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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 SuNational Taiwan
University of Science and Technology, Taiwan
Yong-Yuan KuAutomotive 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 builtvia
the ANSYS FLUENT CFD solver to simulate thephenomenon during each
stroke. The initial conditions andboundary conditions are set based
on experimental dataobtained from a turbo-charged common-rail
diesel enginedeveloped by Mitsubishi. The variables that can be
observedfrom the CFD model include cylinder pressure, gas
velocity,
cylinder temperature, fuel particle tracks, and mass fractionof
cylinder gas components. The simulation results display
the effects of the fuel injection timings on the combustionheat
release process, cylinder pressure and cylindertemperature at
different engine operation conditions. Thepure 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 thespontaneous combustion.
The flow field obtained from
simulation results such as the tumbling motion can be usedto
explain the macroscopic phenomena observed fromexperiment results.
This research also discusses the effect offuel injection timing on
the cylinder pressure. Results showthat as the fuel injection
timing advances, the combustionphasing 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 sufferfrom
the undesired byproduct of pollutants such asparticulate matter
(PM), NOx, CO2, CO, and HC existing inthe combustion exhausts.
Recent progress in actuators suchas the common-rail fuel injection
system not only increasesthe engine power output but also reduces
the emission of
exhaust pollutants. In an effort to understand the mechanismof
the common-rail diesel engine combustion and to achieve
optimized engine performance and reduce emissions,
acomputational fluid dynamics (CFD) model of a dieselengine is
constructed in this paper. The CFD simulationallows us to
investigate the combustion process of dieselengine after
compression ignition and may cut down on thenumber 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 canbe
minimized. Furthermore, formation of pollutants [1] [2]can be
studied numerically. Simulation also offer thepossibility of
conducting fuel injection strategy research [3][4] effectively. The
combustion heat release process andflame 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 modelingis
used to estimate the formation of pollutants in thecombustion
process and to find the optimum injectionstrategy that reduces
exhaust pollutants. The CFD study in[7] shows that soot formation
can be suppressed viamulti-injection. Specifically, the start of
injection timing(SOI) and split-main ratio have strong effect on
the sootformation. Retarded SOI and higher split-main ratio lead
to
more soot pollutant. The CFD simulation results reveal thesoot
formation process and location clearly [8]. The author
in [9] studied the effects of injection timing and
intakepressure on direct injection diesel engines.
Simulationresults show that advanced injection timing
increasescylinder pressure, temperature, heat release rate, and
NOxemission. On the other hand, as the intake pressure isincreases,
the heat release process speeds up and soot
exhaust pollutant is decreased, while NOx emission isincreased
[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 fuelinjection trajectories and mass
fraction of gas componentsare obtained from the CFD simulation. The
effect of the fuelinjection timings on the combustion heat release
process is
investigated and validated against cylinder
pressuremeasurement.
ENGINE SPECIFICATIONS
The engine used in this study is a Mitsubishi
4M42-4AT2turbo-charged common-rail diesel engine as shown in
Figure
1. This engine is equipped with turbo charger, common-raildirect
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 theCFD
model are set based on the experimental data. Due tospace
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/3200rpmRated Torque 294 Nm/1700rpmCylinders
In-line 4 Cylinder/4 StrokeCompression Ratio 171Fuel System Common
Rail (Direct Injection)Injection Pressure 1600 bar max.Nozzle
HoleAir Intake System Turbocharged with intercooler
Emission Control System CAT, PCV, EGRIntake Valve Opening
Intake Valve Closing
Exhaust Valve Opening
Exhaust Valve Closing
Figure 1. MITSUBISHI 4M42-4AT2
CFD MODELING
In order to obtain a realistic model of the engine, a
spareengine was disassembled and parts (such as the piston headand
valves) were cut open to determine the exact dimensionsof various
parts.
CYLINDER DISPLACEMENT
The geometric configuration of the piston-crank drivingmechanism
of our engine is shown in Figure 6. With somesimple geometric
relationships, the cylinder volume V can
be expressed in terms of the rotating angle as follows:
where Vc is the clearance volume, B is the bore diameter,
is the ratio of the length of connecting rod l tocrank 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
thevalve 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,
l2isthe distance between the axis of the cam and the center ofthe
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 l1and
the horizontal, and is the angle between link l4and
thehorizontal. 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 intofour
segments L1, L2, L3, L4, respectively. The radius of L4is 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 straightlines
L1 and L3.
L1 ( ) :
L2 ( ) :
L3 ( ) :
L4 ( ) :
Figure 7. Cam, link and valve geometry
Combining (2) to (7), the valve lift profile can
bedetermined.
Figure 8. Cam geometry
ENGINE MODEL
The engine model is constructed using AUTOCAD based onthe
specifications of the MITSUBISHI diesel engine shownin Table 1 and
measurement of key engine components,such as the piston bowl and
valve stems shown in Figure 9and 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 simulationzone
shown in Figure 12. The complicated manifolds of theinlet 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 themodel
and the movement of each component. The meshingmethod used in this
study follows the works [10] [11] andANSYS FLUENT Guide [12]. As
shown in Figure 13,tetrahedrons method is used to the fill
irregular shape of thepiston crown. The fluid close to cylinder
head and in the
inlet and outlet ports adopts the sweeping method forapplication
of the Layering Mesh Method. The element sizes
are between 1 mm and 2.1 mm. And the number of nodesand elements
are 79181 and 207111 respectively.
Figure 13. Mesh
NUMERICAL METHODS
In order to simulate the flow field in the combustionchamber,
the standard k-epsilon model is adopted in thisstudy. The engine
intake model is pressure inlet. In thismodel, the intake pressure,
temperature and air compositionscan be determined. The plain
orifice atomizer and droplet
collision and breakup are used respectively in injection typeand
spray model to simulate the fuel injection. For the diesel
engine compression ignition, ignition delay model isselected in
the autoignition model. The minimum time step
is As the engine is equippedwith turbo charger and EGR, the
intake pressure and intakegas compositions are determined based on
the experimentaldata.
SIMULATION RESULTS
In the CFD simulation, the cylinder variables in each cycle
are calculated from crank angle 360 to 1080 thatincludes four
strokes such as intake, compression, expansionand exhaust
respectively. The combustion top dead center
(TDC) is located at 720 . Table 2 shows the initialconditions,
boundary conditions and nozzle specification setbased on the
experimental data obtained from the
MITSUBISHI 4M42-4AT2 diesel engine. In the followingfigures, the
left port is inlet and the right port is outlet. TheCFD 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 evaluationof
the engine design and control strategy. Figure 14 andFigure 15 show
the pressure and velocity fields inside the
combustion chamber during the intake stroke. Significantlystrong
tumbling motion can be clearly observed in Figure 16.Figures 17-19
show the temperature, fuel vapor C10H22massfraction and O2mass
fraction during the combustion strokenear top dead center. The
simulation results reveal details ofC10H22and O2 compositions
during the combustion process.
Figure 17 exhibits that the burning starts from the insidewall
of cylinder bowl. Consequently, at the correspondingareas in Figure
18 and Figure 19 it can be observed that the
C10H22and 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 87bBDC
Figure 15. Velocity field at 87bBDC
Figure 16. Tumble at 14.1aBDC
Figure 17. Temperature field at 2
aTDC
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Figure 18. C10H22mass fraction at 2aTDC
Figure 19. O2mass fraction at 2aTDC
The cylinder pressures at various injection timings show
inFigure 20 are validated against experimental measurement.
The cylinder pressures in Figure 20 result from the shift inthe
combustion phasing at different injection timings. Theretarded
injection of fuel leads to lower peak pressure as thecombustion is
initiated late in the expansion stroke. Shift intemperature at
different injection timings is also observed inFigure 21. The peak
temperature is also decrease as the
injection timing retards. The temperature is decreased beforethe
combustion takes place due to the fuel vaporization. This
phenomenon can be observed in Figure 22 and Figure 23 asthe
vaporization of fuel causes the temperature drop in thepiston
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 3bTDC
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Figure 23. Fuel vapor distributes at 3
bTDC
CONCLUSIONS
A CFD model for a turbo-charged common-rail dieselengine is
constructed in this paper. Physical variables suchas cylinder
pressure, gas velocity, cylinder temperature andmass fraction of
cylinder gas components can be easilyobtained from simulations. It
can be used to explain the
macroscopic phenomena such as the cylinder gastemperature
dropping due to the fuel vaporization. Thisresearch also discusses
the effect of fuel injection timing onthe cylinder pressure and
temperature. Results show that asthe fuel injection timing advances
and the combustionphasing advances, and the peak cylinder pressure
and thepeak cylinder temperature raise. In this paper, the initial
andboundary conditions are set based on the measurement ofthe
pressure, temperature and gas compositions. The
simulation results are then validated against cylinderpressure
measurement at various fuel injection timings.
REFERENCES
1. Harun M.I., H.K. Ng, S. Gan, Evaluation of CFDSub-models for
In-Cylinder Light-Duty Diesel EngineSimulation, Proceedings of ICEE
2009 3rdInternational Conference on Energy andEnvironment,7-8
December 2009, Malacca, Malaysia.
2.
K.M. Pang, H.K. Ng, S. Gan, Light-Duty DieselEngine Modelling
with Integrated Detailed Chemistry
in 3-D CFD Study, Proceedings of ICEE 2009 3rdInternational
Conference on Energy and Environment,7-8 December 2009, Malacca,
Malaysia.
3. Hu Mingjiang, Shen Chaoying, Research on PredictingFuel Spray
Characteristics of Diesel Engine, 2010International Conference on
Intelligent ComputationTechnology and Automation.
4. Shang Yong, Liu Fu-shui, Li Xiang-rong, NumericalSimulation
on Forced Swirl Combustion Chamber inDiesel Engine, 2010
International Conference onDigital Manufacturing &
Automation.
5.
Sage L. Kokjohn, Rolf D. Reitz, Investigation of theRoles of
Flame Propagation, Turbulent Mixing, andVolumetric Heat Release in
Conventional and Low
Temperature Diesel Combustion, Journal ofEngineering for Gas
Turbines and Power-Transactionsof the ASME Volume: 133, 2011.
6. Gianluca DErrico, Tommaso Lucchini, Frank Atzler,Rossella
Rotondi, Computational Fluid DynamicsSimulation of Diesel Engines
with SophisticatedInjection Strategies for In-Cylinder Pollutant
Controls,ENERGY & FUELS Volume: 26 Issue: 7 Pages:4212-4223,
2012.
7.
Raouf Mobasheri, Zhijun Peng, Seyed MostafaMirsalim, Analysis
the Effect of Advanced InjectionStrategies on Engine Performance
and Pollutant
Emissions in a Heavy Duty DI-Diesel Engineby CFD Modeling,
International Journal of Heat andFluid Flow Volume: 33 Issue: 1
Pages: 59-69, 2012.
8. Kar Mun Pang,, Hoon Kiat Ng, Suyin Gan, Simulationof Temporal
and Spatial Soot Evolution in anAutomotive Diesel Engine Using the
Moss-Brookes
Soot Model, Energy Conversion andManagement Volume: 58 Pages:
171-184, 2012.
9. B. Jayashankara, V. Ganesan, Effect of Fuel InjectionTiming
and Intake Pressure on the Performance of a DI
Diesel Engine - A Parametric Study Using CFD,Energy Conversion
andManagement Volume: 51 Issue: 10 Pages: 1835-1848, 2010.
10. Harun Mohamed Ismail, HoonKiatNg, SuyinGan,Evaluation of
Non-Premixed Combustion and FuelSpray Models for In-Cylinder Diesel
EngineSimulation,Applied Energy 90 (2012) 271279..
11.,,,, CFD
, 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 ChiangAssistant 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 fromBureau
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
