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High-Speed Flow and Combustion Visualization to Study the Effects of Charge Motion Control on Fuel Spray Development and Combustion Inside a Direct- Injection Spark-Ignition Engine 2011-01-1213 Published 04/12/2011 Mayank Mittal MSU College of Engineering David L.S. Hung, Guoming Zhu and Harold Schock Michigan State Univ. Copyright 2011 SAE International doi:10.4271/2011-01-1213 ABSTRACT An experimental study is performed to investigate the effects of charge motion control on in-cylinder fuel-air mixture preparation and combustion inside a direct-injection sparkignition engine with optical access to the cylinder. Highpressure production injector is used with fuel pressures of 5 and 10 MPa. Three different geometries of charge motion control (CMC) device are considered; two are expected to enhance the swirl motion inside the engine cylinder whereas the third one is expected to enhance the tumble motion. Experiments are performed at 1500 rpm engine speed with the variation in fuel injection timing, fuel pressure and the number of injections. It is found that swirl-type CMC devices significantly enhance the fuel-air mixing inside the engine cylinder with slower spray tip penetration than that of the baseline case without CMC device. Combustion images show that the flame growth is faster with CMC device compared to the similar case without CMC device. INTRODUCTION Improvement in fuel efficiency and reduction in exhaust emissions are the main goals behind the new developments in internal combustion engines (Mittal et al., 2010). The concept of direct-injection spark-ignition (DISI) engine has the potential to achieve such goals. In this technology, fuel is directly injected into the engine cylinder, which offers great flexibility to control the fuel injection timing, its duration and the number of injections. Note that the fuel-air mixture preparation in the combustion chamber is one of the key factors that influence the in-cylinder combustion characteristics and hence the engine performance (Hung et al., 2007). Therefore, optimizing the fuel-air mixture homogeneity is an important parameter for the engine designers. In general, a homogeneous fuel-air mixture is achieved by injecting the fuel during the intake stroke. In addition, the use of a charge motion control (CMC) device is an important factor that affects the flow (Mittal and Schock, 2010) and hence the fuel-air mixing and combustion inside the engine cylinder. It is expected that the CMC device imparts an angular momentum to the charge entering the engine cylinder. Several studies have been reported to investigate the influence of charge motion control on the engine performance. Clarke and Stein (1999) combined the variable valve timing with the charge motion control valve (CMCV). Variable valve timing was obtained using the dual equal variable camshaft timing (VCT) strategy. The combination of dual equal VCT with a CMCV allows an engine to be operated either at or near stoichiometric or at lean conditions, which allows the use of a NOx trap for the purpose of further reducing air pollutants. The authors indicated that the synergy between the CMCV and the dual equal VCT allows the fuel consumption to be less than the fuel consumption during lean operation at standard valve timing. This is due to the fact that CMCV increases the in-cylinder charge motion, and hence improves the combustion and the ability to handle the charge dilution, which occurs from increased levels of internal 1469

exhaust gas recirculation resulting from valve timing retard. Li et al. (2000) investigated the effects of swirl control valve on in-cylinder flow using a laser doppler anemometry technique. Mittal and Schock (2010) used molecular tagging velocimetry to study the influence of charge motion control on in-cylinder flow inside an internal combustion engine assembly. Kim et al. (2005) investigated the effects of injection timing and intake port flow control on fuel wetting inside the engine cylinder. They found that a tumble mixturemotion plate inside the intake port significantly reduced cylinder liner and piston top fuel wetting. This is because the use of the tumble mixture-motion plate provided more turbulence, which effectively enhanced the mixing during the intake process. Lee and Heywood (2006) studied the effects of CMCV on combustion characteristics and hydrocarbon emissions. The authors concluded that CMCV improved mixture preparation due to increased swirl and tumble intensities which enhanced fuel transport, distribution, and evaporation. CMCV in the closed condition allowed reduced fuel injection and retarded spark timing strategies that reduced hydrocarbon emissions significantly during the cold start due to greater fuel evaporation and faster burning rate. Overall, previous investigations show that a charge motion control device is an important factor that controls the combustion process, and hence, influences the engine performance (Mittal and Schock, 2010). However, to the best of authors' knowledge, visualization studies of the charge motion control device on in-cylinder fuel-air mixture preparation and combustion are not available. Therefore, an experimental study is performed to investigate the effects of charge motion control on in-cylinder fuel-air mixture preparation and combustion inside a direct-injection sparkignition engine. Experiments were performed at 1500 rpm engine speed with the variation in fuel injection timing, fuel pressure and the number of injections. In the following sections, a detail of experimental setup is first outlined, followed by the results of various tests performed. Finally, concluding remarks are summarized from this work. EXPERIMENTAL SETUP The engine used in the present work is a four-valve, two intake and two exhaust, 0.4 liter single-cylinder sparkignition engine. It has a bore diameter of 83 mm and stroke length of 73.9 mm. A flat-top piston with optical access is used. This provided a compression ratio of 9.75:1. Mittal et al. (2010) used a custom-designed piston in the same engine, which allowed the compression ratio of 13.5:1. The head accommodates a pressure transducer to record the in-cylinder pressure data. A view of the combustion chamber geometry showing intake and exhaust valves, direct-injector, spark plug and the pressure transducer is illustrated in Figure 1 (Mittal et al., 2010). It should be noted that in this paper 0 crank angle corresponds to the top dead center (TDC) of the compression, and therefore, 180 crank angle degrees (CAD) corresponds to the bottom dead center (BDC) of the intake, i.e. 180 BTDC (before top dead center). Different fuel injection timings (240, 210 and 180 BTDC) are considered with gasoline fuel. A high-pressure direct-injection (HPDI) 7-hole injector (Mittal et al., 2010) is used with the fuel pressures of 5 and 10 MPa. Figure 1. Optical engine combustion chamber SETUP FOR FUEL SPRAY VISUALIZATION Figure 2 shows the experimental rig used for spray visualization tests. The laser is introduced into the cylinder through the flat-top piston with optical access. A Mie scattering technique is used to visualize the liquid phase of the fuel dispersion inside the combustion chamber. A quartz cylinder is used to provide the optical access to the cylinder for high-speed imaging. The fuel spray was imaged with a Photron APX-RS non-intensified highspeed CMOS camera with a Nikon 105 mm AF micro lens. The camera was set to operate at 10 khz, which provided an image size of 512 512 pixels. At 1500 rpm engine speed, each frame corresponds to 0.9 crank angle degrees. A high repetition rate pulsed copper vapor laser, synchronized with the high-speed camera and the fuel injection timing logic, was used to illuminate the liquid fuel dispersion. For each test condition, the engine was first motored to reach the desired rpm, i.e. 1500 rpm. Once the engine was stabilized, a signal from the Cosworth engine controller was sent out to the fuel injector to trigger the start of injection at a specific crank angle position as well as to trigger the camera to start recording the specified number of images in consecutive cycles. The fuel injection duration at each test point is defined to achieve a stoichiometric air-fuel ratio based on gasoline. For each imaging test, five-injection-cycle spray images were recorded to visualize the fuel dispersion with 400 consecutive frames from each cycle. 1470

Downloaded from SAE International by Brought To You Michigan State Univ, Thursday, April 02, 2015 Figure 2. Experimental rig for in-cylinder fuel spray visualization SETUP FOR COMBUSTION VISUALIZATION Figure 3 shows the experimental rig for combustion visualization. Note that the quartz cylinder was replaced with the metal cylinder for the combustion tests. Experiments were performed at 1500 rpm engine speed with part-load condition (0.45 bar MAP). This part-load condition was selected due to optical limitations of the flattop piston. The effects of split (or dual) injection were also studied and compared with the corresponding cases of single injection by maintaining the same relative air-to-fuel ratio (λ), inverse of fuel-to-air equivalence ratio (φ). With the split injection, the second injection was 90 CAD apart from the first injection and the two pulse widths (of fuel injection) were kept the same. The combustion images were captured by Photron APX-RS highspeed camera (operated at 10 khz) through the optical piston. For each test condition, the engine was first motored to reach the desired rpm. Once the engine was stabilized, a signal from the Opal-RT engine controller was sent out to the fuel injector to trigger the start of injection at a specific crank angle position as well as to trigger the camera at the spark timing crank angle position to start recording the specified number of images in consecutive cycles. The fuel injection duration at each test point was selected to achieve the desired relative air-to-fuel ratio. For each imaging test, forty consecutive cycles were recorded to visualize the combustion process with 200 frames from each cycle. Due to optical engine limitations, the fuel supply was cut off as soon as the camera recorded the specified number of cycles. In-cylinder pressure was recorded with one degree of crank angle resolution that has been synchronized with the imaging signal. The Kistler piezoelectric pressure transducer was used with the measurement range varying from 0 to 250 bars. The averaged in-cylinder pressure data is then used to evaluate the engine performance. Mass fraction burned (MFB) and burn durations are determined using the well-known RassweilerWithrow method (Rassweiler and Withrow, 1938). A linear model for the polytropic index during the combustion process is used to evaluate the pressure change due to the volume change (Mittal et al., 2009). Figure 3. Experimental rig for combustion visualization 1471

CHARGE MOTION CONTROL DEVICE The charge motion control device was installed between the intake manifold and the intake port. The nature of CMC device influence depends on its geometry. Three different geometries of charge motion control device were considered (see Figure 4); two are expected to enhance the swirl motion inside the engine cylinder whereas the third one is expected to enhance the tumble motion. Each CMC device used in this study reduced the port cross sectional area by about 75%, i.e., the flow area about 25%. Experiments were performed with both the conditions: CMC device open (i.e. without CMC device) and CMC device closed (i.e. with CMC device). RESULTS AND DISCUSSION Results of fuel spray development; combustion visualization and in-cylinder pressure analyses are presented for a directinjection spark-ignition engine with both open (without) and closed (with) charge motion control devices. Three different geometric configurations of charge motion control device are considered. FUEL SPRAY VISUALIZATION Figure 5 shows the spray development of gasoline with open (left column) and closed CMC devices of all three configurations, i.e. swirl-types 1 (second column) and 2 (third column) and tumble-type (right column). In each case, highpressure direct-injection injector is used with 5 MPa of injection pressure at 1500 rpm engine speed. The start of injection (SOI) is at 240 crank angle degrees (or 240 BTDC). The size of each spray image shown in this paper is 512 512 pixels. The physical size of a pixel is about 0.19 mm. Note that the intake valves are located towards the left side of each spray image. The spray development at 223.8 BTDC shows that the spray tip penetration is faster with tumble-type CMC device compared to the baseline case with open CMC device. However, it is to be noticed that the spray tip penetration is slower with swirl-type CMC devices compared to both open and tumble-type CMC devices. Spray images at 216.6 BTDC clearly show that the spray tip penetration is even slower with swirl type-1 CMC device compared to the swirl type-2 CMC device. It is interesting to note that no significant difference is observed in spray development when the baseline case (with open CMC device) is compared with the tumble-type CMC device. However, the fuel dispersion is wider with swirl-type CMC devices. Also, note that the intensity values in these images (with swirltypes 1 and 2) are relatively low compared to the intensity values in spray images of both open and tumble-type CMC devices (see the presence of more liquid fuel towards the left side of the piston top with both open and tumble-type CMC devices). This clearly shows that the air-fuel mixing improves with swirl-type CMC devices with reduced piston top impingement. Figure 6 shows the spray development of gasoline with open and closed CMC devices of all three configurations. In each case, an HPDI injector is used with 5 MPa of injection pressure at 1500 rpm engine speed. The start of injection is at 180 BTDC. Similar to the results observed with SOI at 240 BTDC, the spray tip penetration is slower with swirl-type CMC devices than that of open and tumble-type CMC devices. Note that the fuel dispersion is wider with swirl- and tumble-type CMC devices than that of open CMC device. Therefore, it is expected that the air-fuel mixing improves with the charge motion control device. SPRAY TIP PENETRATION Figure 7 shows the effects of injection pressure (at 5 and 10 MPa) and the injection timing (at 240 and 180 BTDC with 5 MPa of injection pressure) on spray tip penetration with open CMC device. The penetration length is determined as the axial location of the spray tip from the injector tip. As expected, the spray tip penetration is faster with the injection pressure of 10 MPa than that of 5 MPa. Note that the spray tip penetration is slower when fuel injection starts at 180 BTDC than that of 240 BTDC injection timing due to upward movement of the piston. Figure 8 shows the effects of different configurations of charge motion control devices on spray tip penetration. As observed in spray images (of Figure 5), the spray tip penetration is faster with tumble-type CMC device compared to open and swirl-type CMC devices. Note that the spray tip penetration is slowest with swirl type-1 CMC device compared to open and other configurations (swirl type-2 and tumble-type) of CMC devices. Figure 4. Three different types of charge motion control devices: (a) Swirl type-1, (b) Swirl type-2 and (c) Tumble type CMC devices 1472

Downloaded from SAE International by Brought To You Michigan State Univ, Thursday, April 02, 2015 Figure 5. Spray development with (a) Open, (b) swirl-type 1, (c) swirl-type 2 and (d) tumble-type CMC devices at 5 MPa of injection pressure with SOI at 240 BTDC 1473

Downloaded from SAE International by Brought To You Michigan State Univ, Thursday, April 02, 2015 Figure 6. Spray development with (a) Open, (b) swirl-type 1, (c) swirl-type 2 and (d) tumble-type CMC devices at 5 MPa of injection pressure with SOI at 180 BTDC 1474

Figure 7. Effect of injection pressure and injection timing on spray tip penetration Figure 8. Effect of charge motion control device on spray tip penetration COMBUSTION VISUALIZATION AND IN-CYLINDER PRESSURE ANALYSES Combustion visualization and in-cylinder pressure analyses are presented with open and closed (swirl type-2) CMC devices. The characteristics displayed in the combustion images, such as the flame sizes, shapes and appearance, may provide useful insight into what happens over the combustion period (Aleiferis et al., 2008). It should be pointed out here that the images presented are a two-dimensional representation of the three-dimensional flame development inside the engine cylinder. Also, it is to be noticed that each combustion image shown in this paper is a reduced form of its original image size of 512 512 pixels to 420 420 pixels (for better visibility to the reader) by eliminating the dark area band of pixels outside the cylinder. The physical size of a pixel is about 0.22 mm. Figure 9 shows the stoichiometric combustion images of gasoline with single injection for both open and closed (swirl type-2) CMC devices at 25.2, 28.8, 31.5 and 34.2 after spark timing (AST). High-pressure direct-injection injector is used at 5 MPa of injection pressure. The images are enhanced so that the early flame development and its growth is clearly visible to the reader for comparison purpose. The spark timing (ST) was at 35 BTDC based on MBT. The MBT timing at each test point was determined based on the maximum value of the mean IMEP during the spark sweep. The engine was operated at 1500 rpm with part-load condition. In each case the start of injection was considered at 210 BTDC. Note that Mittal et al. (2010) showed less overall impingement on in-cylinder surfaces in the same engine at this injection timing, and due to this, injection 1475

Figure 9. Flame images of gasoline with single injection (λ =1 and ST = 35 BTDC) using HPDI injector at 5 MPa with open (upper) and closed (lower) CMC devices timing of 210 BTDC is selected. The intake valves in all the combustion images are located towards the upper half of the images. It is evident from the images that the flame growth is slower with open CMC device than that of closed CMC device. It is expected that there will be some cycle-to-cycle variations in the flame development. Figure 10 shows the combustion images of gasoline with split (or dual) injection for both open and closed CMC devices at 25.2, 28.8, 31.5 and 34.2 after the spark timing. In each case, the start of first injection was at 210 BTDC with injection pressure of 5 MPa. The start of second injection was at 120 BTDC (90 CADs apart from the first injection). Note that the total amount of fuel was divided equally in both the injections for stoichiometric air-to-fuel condition. The spark timing was at 32 BTDC based on MBT. The combustion images show that the flame growth is much faster with closed CMC device than that of open CMC device. It is to be noticed that some bright rich spots are also visible in the combustion images (more with open CMC device than that of closed CMC device) of split injection. This may be occurring due to droplet burning (Aleiferis et al., 2008). Early start of the second injection may help to reduce these hot spots by allowing more mixing time. Also, hot in-cylinder conditions of the metal engine may help to reduce these hot spots due to faster evaporation of liquid fuel inside the engine cylinder. Aleiferis et al. (2008) discussed that gasoline is particularly susceptible to these hot spots. Figure 11 shows the averaged in-cylinder pressures for gasoline at stoichiometric conditions for both open and closed CMC devices with both single and split injections. An HPDI injector at 5 MPa is used in each case. It can be observed that the peak in-cylinder pressure increases with the split injection than that of the corresponding case with single injection. Also, the peak in-cylinder pressure is slightly higher with the closed CMC device than that of open CMC device. It is noticed that the crank angle at which the peak in-cylinder pressure occurs is 2 CAD earlier for closed CMC device than that of its corresponding open CMC device case for single injections. Similarly, for open CMC device the peak incylinder pressure location is 2 CAD earlier with split injection than that of its corresponding case of single injection. The mean IMEPs are 2.59 and 2.73 bar with open CMC device for single and split injections, respectively. This shows that the mean IMEP increases with the split injection than that of its corresponding case with the single injection. The mean IMEPs with closed CMC device are 2.52 and 2.58 bar for single and split injections, respectively. This shows that the mean IMEP reduces with closed CMC device than that of its corresponding case with open CMC device. This is expected due to increased pumping power with the closed CMC device. 1476

Downloaded from SAE International by Brought To You Michigan State Univ, Thursday, April 02, 2015 Figure 10. Flame images of gasoline with split injection (λ =1 and ST = 32 BTDC) using HPDI injector at 5 MPa with open (upper) and closed (lower) CMC devices Figure 11. Averaged in-cylinder pressure for gasoline (λ =1) with open and closed CMC devices for both single and split injections at 5 MPa of injection pressure Figure 12. Mass fraction burned for gasoline (λ =1) with both open and closed CMC devices for single and split injections at 5 MPa of injection pressure 1477

Downloaded from SAE International by Brought To You Michigan State Univ, Thursday, April 02, 2015 Figure 13. Flame images of gasoline with single injection (λ =1 and ST = 35 BTDC) using HPDI injector at 10 MPa with open (upper) and closed (lower) CMC device Figure 12 shows the mass fraction burned curves calculated from the averaged in-cylinder pressure data shown in Fig. 11. It can be observed that the burning is faster with closed CMC device than that with open CMC device. Similarly, it is faster with split injection than that of single injection. The 10% burn locations for single injections are at 0 and 1 CAD for open and closed CMC devices, respectively. With split injections, the total burn durations (10% - 90%) are 24 and 27 CAD for closed and open CMC devices, respectively. Therefore, the total burn duration decreases with the closed CMC device more than with the open CMC device. Figure 13 shows the stoichiometric combustion images of gasoline with single injection for both open and closed CMC devices at 25.2, 28.8, 31.5 and 34.2 after the spark timing. An high-pressure direct-injection injector is used at 10 MPa of injection pressure. The spark timing was at 35 BTDC based on MBT. The engine was operated at 1500 rpm with part-load condition. In each case the start of injection was at 210 BTDC. It is evident from the images that the flame growth is much faster at higher injection pressure of 10 MPa than that of lower injection pressure of 5 MPa (see Figure 9 for comparison). Also, at this higher injection pressure of 10 MPa, some bright spots are visible with open CMC device compared to the combustion images with closed CMC device. 1478 Figure 14. Averaged in-cylinder pressure for gasoline (λ =1) with open and closed CMC devices (single injections) at 10 MPa of injection pressure Figure 14 shows the averaged in-cylinder pressures for gasoline at stoichiometric conditions for both open and closed CMC devices with single injection. An HPDI injector was used with 10 MPa pressure in each case. No significant difference is observed in peak in-cylinder pressure values at this higher injection pressure (of 10 MPa) with open and closed CMC devices. However, the crank angle at which the peak in-cylinder pressure occurs is 1 CAD earlier for closed CMC device than that of open CMC device. The peak incylinder pressure is about 14.8 bar for both cases. The mean IMEPs are 2.78 and 2.62 bar with open and closed CMC devices, respectively.

REFERENCES 1. Mittal, M., Hung, D.L.S., Zhu, G. and Schock, H.J., A Study of Fuel Impingement Analysis on In-Cylinder Surfaces in a Direct-Injection Spark-Ignition Engine with Gasoline and Ethanol-Gasoline Blended Fuels, SAE Technical Paper 2010-01-2153, 2010, doi:10.4271/2010-01-2153. 2. Hung, D.L.S., Zhu, G., Winkelman, J.R., Stuecken, T., Schock, H., and Fedewa, A., A High Speed Flow Visualization Study of Fuel Spray Pattern Effect on Mixture Formation in a Low Pressure Direct Injection Gasoline Engine, SAE Technical Paper 2007-01-1411, 2007, doi: 10.4271/2007-01-1411. Figure 15. Mass fraction burned for gasoline (λ =1) with open and closed CMC devices (single injections) at 10 MPa of injection pressure Figure 15 shows the mass fraction burned curves calculated from the averaged in-cylinder pressure data shown in Figure 14. As seen earlier, it can be observed that the burning is faster with closed CMC device than that of open CMC device at this higher injection pressure of 10 MPa. The 10% burn locations are at 2 and 3 CAD for open and closed CMC devices, respectively. The total burn durations (10% - 90%) are 23 and 26 CAD for closed and open CMC devices, respectively. Therefore, similar to the results at 5 MPa of injection pressure, the total burn duration decreases with closed CMC device than that of open CMC device. CONCLUSIONS An experimental study was performed to investigate the effects of charge motion control on in-cylinder fuel-air mixture preparation and combustion of a direct-injection spark-ignition engine. High-pressure production injector was used with fuel pressures of 5 and 10 MPa. Experiments were performed at 1500 rpm engine speed with the variation in fuel injection timing, fuel pressure and the number of injections. It is found that swirl-type charge motion control devices significantly enhance the fuel-air mixing inside the engine cylinder compared to the baseline case with open CMC device. In addition, the spray tip penetration is found to be slower with swirl-type CMC devices compared to the case with open CMC device. The results of combustion visualization show that the flame growth increases with the increased fuel injection pressure. The peak in-cylinder pressure also increases with the increased injection pressure. The effects of CMC device on flame growth are more significant at lower fuel injection pressure of 5 MPa than that of higher injection pressure of 10 MPa. Overall, it can be concluded that charge motion control is an effective way to enhance the fuel-air mixing and hence to improve the engine performance. 3. Mittal, M., and Schock, H.J., 2010, A study of cycle-tocycle variations and the influence of charge motion control on in-cylinder flow in an I.C. engine, ASME Journal of Fluids Engineering, 132(5), 051107, pp. 1-8. 4. Clarke, J. R., and Stein, R. A., 1999, Internal Combustion Engine With Variable Camshaft Timing, Charge Motion Control Valve, and Variable Air/Fuel Ratio, U.S. Patent No. 5,957,096. 5. Li, Y., Liu, S., Shi, S., and Xu, Z., Effect of the Swirl Control Valve on the In-Cylinder Air Motion in a Four-Valve SI Engine, SAE Technical Paper 2000-01-2058, 2000, doi: 10.4271/2000-01-2058. 6. Kim, H., Yoon, S., Xie, X. B., Lai, M. C., Quelhas, S., Boyd, R., Kumar, N., and Moran, C., Effects of Injection Timings and Intake Port Flow Control on the In-Cylinder Wetted Fuel Footprints During PFI Engine Startup Process, SAE Technical Paper 2005-01-2082, 2005, doi: 10.4271/2005-01-2082. 7. Lee, D. and Heywood, J. B., Effects of Charge Motion Control During Cold Start of SI Engines, SAE Technical Paper 2006-01-3399, 2006, doi:10.4271/2006-01-3399. 8. Rassweiler, G. M. and Withrow, L., Motion Pictures of Engine Flames Correlated with Pressure Cards, SAE Technical Paper 380139, 1938, doi:10.4271/380139. 9. Mittal, M., Zhu, G., and Schock, H.J., 2009, Fast mass fraction burned calculation using net pressure method for real-time applications, Proc. IMechE, Part D: J. Automobile Engineering, 223(3), pp. 389-394. 10. Aleiferis, P.G., Malcolm, J.S., Todd, A.R., Cairns, A., and Hoffmann, H., An Optical Study of Spray Development and Combustion of Ethanol, Iso-Octane and Gasoline Blends in a DISI Engine, SAE Technical Paper 2008-01-0073, 2008, doi:10.4271/2008-01-0073. 1479

CONTACT INFORMATION Author for correspondence: Mayank Mittal, PhD Department of Mechanical Engineering Michigan State University East Lansing, MI - 48824, USA mittalma@msu.edu ACKNOWLEDGMENTS This work was supported in part by the U.S. Department of Energy under Grant DE-FC26-07NT43275. DEFINITIONS/ABBREVIATIONS HPDI High-pressure direct-injection IMEP Indicated mean effective pressure MAP MFB RPM Manifold absolute pressure Mass fraction burned Revolutions per minute λ Relative air-to-fuel ratio SOI Start of injection φ Fuel-to-air equivalence ratio ST Spark timing 180 BTDC 180 crank angle degrees before TDC of compression 25.2 AST 25.2 crank angle degrees after spark timing TDC VCT Top dead center Variable camshaft timing BDC Bottom dead center BTDC Before top dead center CAD Crank angle degree CMC Charge motion control CMCV Charge motion control valve DI Direct-injection DISI Direct-injection spark-ignition 1480