Title
Computational study of the influence of in-cylinder flow on spark ignition-controlled auto- ignition hybrid combustion in a gasoline engine
Author names and affiliations:
Xinyan Wang a, b
a State Key Laboratory of Engines, Tianjin University, P.R. China.
b Centre for Advanced Powertrain and Fuels, Brunel University London, United Kingdom, [email protected].
Hui Xie a
a State Key Laboratory of Engines, Tianjin University, P.R. China, [email protected].
Hua Zhao a, b
a State Key Laboratory of Engines, Tianjin University, P.R. China.
b Centre for Advanced Powertrain and Fuels, Brunel University London, United Kingdom, [email protected].
Corresponding author:
Hui Xie, State Key Laboratory of Engines, Tianjin University, P.R. China, Weijin Road 92, Nankai District, Tianjin, China, 300072. Tel: +86-22-27406842-8009, Mobil: +86- 13001344531, Fax: +86-22+27406842-8009. Email: [email protected].
Abstract
In this research, computational fluid dynamics (CFD) simulations were performed to investigate the effect of in-cylinder flow motion on the in-cylinder conditions and SI-CAI hybrid combustion. It is proved in this study that asymmetric intake valve events could be used to generate the swirl dominated flow motion. However, the macroscopic flows, such as the swirl and tumble, show very weak correlations with zone-to-zone conditions and hybrid combustion process. The detailed investigation on the in-cylinder zone-to-zone conditions indicates that the in-cylinder turbulent kinetic energy (TKE) level and the mean flow velocity (Vm) around the spark plug would directly affect the early flame propagation process, which in turn affect the subsequent auto-ignition process through changing the heat transfer between central burned gas and end-gas. In addition, the increased temperature inhomogeneity of the spherical zones caused by the in-cylinder flow motion would prolong the auto-ignition combustion. The structures of the flame front and auto-ignition sites also demonstrate the significant impact of in-cylinder motion on the combustion process. It is found that the combustion mode transition is very sensitive to the in-cylinder TKE, Vm, temperature and its inhomogeneity, indicating that these flow and thermal conditions could be used to optimise the hybrid combustion mode operation. It also proves the fluctuations of the in-cylinder flow and thermal conditions could be the reasons leading to significant cycle-to-cycle variations in SI-CAI hybrid combustion.
Keywords: computational fluid dynamics, hybrid combustion, in-cylinder flow, controlled auto- ignition, gasoline engine
1. Introduction
Controlled auto-ignition (CAI), also known as homogeneous charge compression ignition (HCCI), enables low temperature and high efficiency combustion in gasoline engines1. However, the narrow CAI operating range becomes the hindrance of its practical application2. The hybrid spark ignition (SI) and CAI combustion was developed to assist CAI and bridge transition between the SI mode and pure CAI mode2-6. Although the SI-CAI hybrid combustion shows the potential to enhance the control of CAI and extend the operating range, it was found that the hybrid combustion is sensitive to the in-cylinder conditions and shows significant cycle-to-cycle variations when applied to a practical engine3, 4. Several experimental and numerical studies were carried out to understand the mechanism of the high cycle-to-cycle variations lying behind SI-CAI hybrid combustion. The chemical kinetic modeling study performed by Havstad et.al.4 revealed the fluctuations of exhaust gas recirculation (EGR) level, average cylinder temperature, zone-to-zone temperature and intake pressure are mainly responsible to the combustion oscillations with hybrid combustion mode.
Olesky et.al.5 and Urushihara et.al.6 experimentally found the in-cylinder temperature is one of the most important factors to control the SACI behavior. Wang et.al.7 demonstrated the significant effect of temperature stratification on SI-CAI hybrid combustion process through engine experiments and detailed three dimensional computational fluid dynamics (3-D CFD) simulations. On the other hand, different methods are developed to realize the effective
control of the SI-CAI hybrid combustion and achieve stable combustion. The spark timing5, intake temperature6 and combination of internal and external EGR5, 8 are used to control the SI-CAI hybrid combustion.
Apart from the above methods, the in-cylinder flow is also a promising approach to control the combustion process, which has been previously studied and applied to the control of pure SI and CAI combustion. Long et.al.9 carried out the research on the fundamental interaction between flame and flow and found that the burning velocity increases by up to 10 times the laminar burning velocity in the presence of a highly rotating flow. Laget et.al.10 established the correlations between characteristics of in-cylinder charge motion and SI combustion evolution through the numerical studies and found the addition of a moderate swirl motion to the tumble motions can bring benefits to the combustion. Lee et.al.11 performed the optical study on the effects of tumble and swirl flows on the flame propagation in a gasoline engine and concluded that the combination of tumble and swirl was found to be more effective than pure tumble in effecting rapid and stable combustion under lean mixture conditions. In the case of CAI/HCCI combustion, Christensen et.al.12 experimentally investigated the effect of turbulence on HCCI combustion and found the increased turbulence would increase the HCCI combustion duration. Kong et.al.13 performed numerical study on the effects of geometry generated turbulence on HCCI combustion and found the main effect of turbulence is to affect the wall heat transfer, and hence to change the mixture temperature which, in turn, influences the ignition timing and combustion duration. In order to effectively utilize the in-cylinder flow/turbulence to improve combustion, the intake port geometries11, intake valves14 and piston shapes12, 13, 15 could be optimized to adjust the in-cylinder flow and turbulence. The development of variable valve system provides an effective method to control the in-cylinder flow motion and combustion process16-18. In addition, the asymmetric valve events for the two intake valves19, 20 and intake valve deactivation strategy21, 22 have attracted much attention for their potential to improving the in-cylinder flow and mixing process.
Previous studies have focused on the effect of flow motion on SI and CAI, only limited researches covered the effect on SI-CAI hybrid combustion. Persson et.al.21 experimentally found increasing the turbulence levels could increase the flame expansion speed at both high and low residual rates, while the auto-ignition is delayed by increased turbulence at high residual rates. Joelsson et.al.23 analyzed the effect of initial temperature and turbulence fields on the spark assisted compression ignition (SACI) using LES and found turbulence plays a significant role in the first stage SI flame propagation. Yoo et.al.24 performed two-dimensional direct numerical simulations (DNS) on the effect of turbulence on SACI and found the high flow turbulence significantly enhances the overall combustion by inducing many deflagration waves.
Although these fundamental studies have indicated the significant effect of turbulence on the hybrid combustion, there is still a gap between the in-cylinder flow motion (swirl and tumble) and the SI-CAI hybrid combustion process if different flow control strategies are expected to be applied to control the hybrid combustion. The aim of this study is to provide the fundamental knowledge of SI-CAI hybrid combustion characteristics with different in-cylinder flow motions controlled by the asymmetric intake valve events. In this study, the detailed and systematic investigation of the effect of in-cylinder flow motion on the in-cylinder conditions and combustion process would be performed using 3-D CFD simulations. The asymmetric
intake valve events, one of the promising strategies to control the in-cylinder flow motion, were adopted to generate the in-cylinder flow motion. The in-cylinder flow conditions would be analysed at both macroscopic and microscopic scales at first. Then the accompanying effect of the flow motions on thermal distribution would be investigated and the detailed analysis of SI and CAI combustion process presented. Finally, the correlations will be established to clarify the relationship between in-cylinder flow conditions, thermal conditions and SI-CAI hybrid combustion characteristics.
2. Methodology
2.1. Engine specification
In this study, experiments were carried out on a single cylinder gasoline engine. The engine specifications are shown in Table 1. The engine comprises a Ricardo Hydra single cylinder block and a specially designed cylinder head, which is equipped with a 4-variable valve actuation system (4VVAS) with BMW’s Vanos and Valvetronics on both the intake and exhaust camshafts. The 4VVAS enables the continuous adjustment of intake/exhaust valve lift and the valve timing. The stable SI-CAI hybrid combustion was achieved by utilizing both external EGR and internal EGR achieved by the NVO strategy. In this study, an operation condition with symmetric intake valve events was selected to validate the applicability of the simulation results. The valve parameters for the 4VVAS, including intake/exhaust valve opening timing (IVO/EVO), closing timing (IVC/EVC) and lift (IL/EL), are presented in Table 2.
The other experimental detail could be found in previous works3, 7. 2.2. Numerical models
In this study, a recently developed SI-CAI hybrid combustion engine CFD programme25 was applied to study the effect of in-cylinder flow on the in-cylinder conditions and SI-CAI hybrid combustion process. In order to cover both turbulent mixing effects and chemical kinetics in the SI-CAI hybrid combustion, a set of models for the premixed flame propagation and auto- ignition combustion was constructed to model the hybrid combustion process and implemented in the STAR-CD TM software. The ECFM3Z combustion model26, which uses a flame surface density equation to describe the flame propagation process, was adopted as the framework of the hybrid combustion model. With the ECFM3Z model, the gas state is represented by three mixing zones: a pure fuel zone, a pure air plus possible residual gas zone and a mixed zone. The auto-ignition of unburned gas which is separated by the flame front in the mixed zone could be predicted by an auto-ignition model. The tabulated chemistry approach27 was applied to predict and capture the auto-ignition process in the hybrid combustion using the auto-ignition tendency estimation. A reduced surrogate mechanism for gasoline28 comprising four components (iso-alkanes, n-alkanes, aromatics and olefins) with 312 species was selected for the chemical kinetics calculations to construct the tabulated database under various thermodynamic conditions. Based on these models, the reaction regime of each cell is determined by the flame surface density and the auto-ignition tendency.
Specifically, if the flame surface density of a certain cell is greater than 0, flame would consume the available fuel in the cell according to the ECFM3Z model. By contrast, if the
auto-ignition tendency of a certain cell has achieved 1, the available fuel in the cell will undergo auto-ignition combustion according to the tabulated chemistry approach. Therefore, the SI-CAI hybrid combustion model could identify the reaction regime of each cell and enable the simulation of the transition from pure flame propagation to auto-ignition under SI-CAI hybrid combustion mode.
Reynolds-averaged Navier Stokes (RANS) approach was applied with renormalization group (RNG) k-ε turbulence model in the simulations. Pressure-implicit with splitting of operators (PISO) algorithm was used to solve the equations. The heat transfer was implemented through the general form of the enthalpy conservation equation for the fluid mixture29. The Angelberger wall function30 was used for the simulation of the wall heat transfer. The moving mesh with over 300,000 grids of 1 mm averaged grid size, as shown in Figure 1, was generated based on the single cylinder engine geometry.
2.3 Computational conditions
In this study, the asymmetric intake valve events were adopted to generate the specific in- cylinder flow motion. The baseline case F0 with the symmetric intake valve events was selected from experimental data and used to validate the simulation model. The schematic of the intake valve events for the investigated cases is shown in Figure 2. The IVO/IVC timings and exhaust valve parameters were kept the same for all cases, as shown in Table 2. While the intake lift values for two intake valves were different, as shown in Table 3. The asymmetric coefficient (C) is defined to show the degree of asymmetry of the intake valve lifts. It should be noted the reversed cases beginning with letter “R” (R2, R3, R4) just reversed the two valve lifts of the forward cases beginning with letter “F” (F2, F3, F4). With the current valve setups, the in-cylinder mass is almost the same for all cases, which could eliminate its effect on combustion characteristic. Because the fuel injection type was the port fuel injection in this study, the intake mixture was set as the homogeneous stoichiometric fuel/air mixture.
The intake boundary conditions for CFD simulations were estimated by the corresponding one-dimensional simulations of the baseline case (F0) based on GT-power. Specifically, the intake pressure and intake temperature were set as 95 kPa and 320K for all cases. The wall temperature for cylinder head, piston top and cylinder liner were 420K, 482K and 377K respectively. The CFD simulations were carried out from the exhaust stroke of last cycle to the combustion stroke of the current cycle.
In order to illustrate the applicability of the CFD engine simulation model, the predicted and measured in-cylinder pressure and heat release rate profiles are plotted in Figure 3. As it can be seen that the hybrid combustion engine simulation model could reproduce the SI-CAI hybrid combustion process and shows good agreement with the experimental data.
3. Results and discussions
3.1. In-cylinder flow conditions
Figure 4 compares the in-cylinder flow streamlines at 640 ºCA ATDC after the closing of intake valves. 22 “seeds” distributed under intake and exhaust valves are selected to generate the streamlines. The clockwise tumble motion along the cylinder head could be
faintly observed at the first row in Figure 4 and becomes chaotic at the second row. For the baseline case F0, a very weak clockwise swirl could be observed. However, the apparent anti-clockwise swirl flow is observed for the forward cases with asymmetric intake valve events, and the swirl flow is gradually enhanced from F1 to F5. On the other hand, the obvious clockwise swirl flow is observed in the reversed cases and gradually strengthened from R2 to R4. The vortex center moves from the top right corner in F1 to the middle between intake and exhaust valves in F3, and finally to the center of the cylinder in F5. The vortex centers for the reversed cases locate at the +y direction and are almost symmetric with that in the corresponding forward cases along x axis. It can be concluded from all these results that the in-cylinder flow is spinning helically upwards from the piston top to form an obvious swirl flow near the cylinder head. The gradually increased velocity magnitude demonstrates the asymmetric intake valve events significantly enhance the in-cylinder flow motion characterized with the obvious swirl motion. As expected, the flow velocity becomes higher further out from the swirl center. The change of the swirl center in different cases causes the velocity distribution to vary from case to case, which would affect the later combustion process.
In order to quantitatively investigate in-cylinder macroscopic flow conditions, swirl ratio (SR), tumble ratio (TR) and cross tumble ratio (CTR), as defined in Figure 1, are calculated from 400 ºCA to 670 ºCA ATDC. The equation used to calculate the swirl ratio is given as following31:
2
( ) ( ) ( ) ( ) ( ) 2 ( ) ( ) ( )
60
m
i i i i
i m
i i i
i
v r V
SR n r V (1)
where n is engine speed, the crank angle, i the cell number, Vi( )
the cell volume,
i( )
the cell density, vi( )
and ri( )
are the tangential velocity and radius respectively in the cylindrical coordinate with Z axis as the swirl axis.
With the same method, the swirl axis along with the cylindrical coordinate system in Equation (1) is then replaced as the tumble/cross tumble axis to evaluate TR/CTR. The tumble/cross tumble axis is parallel to y/x axis and crosses the central point between maximum and minimum z value of the combustion chamber.
Figure 5 shows the evolution of SR, TR and CTR. It should be noted that the positive values represent the motion in the clockwise direction as shown in Figure 1. It can be seen that the SR for the forward cases gradually increases after the intake valve opening and then decreases after reaching the peak value. The SR for the reversed cases seems to show the opposite trend as shown in Figure 5. With increasing asymmetry coefficient, the peak swirl ratio (SR) changes significantly from -1.8 in R4 to 2.5 in F5. The change of peak cross tumble ratio (CTR) is also very large from -3.6 in F5 to 2.9 in R4. The evolutions of TR for both forward cases and reversed cases show a similar trend of initial decline followed by a rise afterwards. The peak tumble ratio (TR) shows no distinct trend among all cases and varies slightly between -2.2 and -2.8. After the closing of the intake valves, the swirl motion decays slowly, leading to apparent differences among cases at 670 °CA ATDC just before spark
discharge. In addition, Figure 4 shows that the swirl structure is very obvious for the cases with asymmetric intake valve events. However, the tumble and cross tumble motion decays significantly, leading to similar flow fields at 670 °CA ATDC in all cases. The significant decomposition of the tumble motion also explains the chaotic flow field from +y view in Figure 4.
In theory, the in-cylinder flow conditions just before spark discharge are more important and influential to the later combustion process. Figure 6 directly compares the SR, TR and CTR for all cases at 670 °CA ATDC just before spark discharge. The reversed cases are distinguished from the forward cases with the hollow symbols in the figure. As shown in Figure 6, the swirl motion still prevails after the intake valve closing and ranges from -1.3 in R4 to 1.7 in F5 at 670 °CA ATDC. The absolute value of SR increases monotonously from case F0 to F5, which is expected to enhance the early flame propagation. The absolute values of TR show a trend of decreasing from F0 to F5 and range between 0.06 and 0.61 among cases. Although the CTR reaches considerable values during the intake process, it decays dramatically and fluctuates around 0 just before spark discharge.
Apart from the estimation of the macroscopic flow conditions using SR, TR and CTR, the investigation of the zone-to-zone flow conditions is also involved in this study. Seven concentric zones along the cylinder axis are defined as shown in Figure 7 to investigate the zone-to-zone flow conditions. It should be noted that the spark plug locates in Zone 1.
Figure 8 gives the average velocity magnitude (Vm) for the whole combustion chamber, Zone 1 and Zone 7 at 670 °CA ATDC. It should be noted that only the representative Zone 1 and Zone 7 are compared in this section for simplicity. The Vm in Zone 1 is obviously smaller than that in Zone 7. This could be demonstrated from Figure 4 that the swirl center with slower velocity is closer to Zone 1. The average Vm for the whole combustion chamber lies between Zone 1 and Zone 7. More importantly, it is found the Vm for Zone 1 does not increase monotonously as that for the whole chamber and Zone 7, but falls dramatically in F3 (R3) and then increases again. The reason accounting for this could be found in Figure 4 in that the swirl structure in F3 (R3) creates a lower velocity zone (dark color of the streamlines) just around the central area (Zone 1). With the swirl center continuous moving into the central area after case F3, the spark plug gradually moves out of the lower velocity zone and exposes to the higher Vm in Zone 1 for later cases (F4, R4, and F5). It again proves that the swirl ratio and the location of swirl center codetermine the in-cylinder flow velocity and its distribution in the current studied cases mainly dominated by swirl flow motion. Therefore, although the average velocity magnitude for the whole combustion chamber increases from case F0 to F5, the velocity in Zone 1, where the spark plug resides, changes non- monotonously.
The turbulence kinetic energy (TKE) also plays an important role in the flame propagation process. Figure 9 shows the average TKE for the whole combustion chamber at 670 °CA ATDC. Similar to Vm, the TKE increases at first, and suddenly drops in F3 (R3) before it increases again. It is found that TKE in Zone 7 is significantly lower than that in Zone 1 and shows tiny differences between different cases. Therefore, the average TKE for the whole combustion chamber shows similar trend to that in Zone 1. It seems that there is no direct link between a specific macroscopic flow condition (SR, TR or CTR) and the in-cylinder turbulence level, especially the zone-to-zone flow conditions.
3.2. In-cylinder thermal conditions
The in-cylinder flow motion would affect the mixing process of the hot in-cylinder residual gas and cold fresh mixture. Meanwhile, the flow also changes the heat transfer process between the in-cylinder mixture and chamber wall. This would finally lead to considerably impact on in- cylinder thermal conditions which also affect the combustion process. In this section, the in- cylinder thermal condition is analysed so that the effect of in-cylinder flow on the SI-CAI hybrid combustion can be understood.
Figure 10 plots the in-cylinder temperature distributions of different cases. It is noted that both the position and area of the high temperature zone vary from case to case. It is also found that the temperature of the central zone where the spark ignition will take place changes significantly among cases. As shown in Figure 11 the average temperature of the whole cylinder charge decreases slightly from 572 K in F0 to 569 K in F5. The reason could be attributed to the increased heat transfer to the chamber wall caused by the enhanced macroscopic flow motion. The increased heat transfer also results in the gradually reduced area of the high temperature zone with increasing valve lifts asymmetry, as shown in Figure 10. The average temperature of Zone 1 shows an obvious drop from 572 K in F0 to 567 K in F5 and is generally higher than that of Zone 7 near the wall. The average temperature of Zone 7 among different cases exhibits a negligible fluctuation within 3 K. The reduction in Zone 1 temperature is mainly caused by the weak mixing of the cold fresh intake mixture with the hot residual gas in the central area. It can be seen in Figure 12 that the residual gas fraction (RGF) of the central area (Zones 1 and 2) at 670 °CA ATDC gradually decreases from F0 to F5. The swirl flow inhibits the hot residual gas coming into the central area with the increasing swirl ratio.
Temperature inhomogeneity (Ti) in Figure 13 is determined as the standard deviation over all the cells in the cylinder at 670 °CA ATDC. Therefore, a smaller inhomogeneity indicates a greater mixing effect. It can be seen in Figure 13 that the temperature for the whole combustion chamber tends to be homogeneous from F0 to F2 (R2) and then becomes inhomogeneous again. The zone-to-zone inhomogeneity shows different trends. In Zone 1, the Ti is significantly lower before case F3, indicating better mixing in central area for case F0 to F2 (R2). In comparison, the Ti for Zone 7 in F0 is 13.7, significantly higher than the cases with the asymmetric valve lifts whose inhomogeneity fluctuates between 9.5 and 11.2.
According to the effect of temperature inhomogeneity on auto-ignition31, the temperature inhomogeneity variation in Zone 7 would affect auto-ignition process. Generally, the zone-to- zone temperature inhomogeneity variation gradually reduces to the average level after Case F3, indicating relatively homogeneous in-cylinder thermal distribution. However, the monotonous change of average temperature inhomogeneity among cases confirms that although the asymmetric valves could generate regular macroscopic flow (i.e. increased swirl ratio), the mixing between intake fresh air and in-cylinder residual gas is not as homogeneous as expected. It also proves the moderate swirl flow combined with the tumble motion is preferable to achieve the most effective mixing effect11.
3.3. SI-CAI hybrid combustion process
The previous sections have investigated the in-cylinder flow and thermal conditions among
cases. In this section, the effect of in-cylinder flow and thermal conditions on the two-stage hybrid combustion process is discussed.
Figure 14 shows the crank angles, CA10, CA50 and CA90, corresponding to 10%, 50% and 90% of total mass fraction burned (MFB). The crank angle corresponding to the mode transition (CAT) from SI to CAI and the corresponding MFB are also marked in the figure. In this study, CAT is defined as the crank angle when over 2% of the cells auto-ignites. It is found that the combustion characteristics also change non-monotonously. The combustion phasing (CA50) advances from 726 °CA in F0 to 723 °CA ATDC in R2. After that CA50 in F3 suddenly delays to 727 °CA and gradually advances again. It could also be found the transition crank angle (CAT) correlates well with CA50 in this study and shows a slightly delay after CA50 with the asymmetric valve events. The MFB at CAT increases from F0 to F5. This is because the occurrence of auto-ignition needs more energy from flame propagation to compensate for the heat loss to the chamber wall for the case with asymmetric valve events.
It could be deduced from this angle that the in-cylinder flow has opposite effects on SI flame and CAI combustion. On the one hand, the enhanced flow would speed up the turbulent flame velocity. On the other hand, the flow would also increase the heat loss and inhibit the auto- ignition to some extent.
In order to investigate the impact of in-cylinder flow on the two-stage combustion process separately, two parameters D1 (durations from CA10 to CAT) and D2 (durations from CAT to CA90) are selected here to characterize the flame propagation dominated combustion process and auto-ignition dominated combustion process. Figure 15 shows that the flame dominated combustion duration D1 is significantly longer than the auto-ignition dominated combustion duration D2 for all cases. Both D1 and D2 decrease at first and suddenly increase in case F3. After that, they gradually decrease again. Therefore, the impact of the in-cylinder flow is effective throughout the whole combustion process. Certainly, the development of the early flame propagation process would play an important role on the later auto-ignition process. However, the effect of flow on end-gas condition still could not be neglected, which is demonstrated from the combustion duration shown in case F5. It is found that the flame dominated combustion duration D1 is reduced from 13.2 °CA in F4 to 13 °CA in F5. However, the auto-ignition dominated combustion duration D2 shows opposite trend, increasing from 4.7 °CA in F4 to 5 °CA in F5. This opposite trend of D1 and D2 in F5 could be attributed to the relatively higher temperature inhomogeneity of the end-gas, i.e. the Ti of Zone 7 in F5 as shown Figure 13.
From the analysis of the combustion characteristics, it is confirmed that the in-cylinder flow shows significant effects on both SI and CAI combustion process. Figure 16 shows the section views of the flame structures at CA10 for all cases. In order to estimate the flame front structure, the 80% value of the maximum flame average density is set as the low criteria of the displaying scale in Figure 16. It is obvious that the development of the flame structure shows different behaviours among cases. The arrows in Figure 16 indicate that the flame front develops toward different directions among cases under the effect of in-cylinder flow.
The flame fronts concentrate mostly in the center of the combustion chamber from case F0 to F2 (R2), while the flame fronts gradually become irregular after case F3. It seems that the flame fronts prefer to develop towards the swirl center where the swirl terminates (as shown in Figure 4) under the impact of the in-cylinder flow motion. Specially, the flame front center
almost overlaps the swirl center under the severe swirl motion in F5. Therefore, it is easy to understand that the directions in the F3 and R3, F4 and R4 are symmetric to the x axis, in accordance with the locations of swirl center as shown in Figure 4. It should be noted that Figure 16 only shows the section views, therefore the flame area could not be used to quantitate the total heat release rate from flame propagation.
Figure 17 shows the iso-surface of the auto-ignited cells at crank angle with MBF around 60%
(CA60). Actually, the early flame propagation shows significant effect on the later auto-ignition process25. The flame propagation in the central region would elevate the pressure and temperature of the unburned mixture, finally leading to the subsequent auto-ignition of the end-gas. The effect of pressure increase on later auto-ignition is equal for all the unburned mixture, while the impact of the component transportation and heat transfer caused by the early flame propagation is only effective for the mixture near the flame front, and the interaction between the heat transfer from the central burned area and heat loss to the cylinder wall leads to the initial auto-ignition probably occurred at the rim of the flame front instead of the other positions. Therefore, the end-gas auto-ignition is codetermined by flame propagation process and the end-gas condition and shows high sensitivity to the in-cylinder flow. This in turn leads to distinct shapes among cases as shown in Figure 17. However, it is still possible to identify the effect of in-cylinder flow on the auto-ignition sites from F3, R3, F4 and R4. The arrows in the figure indicate apparent hollow parts of the wrinkle. It is found that these parts in F3 and R3, F4 and R4 still shows apparent symmetry to x axis, indicating the effect of in-cylinder flow on end-gas auto-ignition.
3.4. Correlations between in-cylinder condition and hybrid combustion process
In order to quantify the effect of in-cylinder flow on SI-CAI hybrid combustion, extensive analyses were performed to create the correlations between the in-cylinder conditions and the hybrid combustion characteristics.
Figure 18 shows the Pearson correlation coefficients between macroscopic flow parameters (SR, TR and CTR) and microscopic flow parameters and thermal conditions at 670 °CA ATDC. Because the plus-minus signs for SR, TR and CTR only indicate the direction, the absolute values of SR, TR and CTR are used to establish the correlations. For simplicity, only the parameters with strong correlations (at least one of the absolute coefficients is over 0.8) are shown in the figure. It is found Vm for the peripheral zones has strong positive correlation with SR, while the Vm for central zones have weaker correlation with SR. The reason is that the SR directly affects the peripheral velocity as shown in Figure 4. As discussed in the “in- cylinder thermal conditions” section, the strong swirl ratio would increase the heat transfer and inhibit the mixing of the fresh mixture and hot residual gas in central area, leading to strong negative correlation between SR and average temperature (T) in Zone 1. On the other hand, the TKE seems have some degree of positive correlation with the TR. One of the possible reason is the tumble gradually decays and transfers the kinetic energy into turbulence. Meanwhile, the increased turbulence would enhance the local mixing, leading to more homogeneous temperature. Therefore, it is also observed from Figure 18 that Ti, especially in Zone 1, has strong negative correlation with TR. The CTR is relatively small and shows little correlation with the microscopic flow parameters and thermal conditions.
Figure 19 shows the correlation coefficients between the mode transition crank angle and in- cylinder conditions at 670 °CA ATDC. Overall, the macroscopic flow, especially the swirl flow, shows weak correlation with CAT. The Vm in Zone 1 shows particularly stronger negative correlation with CAT compared to that for other zones, demonstrating the significant effect of the flow condition around spark plug on flame kernel development. The TKE for all zones show extremely strong negative correlation with CAT, indicating the dominated effect on flame propagation. It is interesting to find the average temperature for the peripheral zones has stronger positive correlation with CAT. Because the differences of average temperature among cases are not very significant as shown in Figure 11, the higher temperature for the end-gas in the peripheral zones indicates lower temperature for the central zones, leading to delayed mode transition.
In order to illustrate the effect of in-cylinder conditions on the flame propagation and auto- ignition process separately, Figure 20 and Figure 21 show the correlation coefficients between D1, D2 and all in-cylinder conditions at 670 °CA ATDC. The correlations between D1 and in-cylinder conditions are almost the same as that of CAT and in-cylinder conditions in Figure 19. As shown in Figure 20, there is a relatively strong positive correlation between temperature inhomogeneity for the central zones and D1, indicating that the higher temperature inhomogeneity also has the potential to slow the flame propagation process.
The correlations between auto-ignition dominated combustion duration D2 and in-cylinder conditions at 670 °CA are shown in Figure 21. Considering the effect of the early flame propagation on the in-cylinder conditions, the correlations between D2 and in-cylinder conditions at 720 °CA just before auto-ignition are also added in Figure 21 with the hollow symbols. The macroscopic flow parameters also show weak correlation with D2, while the Vm and TKE for the central zones show higher negative correlation with D2. This is mainly because the central flow condition significantly affects the flame propagation process which in turn affects the subsequent auto-ignition process through heat transfer between central burned gas and end-gas. Therefore, it is found in Figure 21 that the temperature at 720 °CA shows high negative correlation with D2. The auto-ignition duration D2 shows slightly strong correlation with the temperature inhomogeneity of the end-gas at 670 °CA, which proves the analysis about case F5 that large temperature inhomogeneity in Figure 13 leads to longer auto-ignition process in Figure 15. However, the correlation between temperature inhomogeneity at 720 °CA and D2 is not positive as the previous analysis. In fact, the higher temperature inhomogeneity at 720 °CA indicates fast flame propagations, which significantly enhances the auto-ignition. As shown in Figure 22, the flame propagation dominated combustion duration D1 and auto-ignition dominated combustion duration D2 do have positive correlation. This also demonstrates that the in-cylinder flow which shows significant impact on flame propagation would also significantly affects the subsequent auto-ignition. Figure 23 shows the strong positive correlations between CAT and auto-ignition dominated combustion duration D2. This indicates that the auto-ignition process would be prolonged with the delayed CAT. However, it can be seen in Figure 14 that CAT of F5 is a little earlier than that of F4, but D2 of F5 is still a little longer than that of F4 in Figure 15, indicating the significant effect of temperature inhomogeneity on the later auto-ignition process.
3.5. Sensitivity of the mode transition to the initial in-cylinder flow condition.
The scatter distributions and the correlation curves between CAT and the macroscopic flow are plotted in Figure 24. Overall, the scatter distributions of the macroscopic flow parameters (especially SR) show large dispersions. Figure 25 shows the scatter distributions and the correlation curves between CAT and the selected microscopic flow parameters with the highest correlation coefficients (Vm1 and TKE0) in Figure 19. It is obvious that the scatters in Figure 25 concentrate around the correlation curves, showing strong correlations. The scatter curves in Figure 26 show strong correlations between CAT and in-cylinder charge thermal conditions (T0, T7 and Ti0) at 670 °CA ATDC. In order to evaluate the sensitivity of the effect of the mode transition in SI-CAI hybrid combustion to the in-cylinder flow and thermal conditions, the slope of the correlation curves in Figure 24, 25 and 26 are shown in Figure 27.
For the microscopic flow conditions, it is found that CAT is most sensitive to the in-cylinder TKE0. A 1 m2/s2 increase of TKE0 would advance the CAT point by 0.45 °CA. Similarly, the mode transition is also very sensitive to the average Vm in Zone 1 with a slope of -0.41. For the macroscopic flow parameters, the TR has the strongest impact on CAT with a slope of 0.07. For the thermal conditions, the average temperature of the total charge (T0) shows strong impact on CAT with a slope of 0.34, while the effect of the temperature of the end-gas (T7) on CAT is more significant with a slope of 0.64. In addition, the temperature inhomogeneity also shows pronounced impact on CAT with the slope of 0.62.
These results indicate that the in-cylinder flow and thermal conditions, e.g. TKE, Vm, temperature and its inhomogeneity, could be used to effectively control the hybrid combustion mode transition. The fluctuations of these in-cylinder flow and thermal conditions could be the reasons leading to significant cycle-to-cycle variations in SI-CAI hybrid combustion.
4. Conclusions
In this study, the detailed and systematic investigation of the effect of in-cylinder flow motion on the in-cylinder condition and combustion process were performed using 3-D CFD simulations. The asymmetric intake valve events were adopted to generate the in-cylinder flow motion. The findings can be summarized as follows.
(1) It is proved in this study that asymmetric intake valve events could be used to control the in-cylinder macroscopic flow motion. The swirl flow could prevail after the intake valve closing and is gradually enhanced with increasing valve lifts asymmetry whilst the absolute values of tumble ratio (TR) only ranges between 0.06 and 0.61 among cases just before spark discharge.
(2) The average temperature of the whole cylinder exhibits a slight reduction with increasing swirl because of the increased heat loss, whilst the average temperature in the central zone experiences more temperature drop due to less mixing between fresh charge and hot residual gas. It is found in this study that the strong swirl flow would significantly inhibit the mixing in the central area although it enhances the mixing in the outer zones. The in-cylinder temperature inhomogeneity proves that the moderate swirl flow combined with the tumble motion is preferable to achieve the most effective mixing effect.
(3) Strong correlations between some macroscopic and zone-to-zone conditions are found in this study. Swirl ratio (SR) has strong positive correlations with the average flow velocity (Vm) in the peripheral zones and average temperature in the centre. TR shows positive correlation
with TKE and negative correlation with Ti in the centre. However, the average Vm and turbulent kinetic energy (TKE) in the central zone show weak correlations with SR and TR.
CTR is relatively small and shows no considerable correlations with the microscopic flow parameters and thermal conditions.
(4) The flame fronts prefer to develop towards the swirl center where the swirl terminates under the impact of the in-cylinder flow motion. The structure characteristics of auto-ignition sites show apparent symmetry along x axis between forward and reversed cases, indicating the effect of in-cylinder flow on end-gas auto-ignition sites.
(5) The effect of flow on SI-CAI hybrid combustion depends on the relative heat released by spark ignited flame propagation and auto-ignition combustion. As the swirl ratio is increased initially, the flame propagation is enhanced by the higher turbulence. But further increase in the swirl ratio can lead to retarded auto-ignition combustion due to higher heat loss of the unburned mixture characteristics and macroscopic flows. This is reflected by the increased value of mass fraction burned at the crank angle when transition from flame to auto-ignition.
(6) The significant effect of in-cylinder flow and thermal conditions, e.g. TKE, Vm, temperature and its inhomogeneity, on SI-CAI hybrid combustion found in this study indicates that the fluctuations of the in-cylinder flow and thermal conditions could be the reasons leading to significant cycle-to-cycle variations in SI-CAI hybrid combustion.
Funding
The study is a part of the National Science Fund project (Grant 51206118) supported by the National Science Fund Committee of China, and State Key Project of Fundamental Research Plan (Grant 2013CB228403) supported by the Ministry of Science and Technology of China.
The authors would also like to acknowledge the China Scholarship Council (CSC), which funded the first author for one year study at Brunel University.
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Figures:
Figure 1. The engine mesh and the definition of the swirl ratio (SR), tumble ratio (TR) and cross tumble ratio (CTR).
Figure 2. Schematic of the asymmetric intake valve events.
Figure 3. In-cylinder pressure and heat release rate from experiment and simulation (F0).
Figure 4. In-cylinder streamlines viewed from +y and +z direction at 640 ºCA ATDC.
Figure 5. Evolution of in-cylinder swirl ratio (SR), tumble ratio (TR) and cross tumble ratio (CTR) from 400 ºCA to 670 ºCA ATDC.
Figure 6. The swirl ratio (SR), tumble ratio (TR) and cross tumble ratio (CTR) at 670 °CA ATDC.
Figure 7. Schematic of the zones defined to reveal the in-cylinder zone-to-zone conditions.
Figure 8. Average velocity magnitude (Vm) for the whole combustion chamber, Zone 1 and Zone 7 at 670 °CA ATDC.
Figure 9. Average turbulent kinetic energy (TKE) for the whole combustion chamber, Zone 1 and Zone 7 at 670 °CA ATDC.
Figure 10. In-cylinder section views of the temperature distribution at 670 °CA ATDC.
Figure 11. Average temperature for whole chamber, Zone 1 and Zone 7 at 670 °CA ATDC.
Figure 12. The residual gas fraction (RGF) for whole chamber, Zone 1 and Zone 2 at 670 °CA
ATDC.
Figure 13. Temperature inhomogeneity for whole chamber, Zone 1 and Zone 7 at 670 °CA ATDC.
Figure 14. CA10, CA50, CA90, CAT and the MFB at CAT for all cases.
Figure 15. Flame propagation dominated combustion duration D1 (from CA10 to CAT) and auto-ignition dominated combustion duration D2 (from CAT to CA90) for all cases.
Figure 16. Section views of the average flame density distribution at CA10.
Figure 17. Iso-surface of the newly auto-ignited cells at CA60.
Figure 18. Correlations between macroscopic flow parameters (SR and TR) and microscopic flow parameters and thermal conditions at 670 °CA ATDC.
Figure 19. Correlations between CAT and in-cylinder flow and thermal conditions at 670 °CA ATDC.
Figure 20. Correlations between D1 and in-cylinder flow and thermal conditions at 670 °CA ATDC.
Figure 21. Correlations between D2 and in-cylinder flow and thermal conditions at 670 °CA ATDC (solid symbols) and 720 °CA ATDC (hollow symbols).
Figure 22. Correlations between flame propagation dominated combustion duration D1 and auto-ignition dominated combustion duration D2.
Figure 23. Correlations between CAT and auto-ignition dominated combustion duration D2.
Figure 24. The scatter distribution and correlation curves between CAT and the macroscopic flow conditions (SR, TR and CTR) at 670 °CA ATDC.
Figure 25. The scatter distribution and correlation curves between CAT and the microscopic flow conditions (Vm1 and TKE0) at 670 °CA ATDC.
Figure 26. The scatter distribution and correlation curves between CAT and the thermal conditions (T0, T7 and Ti0) at 670 °CA ATDC.
Figure 27. Slopes of the correlation curves between CAT and in-cylinder conditions shown in Figure 24, 25 and 26.
Tables:
Table 1. Engine specifications.
Bore 86 mm
Stroke 86 mm
Displacement 0.5 L
Compression ratio 10.66
Combustion chamber Pent roof / 4 valves Fuel injection Port fuel injection
Fuel Gasoline 93 RON
Intake pressure Naturally aspirated
Throttle WOT
Table 2. Engine oprating conditions.
EVO/ EVC [ºCA*] 108 / 326
EL [mm] 5.5
IVO / IVC [ºCA*] 390 / 604
IL [mm] 5.1
Spark Timing [ºCA*] 671 Fueling Rate [mg/cycle] 25 Intake temperature [ºC] 40 Coolant temperature[ºC] 85 Oil temperature [ºC] 55 Engine speed [r/min] 1500
* The reference is the combustion top dead center.
Table 3. Intake valve events and in-cylinder mass.
Case IL1 [mm] IL2 [mm] C Mass [mg]
F0 5.1 5.1 1 442
F1 5.6 4.6 1.1 443
F2 6.1 4.1 1.2 443
R2 4.1 6.1 0.8 443
F3 6.4 3.8 1.25 442
R3 3.8 6.4 0.75 442
F4 6.6 3.6 1.3 442
R4 3.6 6.6 0.7 442
F5 7.1 3.1 1.4 442
