.
[20] Miles, P., Megerle, M., Hammer J., Nagel. Z., Reitz,
R.D., and Sick V., 2002, “Late-Cycle Turbulence Generation in
Swirl-Supported, Direct-Injection Diesel Engines”, SAE Paper
2002-01-0891.
[21] Kuensberg-Sarre, C. v., Kong, S. C., and Reitz, R. D.,
1999, “Modeling the Effects of Injector Nozzle Geometry on
Diesel Sprays”, SAE Paper 1999-01-0912, 1999.
[22] Nishimura, A. and Assanis, D. N., 2000, “A Model for
Primary Diesel Fuel Atomization Based on Cavitation Bubble
Collapse Energy”, ILASS2000, Pasadena.
[23] Heywood J. B., 1988, Internal Combustion Engine
Fundamentals, Mc GRAW-HILL Book Company, New York.
[24] Turns S. R., 2000, An Introduction to Combustion, Mc
GRAW-HILL Book Company, New York.

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