CFD Modeling of a Turbo-charged Common-rail Diesel Engine

Transcription

1 JSAE / SAE 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 (C 10 H 22 ) 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

2 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 4M42-4AT2 Displacement 2977 c.c. Rated Power 92.0 kw/3200rpm Rated Torque 294 Nm/1700rpm Cylinders In-line 4 Cylinder/4 Stroke Compression Ratio 17:1 Fuel System Common Rail (Direct Injection) Injection Pressure 1600 bar max. Nozzle Hole Air Intake System Turbocharged with intercooler Emission Control System CAT, PCV, EGR Intake Valve Opening 13 btdc Intake Valve Closing 46 abdc Exhaust Valve Opening 60 bbdc Exhaust Valve Closing 12 atdc where V c 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 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. 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, l 4 is the distance between the pivot point (o) and the center of the roller, l 2 is the distance between the axis of the cam and the center of the roller, l 3 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 l 1 and the horizontal, and is the angle between link l 4 and the horizontal. Based on the geometry in Figure 7, the relationship between and can be found and the valve lift can be calculated. 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: 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

3 centers. The curves L2 and L4 are connected by two straight lines L1 and L3. L1 ( ) : L2 ( ) : L3 ( ) : Figure 8. Cam geometry ENGINE MODEL L4 ( ) : 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 7. Cam, link and valve geometry Figure 9. The profile and coordinate of piston and piston model Combining (2) to (7), the valve lift profile can be determined. Figure 10. Valve and valve model

4 Figure 11. Engine model Figure 13. Mesh NUMERICAL METHODS MESH Figure 12. Fluid model 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 and respectively. 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

5 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 show the temperature, fuel vapor C 10 H 22 mass fraction and O 2 mass fraction during the combustion stroke near top dead center. The simulation results reveal details of C 10 H 22 and O 2 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 C 10 H 22 and O 2 are consumed more than other areas. Table 2. Engine conditions for simulation Engine Speed 2000 rpm Intake Pressure bar Intake Temperature 301 K Initial Cylinder Pressure bar Initial Cylinder Temperature 484 K Initial Gas Composition O2 : 8.85% CO2 : 7.11% H2O : 2.626% Number of injector holes 7 Injector hole diameter mm Fuel flow rate of each hole kg/s Injection duration s Figure 15. Velocity field at 87 bbdc Figure 16. Tumble at 14.1 abdc Figure 14. Pressure field at 87 bbdc Figure 17. Temperature field at 2 atdc

6 Figure 18. C 10 H 22 mass fraction at 2 atdc Figure 20. The effects of the fuel injection timing on cylinder gas pressure (TDC is at 720 ) Figure 19. O 2 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 21. The effects of the fuel injection timing on cylinder gas temperature (TDC is at 720 ) Figure 22. In-cylinder temperature at 3 btdc

7 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 1. Harun M.I., H.K. Ng, S. Gan, Evaluation of CFD Sub-models for In-Cylinder Light-Duty Diesel Engine Simulation, Proceedings of ICEE rd International Conference on Energy and Environment,7-8 December 2009, Malacca, Malaysia. 2. K.M. Pang, H.K. Ng, S. Gan, Light-Duty Diesel Engine Modelling with Integrated Detailed Chemistry in 3-D CFD Study, Proceedings of ICEE rd 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 Automation. 4. Shang Yong, Liu Fu-shui, Li Xiang-rong, Numerical Simulation on Forced Swirl Combustion Chamber in Diesel Engine, 2010 International Conference on Digital Manufacturing & Automation. 5. Sage L. Kokjohn, Rolf D. Reitz, Investigation of the Roles of Flame Propagation, Turbulent Mixing, and Volumetric Heat Release in Conventional and Low Temperature Diesel Combustion, Journal of Engineering for Gas Turbines and Power-Transactions of the ASME Volume: 133, Gianluca D Errico, Tommaso Lucchini, Frank Atzler, Rossella Rotondi, Computational Fluid Dynamics Simulation of Diesel Engines with Sophisticated Injection Strategies for In-Cylinder Pollutant Controls, ENERGY & FUELS Volume: 26 Issue: 7 Pages: , 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 Flow Volume: 33 Issue: 1 Pages: 59-69, 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 Management Volume: 58 Pages: , B. Jayashankara, V. Ganesan, Effect of Fuel Injection Timing and Intake Pressure on the Performance of a DI Diesel Engine - A Parametric Study Using CFD, Energy Conversion and Management Volume: 51 Issue: 10 Pages: , Harun Mohamed Ismail, HoonKiatNg, SuyinGan, Evaluation of Non-Premixed Combustion and Fuel Spray Models for In-Cylinder Diesel Engine Simulation, Applied Energy 90 (2012) 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 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 m @mail.ntust.edu.tw Chia-Jui Chiang Assistant Professor National Taiwan University of Science and Technology, Taipei, Taiwan cjchiang@mail.ntust.edu.tw Yu-Hsuan Su Assistant Professor National Taiwan University of Science and Technology, Taipei, Taiwan ysu@mail.ntust.edu.tw Yong-Yuan Ku

8 Project Engineer Automotive Research and Testing Center, Taiwan 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 SOI EGR TDC computational fluid dynamics start of injection exhaust gas recirculation top dead center btdc before top dead center atdc after top dead center bbdc before bottom dead center abdc after bottom dead center