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JSAE 20139103 / SAE 2013-32-9103
CFD Modeling of a Turbo-charged Common-rail Diesel Engine
Guan-Jhong Wang, Chia-Jui Chiang, Yu-Hsuan Su National Taiwan
University of Science and Technology, Taiwan
Yong-Yuan Ku Automotive Research and Testing Center, Taiwan
Copyright © 2013 SAE Japan and Copyright © 2013 SAE
International
ABSTRACT
In this study, a single cylinder diesel engine model is built
via the ANSYS FLUENT CFD solver to simulate the phenomenon during
each stroke. The initial conditions and boundary conditions are set
based on experimental data obtained from a turbo-charged
common-rail diesel engine developed by Mitsubishi. The variables
that can be observed from the CFD model include cylinder pressure,
gas velocity, cylinder temperature, fuel particle tracks, and mass
fraction of cylinder gas components. The simulation results display
the effects of the fuel injection timings on the combustion heat
release process, cylinder pressure and cylinder temperature at
different engine operation conditions. The pure diesel (C10H22) is
adopted in this simulation study. In the FLUENT setup, k – epsilon
is used in the viscous model, and the autoignition model is used to
simulate the spontaneous combustion. The flow field obtained from
simulation results such as the tumbling motion can be used to
explain the macroscopic phenomena observed from experiment results.
This research also discusses the effect of fuel injection timing on
the cylinder pressure. Results show that as the fuel injection
timing advances, the combustion phasing advances, resulting in
higher peak cylinder pressure and peak cylinder temperature.
INTRODUCTION
With the excellent fuel efficiency and high torque output,
diesel engines are particularly suitable for heavy-duty vehicle
applications. However, diesel engines also suffer from the
undesired byproduct of pollutants such as particulate matter (PM),
NOx, CO2, CO, and HC existing in the combustion exhausts. Recent
progress in actuators such as the common-rail fuel injection system
not only increases the engine power output but also reduces the
emission of exhaust pollutants. In an effort to understand the
mechanism of the common-rail diesel engine combustion and to
achieve optimized engine performance and reduce emissions, a
computational fluid dynamics (CFD) model of a diesel engine is
constructed in this paper. The CFD simulation allows us to
investigate the combustion process of diesel engine after
compression ignition and may cut down on the number of costly
in-situ experiments. In addition, more design parameters can be
varied in simulation, the design
cycle can be shortened, and product development cost can be
minimized. Furthermore, formation of pollutants [1] [2] can be
studied numerically. Simulation also offer the possibility of
conducting fuel injection strategy research [3] [4] effectively.
The combustion heat release process and flame propagation [5] can
be identified and studied.
CFD has been used to study the effect of different injection
strategy in diesel engines [6, 7, 8, 9]. In [6], CFD modeling is
used to estimate the formation of pollutants in the combustion
process and to find the optimum injection strategy that reduces
exhaust pollutants. The CFD study in [7] shows that soot formation
can be suppressed via multi-injection. Specifically, the start of
injection timing (SOI) and split-main ratio have strong effect on
the soot formation. Retarded SOI and higher split-main ratio lead
to more soot pollutant. The CFD simulation results reveal the soot
formation process and location clearly [8]. The author in [9]
studied the effects of injection timing and intake pressure on
direct injection diesel engines. Simulation results show that
advanced injection timing increases cylinder pressure, temperature,
heat release rate, and NOx emission. On the other hand, as the
intake pressure is increases, the heat release process speeds up
and soot exhaust pollutant is decreased, while NOx emission is
increased [9].
In this paper, a CFD diesel engine model is developed to study
the common rail diesel engine combustion. Specifically, the
characteristics of the gas velocity, in-cylinder pressure and
temperature, common-rail fuel injection trajectories and mass
fraction of gas components are obtained from the CFD simulation.
The effect of the fuel injection timings on the combustion heat
release process is investigated and validated against cylinder
pressure measurement.
ENGINE SPECIFICATIONS
The engine used in this study is a Mitsubishi 4M42-4AT2
turbo-charged common-rail diesel engine as shown in Figure 1. This
engine is equipped with turbo charger, common-rail direct injection
system and exhaust gas recirculation (EGR) system. The engine
specifications are summarized in Table
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1. The initial conditions and boundary conditions in the CFD
model are set based on the experimental data. Due to space
limitation, only the results at 2000rpm without EGR are show in
this paper.
Table 1. Engine specifications
Engine Type MITSUBISHI 4M42-4AT2Displacement 2977 c.c. Rated
Power 92.0 kW/3200rpm Rated Torque 294 Nm/1700rpm Cylinders In-line
4 Cylinder/4 StrokeCompression Ratio 17：1Fuel System Common Rail
(Direct Injection)Injection Pressure 1600 bar max. Nozzle Hole Air
Intake System Turbocharged with intercoolerEmission Control System
CAT, PCV, EGR Intake Valve Opening 13 bTDCIntake Valve Closing 46
aBDCExhaust Valve Opening 60 bBDCExhaust Valve Closing 12 aTDC
Figure 1. MITSUBISHI 4M42-4AT2
CFD MODELING
In order to obtain a realistic model of the engine, a spare
engine was disassembled and parts (such as the piston head and
valves) were cut open to determine the exact dimensions of various
parts.
CYLINDER DISPLACEMENT
The geometric configuration of the piston-crank driving
mechanism of our engine is shown in Figure 6. With some simple
geometric relationships, the cylinder volume V can be expressed in
terms of the rotating angle θ as follows:
where Vc is the clearance volume, B is the bore diameter,
is the ratio of the length of connecting rod l to crank radius
a.
Figure 6. Engine geometry
VALVE LIFT PROFILES
There are 4 four valves on the top of cylinder of our engine,
including 2 inlet and 2 outlet valves. The motion of the valve head
is controlled by the cam-link-valve assembly. Figure 8 shows the
cam, links and valve geometry. The parameter is the cam rotation
angle, l4 is the distance between the pivot point (o) and the
center of the roller, l2 is the distance between the axis of the
cam and the center of the roller, l3 is the distance between the
valve tip and the
pivot point (o), d is the radius of roller, is the angle between
radius r and d, is the angle between link l1 and
the horizontal, and is the angle between link l4 and the
horizontal. Based on the geometry in Figure 7, the relationship
between and can be found and the valve lift can be calculated.
Figure 8 shows the cam geometry in the polar coordinate. For
convenience, the profile of the cam is divided into four segments
L1, L2, L3, L4, respectively. The radius of L4 is s, the radius of
L2 is m, and n is the distance between two
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centers. The curves L2 and L4 are connected by two straight
lines L1 and L3.
L1 ( ) :
L2 ( ) :
L3 ( ) :
L4 ( ) :
Figure 7. Cam, link and valve geometry
Combining (2) to (7), the valve lift profile can be
determined.
Figure 8. Cam geometry
ENGINE MODEL
The engine model is constructed using AUTOCAD based on the
specifications of the MITSUBISHI diesel engine shown in Table 1 and
measurement of key engine components, such as the piston bowl and
valve stems shown in Figure 9 and Figure 10 respectively. The
combination of cylinder, piston and valves forms the engine model
shown in Figure 11. The gas filling volume forms the numerical
simulation zone shown in Figure 12. The complicated manifolds of
the inlet ports are not modeled at current stage.
Figure 9. The profile and coordinate of piston and piston
model
Figure 10. Valve and valve model
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Figure 11. Engine model
Figure 12. Fluid model
MESH
The choice of meshing methods depends on the shape of the model
and the movement of each component. The meshing method used in this
study follows the works [10] [11] and ANSYS FLUENT Guide [12]. As
shown in Figure 13, tetrahedrons method is used to the fill
irregular shape of the piston crown. The fluid close to cylinder
head and in the inlet and outlet ports adopts the sweeping method
for application of the Layering Mesh Method. The element sizes are
between 1 mm and 2.1 mm. And the number of nodes and elements are
79181 and 207111 respectively.
Figure 13. Mesh
NUMERICAL METHODS
In order to simulate the flow field in the combustion chamber,
the standard k-epsilon model is adopted in this study. The engine
intake model is pressure inlet. In this model, the intake pressure,
temperature and air compositions can be determined. The plain
orifice atomizer and droplet collision and breakup are used
respectively in injection type and spray model to simulate the fuel
injection. For the diesel engine compression ignition, ignition
delay model is selected in the autoignition model. The minimum time
step
is As the engine is equipped with turbo charger and EGR, the
intake pressure and intake gas compositions are determined based on
the experimental data.
SIMULATION RESULTS
In the CFD simulation, the cylinder variables in each cycle
are calculated from crank angle 360 to 1080 that includes four
strokes such as intake, compression, expansion and exhaust
respectively. The combustion top dead center
(TDC) is located at 720 . Table 2 shows the initial conditions,
boundary conditions and nozzle specification set based on the
experimental data obtained from the MITSUBISHI 4M42-4AT2 diesel
engine. In the following figures, the left port is inlet and the
right port is outlet. The CFD simulation provides the in-cylinder
pressure, velocity, temperature and gas components. In other words,
cylinder
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conditions during each stroke can be obtained for evaluation of
the engine design and control strategy. Figure 14 and Figure 15
show the pressure and velocity fields inside the combustion chamber
during the intake stroke. Significantly strong tumbling motion can
be clearly observed in Figure 16. Figures 17-19 show the
temperature, fuel vapor C10H22 mass fraction and O2 mass fraction
during the combustion stroke near top dead center. The simulation
results reveal details of C10H22 and O2 compositions during the
combustion process. Figure 17 exhibits that the burning starts from
the inside wall of cylinder bowl. Consequently, at the
corresponding areas in Figure 18 and Figure 19 it can be observed
that the C10H22 and O2 are consumed more than other areas.
Table 2. Engine conditions for simulation
Engine Speed 2000 rpm Intake Pressure 1.40026 bar Intake
Temperature 301 K Initial Cylinder Pressure 2.49326 bar Initial
Cylinder Temperature 484 K
O2 : 8.85% CO2 : 7.11% Initial Gas Composition
H2O : 2.626% Number of injector holes 7 Injector hole diameter
0.132 mm Fuel flow rate of each hole 0.000404344 kg/s Injection
duration 0.00055 s
Figure 14. Pressure field at 87∘bBDC
Figure 15. Velocity field at 87∘bBDC
Figure 16. Tumble at 14.1∘aBDC
Figure 17. Temperature field at 2∘aTDC
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Figure 18. C10H22 mass fraction at 2∘aTDC
Figure 19. O2 mass fraction at 2∘aTDC
The cylinder pressures at various injection timings show in
Figure 20 are validated against experimental measurement. The
cylinder pressures in Figure 20 result from the shift in the
combustion phasing at different injection timings. The retarded
injection of fuel leads to lower peak pressure as the combustion is
initiated late in the expansion stroke. Shift in temperature at
different injection timings is also observed in Figure 21. The peak
temperature is also decrease as the injection timing retards. The
temperature is decreased before the combustion takes place due to
the fuel vaporization. This phenomenon can be observed in Figure 22
and Figure 23 as the vaporization of fuel causes the temperature
drop in the piston crown.
Figure 20. The effects of the fuel injection timing on
cylinder gas pressure (TDC is at 720 )
Figure 21. The effects of the fuel injection timing on
cylinder gas temperature (TDC is at 720 )
Figure 22. In-cylinder temperature at 3∘bTDC
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Figure 23. Fuel vapor distributes at 3∘bTDC
CONCLUSIONS
A CFD model for a turbo-charged common-rail diesel engine is
constructed in this paper. Physical variables such as cylinder
pressure, gas velocity, cylinder temperature and mass fraction of
cylinder gas components can be easily obtained from simulations. It
can be used to explain the macroscopic phenomena such as the
cylinder gas temperature dropping due to the fuel vaporization.
This research also discusses the effect of fuel injection timing on
the cylinder pressure and temperature. Results show that as the
fuel injection timing advances and the combustion phasing advances,
and the peak cylinder pressure and the peak cylinder temperature
raise. In this paper, the initial and boundary conditions are set
based on the measurement of the pressure, temperature and gas
compositions. The simulation results are then validated against
cylinder pressure measurement at various fuel injection
timings.
REFERENCES
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for In-Cylinder Light-Duty Diesel Engine Simulation,” Proceedings
of ICEE 2009 3rd International Conference on Energy and
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Modelling with Integrated Detailed Chemistry in 3-D CFD Study,”
Proceedings of ICEE 2009 3rd International Conference on Energy and
Environment, 7-8 December 2009, Malacca, Malaysia.
3. Hu Mingjiang, Shen Chaoying, “Research on Predicting Fuel
Spray Characteristics of Diesel Engine,” 2010 International
Conference on Intelligent Computation Technology and
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4. Shang Yong, Liu Fu-shui, Li Xiang-rong, “Numerical Simulation
on Forced Swirl Combustion Chamber in Diesel Engine,” 2010
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Temperature Diesel Combustion,” Journal of Engineering for Gas
Turbines and Power-Transactions of the ASME Volume: 133, 2011.
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with Sophisticated Injection Strategies for In-Cylinder Pollutant
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7. Raouf Mobasheri, Zhijun Peng, Seyed Mostafa Mirsalim,
“Analysis the Effect of Advanced Injection Strategies on Engine
Performance and Pollutant Emissions in a Heavy Duty DI-Diesel
Engine by CFD Modeling,” International Journal of Heat and Fluid
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8. Kar Mun Pang,, Hoon Kiat Ng, Suyin Gan, “Simulation of
Temporal and Spatial Soot Evolution in an Automotive Diesel Engine
Using the Moss-Brookes Soot Model,” Energy Conversion and
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10. Harun Mohamed Ismail, HoonKiatNg, SuyinGan, “Evaluation of
Non-Premixed Combustion and Fuel Spray Models for In-Cylinder
Diesel Engine Simulation,”Applied Energy 90 (2012) 271–279..
11. 张志荣,冉景煜,张力,蒲舸,“内燃机缸内气体 CFD 瞬态分析中动态网格划分技术,” Journal of
Chongqing University (Natural Science Edition) Vol.28 No.11.
12. ANSYS FLUENT Guide, Copyright ○c 2009 by ANSYS, Inc.
CONTACT INFORMATION
Guan-Jhong Wang Graduate Student Research Assistant National
Taiwan University of Science and Technology, Taipei, Taiwan
[email protected]
Chia-Jui Chiang Assistant Professor National Taiwan University
of Science and Technology, Taipei, Taiwan
[email protected]
Yu-Hsuan Su Assistant Professor National Taiwan University of
Science and Technology, Taipei, Taiwan [email protected]
Yong-Yuan Ku
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Project Engineer Automotive Research and Testing Center, Taiwan
[email protected] ACKNOWLEDGMENTS
The authors would like to thank the funding support from Bureau
of Energy, Ministry of Economic Affairs, Taiwan, R.O.C., under
contract 102-D0107.
ABBREVIATIONS
CFD computational fluid dynamics
SOI start of injection
EGR exhaust gas recirculation
TDC top dead center
bTDC before top dead center
aTDC after top dead center
bBDC before bottom dead center
aBDC after bottom dead center
