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biodiesel fueled RCCI engine at high speed
Energy Conversion and Management 112 (2016) 359–368
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Modeling study on the effect of piston bowl geometries in a gasoline/biodiesel fueled RCCI engine at high speed J. Li, W.M. Yang ⇑, D.Z. Zhou Department of Mechanical Engineering, Faculty of Engineering, National University of Singapore, Singapore 117575, Singapore
a r t i c l e
i n f o
Article history: Received 23 November 2015 Accepted 16 January 2016
Keywords: RCCI Biodiesel Gasoline Bowl geometry KIVA
a b s t r a c t This paper reports the numerical investigation on the effects of three bowl geometries on a gasoline/ biodiesel fueled RCCI engine operated at high engine speed. The three bowl geometries are HCC (Hemispherical Combustion Chamber), SCC (Shallow depth Combustion Chamber) and OCC (Omega Combustion Chamber). To simulate the combustion in an RCCI engine, coupled KIVA4–CHEMKIN code was used. One recently developed reaction mechanism, which contains 107 species and 425 reactions, was adopted in this study to mimic the combustion of gasoline and biodiesel. During the simulation, the engine speed was fixed at 3600 rpm. The low reactivity fuel gasoline was premixed with air with energy percentages of 20% and 40%; accordingly, to maintain the same energy input, the percentages of biodiesel were 80% and 60% (B80 and B60). In addition, the SOI timing was varied at three levels: 11, 35 and 60 deg ATDC for B80 and B60, respectively. With SOI timing of 11 deg ATDC, the combustion is mixing-controlled; in contrast, advancing SOI timing to 60 deg ATDC, the combustion turns into the reactivity-controlled. Comparing the results on combustion characteristics, engine performance and emissions among different bowl geometries, it is concluded that the original OCC design for Toyota diesel engine is better for mixing-controlled combustion; whereas, SCC is the most suitable piston design for RCCI combustion among the three selected geometries under the investigated operating conditions of the engine. With SCC, better combustion and performance can be achieved while maintaining relatively lower CO, NO and soot emissions. Ó 2016 Elsevier Ltd. All rights reserved.
1. Introduction Energy shortage and air pollution are the two major issues that intrigue the engine researchers to find advanced in-cylinder combustion with high efficiency and low emissions. RCCI (reactivity controlled compression ignition), which was proposed in 2009 by Kokjohn et al. [1], has been considered as an advanced combustion scenario. Compared to CDC (conventional diesel combustion), RCCI can reduce soot and NOx emissions to ultra-low levels. It has been investigated by adopting gasoline/diesel fuel combination by some researchers [2–7]. Recently, biodiesel, which is derived from plant or animal sources, and thus renewable, has been tested in CDC engines [8,9]. The oxygen-borne biodiesels can enhance the oxidation of fuels, and consequently bring more complete combustion to CDC engines. However, increased NOx emissions were usually observed in biodiesel fueled diesel engine due to the contained oxygen [9,10]. Therefore, RCCI technique can be applied to alter the combustion mode of CDC engines fueled with biodiesels, ⇑ Corresponding author. Tel.: +65 6516 6481. E-mail address: [email protected] (W.M. Yang). http://dx.doi.org/10.1016/j.enconman.2016.01.041 0196-8904/Ó 2016 Elsevier Ltd. All rights reserved.
consequently to further improve the thermal efficiency and meanwhile lower the NOx emissions. To realize the RCCI combustion in this study, biodiesel was chosen as the high reactivity fuel; whereas gasoline was selected as the low reactivity fuel. Our group has conducted one study on gasoline/biodiesel fueled RCCI combustion [11], where the effects of gasoline ratio and SOI timing were investigated. The results showed that less NOx emissions were generated in gasoline/biodiesel fueled RCCI combustion compared to CDC fueled with pure biodiesel. In this study, to further optimize the gasoline/biodiesel fueled RCCI combustion, the effect of piston bowl geometry was taken into consideration. The piston bowl plays an essential role in facilitating the mixing of air and fuel within the combustion chamber [12,13]. Splitter et al. [14] suggested to decrease the surface to volume ratio while designing the bowl geometry of RCCI engines for the purpose of reducing heat losses. Hanson et al. [2] compared the designed RCCI piston with original diesel piston. It was found that the RCCI piston, which has a lower surface to volume ratio, can provide 3% of increases in thermal efficiency with ultra-low levels of NOx and PM emissions at the engine speed of 2600 rpm and IMEP of 6.9 bar. Benajes et al. [15] also compared three bowls
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Table 1 Models in KIVA4–CHEMKIN. Name
Model
Turbulent model Spray model Breakup model Collision model Combustion model
RNG k—e model Discrete droplet model KH–RT model O’Rourke model CHEMKIN II solver
which are stock, stepped, and bathtub at the engine speed of 1200 rpm under low, medium and high loads. The stepped bowl was proved to be the most suitable piston for the tested RCCI engine. In the past investigation undertaken by our group [16], three types of bowl geometries, namely HCC (Hemispherical Combustion Chamber), SCC (Shallow depth Combustion Chamber) and OCC (Omega Combustion Chamber), were selected and tested at 1200, 2400 and 3600 rpm in a conventional diesel engine. In this study, the objective is to find the suitable piston for the RCCI engine which was modified from the diesel engine. It may be assumed that the piston bowl geometry designed for CDC may not be suitable for RCCI combustion. Moreover, the studies on bowl geometries of RCCI engines which have been carried out so far were operated at low engine speeds [2,14,15], at which the combustion process much relies on the piston surface to volume ratio [16]. The surface to volume ratio could significantly affect the heat transfer process and
(a) HCC
consequently play a dominant role in the combustion process. Thus, in this study, the gasoline/biodiesel fueled RCCI engine was operated at high engine speed to investigate whether the surface to volume ratio of a piston is still crucial for the combustion process. Conclusively, there is a need to study the effect of bowl geometry on a gasoline/biodiesel fueled RCCI engine at high engine speed. In this study, simulations with a newly developed mechanism were conducted based on the high speed (3600 rpm) operating condition in [16], 20% and 40% of gasoline was premixed while maintaining the same energy input. Moreover, the SOI (start of injection) timing was varied from 11 to 60 deg ATDC to transform the combustion mode from the mixing-controlled to the reactivity-controlled. Finally, the results on combustion characteristics including in-cylinder pressure, HRR (heat release rate), PRR (pressure rise rate) and combustion duration, engine performance and emissions formation on CO, NO and soot were compared.
2. Methodology 2.1. KIVA4–CHEMKIN A multi-component fuel combustion model [17], which coupled CHEMKIN II solver with KIVA-4 code, was used in this study. KIVA4 [18] is a computer code for the numerical calculation of the transient, two and three dimensional, multiphase, multicomponent chemically reactive flows with sprays. However, in the original
(b) SCC
(c) OCC
Fig. 1. Generated grids of bowl geometries.
Table 2 The engine specifications and operating conditions for validation cases. (a) Pure biodiesel Engine type Number of cylinders Bore, mm Stroke, mm Displacement, L Compression ratio Injector nozzle number Intake valve close, deg ATDC Exhaust valve open, deg ATDC
GW4D20 1 83.1 92 0.5 16.7 7 130 94
Engine speed, rpm Pressure at IVC, bar Temperature at IVC, K Biodiesel consumption, kg/h Gasoline consumption, kg/h SOI (ATDC)
1400 1.1 360 0.45 0.0 7.0
Piston bowl at TDC
(b) Gasoline/biodiesel
(c) Pure biodiesel Toyota 2KD-FTV 4 92 93.8 2.494 18.5 6 149 150
1400 1.10 362 0.26 0.24 9.0
3600 1.89 370 11.34 0.0 11.0
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(a) Pure biodiesel in GW4D20
Fig. 3. In-cylinder pressure and HRR of B80.
(b) Gasoline/biodiesel in GW4D20
Fig. 4. Temperature distribution and fuel spray of OCC with B80 and SOI timing of 11 deg at TDC.
Rayleigh–Taylor (KH–RT) hybrid model. RNG k–e model is selected for the consideration of turbulence.
(c) Pure biodiesel in Toyota 2KD-FTV Fig. 2. Validation of in-cylinder pressure and HRR of the pure biodiesel and gasoline/biodiesel combustion.
Table 3 Validation on NOx emissions. NOx (g/kW h)
Experiment Simulation
Case (a)
Case 2 (b)
2.29 3.29
1.61 1.86
KIVA-4, only a global reaction and 12 species were included. To realize the detailed chemistry calculation, CHEMKIN II was coupled with KIVA-4. Then a newly developed reaction mechanism by our group [11,19] was used to model the chemical kinetics of gasoline and biodiesel. In this mechanism, there are 107 species and 425 reactions. Among the species, iso-octane is considered as the surrogate of gasoline, and the mixture with 25% MD (methyldecenoate), 25% MD9D (methyl-9-decenoate) and 50% n-heptane in mole represents biodiesel. Table 1 lists the models used in this study. The original breakup model in KIVA4 was replaced with Kelvin–Helmholtz and
2.2. Piston bowl geometry The effect of three different pistons on a diesel engine fueled with pure biodiesel has been investigated in our previous study [16]. Same pistons were selected in this study to examine their effects on the gasoline/biodiesel fueled RCCI engine. Fig. 1 demonstrates the geometries and generated grids of the three pistons: HCC, SCC and OCC. The compression ratio is fixed at 18.5. From the previous study [16], it is known that during the compression stroke in the bowl, OCC can form the strongest squish flow, then followed by HCC; SCC generates the weakest among the three. The piston surface area of OCC is 10,456 mm2 which is the largest. HCC and SCC have similar surface areas which are 8118 and 8192 mm2, respectively. In addition, the throat diameter of SCC is the largest among the three, followed by HCC, and the shortest is OCC. 2.3. Validation To validate the models, experimental data were firstly compared with numerical results under two conditions, fueling
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(a) HCC
(b) SCC
(c) OCC
Fig. 5. Vector (Length(cm)/magnitude = 0.0002) for B80 with SOI timings of
gasoline/biodiesel and pure biodiesel, respectively. Table 2 lists the engine specifications and operating conditions for the validations. The experimental data was obtained from GW4D20 and Toyota 2KD-FTV. The experimental setup of GW4D20 is at Tsinghua University [20]. The in-cylinder pressure is measured by AVL GH14P sensor. The data is recorded in 150 consecutive cycles for combustion analysis. The NOx emission is measured using AVL CEB2 gas analyzer. The measurement on soot is conducted using AVL 439 opacimeter. The NOx and soot are averaged for 30s when the engine condition is stable. The Toyota 2KD-FTV engine setup is at National University of Singapore. The cylinder pressure was measured by AVL GH13P water-cooled pressure transducer with a resolution of 1 °CA. The measured data is obtained from averaged 50 cycles for each condition. The cooling water temperature is maintained at 70 °C. The NO and NO2 are measured by a portable gas analyzer (E instrument 4400 N). Fig. 2 compares the in-cylinder pressure and HRR between experiments and simulation for both pure biodiesel and gasoline/ biodiesel combustion. Though a slightly over-prediction on pressure and HRR can be observed, the discrepancies are still
(a) Temperature (Kelvin) iso-surface
60 deg ATDC at CA50.
acceptable. The abrupt HRR of the simulation on gasoline/biodiesel fueled combustion maybe due to the assumption on the homogeneously premixed gasoline during the simulation. In reality, the gasoline distribution cannot be homogeneous. Thus, the HRR of dual-fuel simulation is faster compared to experiment. Thus, the models and reaction mechanism is validated. The validation on emissions formation is conducted on GW4D20 engine (case (a) and case (b) in Table 3). As mentioned, the NOx and soot are measured by AVL CEB-II and AVL 439 opacimeter, respectively. Table 3 compares the NOx emissions under conditions of validation case (a) and (b). It can be seen that with increase in gasoline ratio, the NOx was reduced for both experiment and simulation cases. Thus, the model could predict the NOx formation trend by increasing the gasoline fraction. The experimental data on soot is not suitable to validate the simulation because the measured smoke by an opacimeter includes not only soot, but also solids and liquids etc. However, it is found that with the increase in gasoline, the predicted soot increases. This trend is consistent with that observed from experiment.
(b) Equivalence ratio iso-surface
Fig. 6. Iso-surfaces of (a) temperature and (b) equivalence ratio of three bowl geometries with B80 and SOI timing of
60 deg at CA50.
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2.4. Study cases As mentioned, the studies carried out so far were focused on the RCCI combustion at low engine speed where the surface area of a piston can play a crucial role on the combustion process [16]. To study the effect of piston bowl geometry on RCCI combustion at high engine speed, the validation case with pure biodiesel was selected as the base case. Thus, OCC is the original design piston for the Toyota 2KD-FTV engine. Then the total energy input remained constant, gasoline was premixed in the cylinder with the energy percentages of 20% and 40%, respectively; accordingly, the percentages of biodiesel decreased to 80% (B80) and 60% (B60). It should be noted that to further increase gasoline ratio, the abrupt peak pressure rise rate is unacceptable due to the absence of EGR (exhaust gas recirculation). Thus, the results with more premixed gasoline (40% above) will not be presented in this study. Moreover, the SOI timing of the base case is 11 deg ATDC, which results in the mixing-controlled combustion [21]. To form the reactivity-controlled combustion, the SOI timing was further advanced to 35 and 60 deg ATDC. Thus, for each bowl geometry, numerical simulations were conducted for B80 and B60, with SOI timings of 11, 35, and 60 deg ATDC, respectively.
(a) B80
3. Results and discussion 3.1. Combustion characteristics 3.1.1. In-cylinder pressure and heat release rate Fig. 3 shows the in-cylinder pressure and HRR of B80 with SOI timings of 11, 35 and 60 deg ATDC. As can be observed, with SOI timing of 11 deg, the peak pressures from the highest to the lowest are of OCC, HCC and SCC, which is associated with the intensity of squish flow formed in piston bowl. This is because the fuel spray targets at piston bowl area and subsequently the combustion happens in the bowl, as shown in Fig. 4. Thus, with late SOI timing, the combustion process is strongly dependent on the
(b) B60 Fig. 8. Maximum pressure rise rate.
Fig. 7. In-cylinder pressure and HRR of B60.
mixing process of fuel and premixed mixture, and the OCC can produce better combustion due to its strong squish formed in the bowl. By advancing SOI timing, the combustion mode gradually transits from the mixing-controlled to the reactivity-controlled [21]. With SOI at 35 deg ATDC, OCC still shows the highest peak pressure; however, HCC could not show superiority compared to SCC. Further advancing SOI timing to 60 deg ATDC, SCC can achieve a slightly higher peak pressure than HCC and OCC. This indicates that the squish flow in piston bowl cannot prevail the combustion process when the mode is reactivity-controlled. This is because for the RCCI combustion with SOI of 60 deg ATDC, the fuel spray will target at squish region and part of the fuel will deposit on the piston surface and cylinder liner. The slightly higher peak pressure of SCC may be due to its shallow depth and wider throat diameter by which the squish flow formed in the bowl can be pushed much far away from the centerline of the bowl, as shown in Fig. 5, which displays the vector of squish flow. Fig. 6 displays the iso-surfaces of temperatures at 1300, 1500, 2000 K and equivalence ratio of 0.14, 0.5 and 1.0 for three bowl geometries with B80 and SOI timing of 60 deg at CA50. CA50 is the crank angle where 50% of heat is released. It can be seen that due to the early injection, high equivalence ratio is observed near
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Fig. 10. In-cylinder average temperature profile with B80 and SOI timing of 60 deg ATDC.
(a) B80
reducing. When SOI timing is advanced to 35 deg ATDC, the peak pressure of OCC reduces. With SOI timing of 60 deg ATDC, same trend on peak pressure is observed with those of B80. SCC can achieve the highest peak pressure during RCCI combustion.
(b) B60 Fig. 9. Combustion duration with SOI timings of
11,
35,
60 deg ATDC.
the cylinder liner (red1 surface in Fig. 6(b)). SCC bowl has a shallow depth and larger diameter, as such, the far-away squish flow formed in the bowl could be more effectively directed to the near-wall region and entrain part of the biodiesel in the bowl (higher equivalence ratio in Fig. 6(b)), thereby enhancing the fuel mixing process in this region subsequently resulting in better combustion (Fig. 6(a)). This also explains the higher HRR of SCC compared to HCC and OCC. Moreover, the highest peak pressure of SCC may be also partly due to the reduced heat losses caused by the less surface area. Also, it is worth noting that by advancing SOI timing from 11 to 60 deg ATDC, the combustion mode transits from slow single-stage combustion to fast single-stage combustion and then two-stage combustion [22]. Fig. 7 displays the in-cylinder pressure and HRR of B60 with SOI timings of 11, 35 and 60 deg ATDC. Compared to B80, B60 can create relatively higher reactivity difference. With SOI timings of 11 deg ATDC, the peak pressures are still following the trend as shown in B80 which indicates the combustion is still mixingcontrolled; however, the gaps among the peak pressures are
1
For interpretation of color in Fig. 6, the reader is referred to the web version of this article.
(a) B80
(b) B60 Fig. 11. IMEP with SOI timings of
11,
35,
60 deg ATDC.
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(a) B80
(a) B80
(b) B60
(b) B60
Fig. 12. CO emissions with SOI timings of
11,
35,
60 deg ATDC.
3.1.2. Maximum peak pressure rise rate Fig. 8 compares the maximum PRR. The acceptable limit of maximum PRR is chosen as 10 bar/deg. As shown in the figure, only B80 with SOI timing of 11 and 60 deg ATDC can satisfy the criterion. The other operating conditions would suffer the unacceptable engine knocking. The high maximum PRR for B80 with SOI timing of 35 deg ATDC is caused by the fast single heat release combustion. When the premixed gasoline increases to 40%, i.e., B60, the maximum PRRs all exceed 10 bar/deg regardless of the SOI timings. The fast combustion with higher gasoline ratio is due to the fast
(a) HCC
Fig. 14. NO emissions with SOI timings of
11,
35,
60 deg ATDC.
combustion resulted from gasoline–air mixture. Also, it should be noted that EGR is not applied to all the cases in this study.
3.1.3. Combustion duration Fig. 9 shows the combustion duration for (a) B80 and (b) B60 with SOI timings of 11, 35, 60 deg ATDC. Combustion duration is defined as the interval between CA5 and CA90, which are the crank angles when 5% and 90% of heat release are completed, respectively. From Fig. 9(a) and (b), it can be seen that with SOI
(b) SCC
Fig. 13. Spatial CO distribution and vector (Length(cm)/magnitude = 0.0001) for B80 with SOI timings of
(c) OCC 35 deg ATDC at crank angle of 25 deg.
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timing of 11 deg ATDC, the combustion duration has a negative correlation with the squish strength during compression stroke in the bowl. SCC which has the weakest squish flow can achieve the longest combustion duration. However, when SOI timing is advanced to 60 deg ATDC, the combustion duration of HCC and OCC can be extended to more than 100 degree; in contrast, the duration of SCC is about 60 degree. The relatively short combustion duration with SOI timing of 11 deg ATDC is due to the mixingcontrolled combustion as mentioned. The stronger squish flow in OCC results in a better mixing process between biodiesel and gasoline–air mixture, subsequently completing combustion at a rapid rate. When the combustion mode transits to the reactivitycontrolled, SCC achieves relatively short combustion duration, though this duration is longer compared to that of all the pistons with SOI timings of 11 and 35 deg ATDC. This short combustion duration of SCC with SOI of 60 deg ATDC can be attributed to the stronger squish flow far away from the centerline caused by the wide throat diameter. This far-away strong flow can be effectively directed to the near wall region, providing better mixing as well as better combustion and high temperature. The higher temperature leads to the fast consumption of fuel and consequently brings short combustion duration. In addition, the increased in-cylinder temperature caused by less heat loss of SCC leads to shorter combustion duration. Fig. 10 shows the in-cylinder average temperature profile with B80 and SOI timing of 60 deg ATDC. It can be observed that the temperature of SCC is higher compared to HCC and OCC, which could consolidate the inference. 3.2. Engine performance Fig. 11 displays the IMEP (indicated mean effective pressure) for (a) B80 and (b) B60 with SOI timings of 11, 35 and 60 deg ATDC. As can be observed from Fig. 11(a) and (b), with B80, OCC and HCC, which have stronger squish flow, can generate the highest IMEP when SOI timing is 11 deg ATDC. This result can be inferred from the peak pressure observed in Fig. 3. Thus, the strong squish flow of a bowl geometry could achieve better performance for a mixing-controlled combustion. However, when the
SOI timing is advanced to 60 deg ATDC, the best performance is observed for SCC. The degraded performance for OCC and HCC may be due to the longer combustion duration.
3.3. Emissions formation 3.3.1. CO emissions Fig. 12 compares the CO emissions for B80 and B60 with SOI timings of 11, 35, 60 deg ATDC. Similar trends are observed for B80 and B60. With SOI timing of 11 deg ATDC, CO levels for all three pistons are almost the same. With advancing SOI timing to 35 deg ATDC, CO of HCC significantly increases whereas the other two remains at the low levels. Further advancing SOI to 60 deg ATDC, both HCC and OCC generates high CO emissions. Fig. 13 shows the spatial CO distribution and the vector for B80 with SOI timings of 35 deg ATDC at the crank angle of 25 deg. The CO of HCC and SCC are mainly generated in the bowl region; in contrast, the CO of OCC is mainly at squish region. This indicates that the combustion of HCC and SCC happens in the bowl and that of OCC happens in the squish region. This is because when SOI timing is 35 deg ATDC, the fuel spray still targets at bowl region of HCC and SCC due to their wider throat diameter compared to OCC. However, the CO of HCC generated in the bowl could not be further oxidized due to the weak mixture flow as implied from the length of the vector, as shown in Fig. 13(a). In contrast, the CO generated in SCC can be further oxidized because of the stronger flow during combustion stroke. With SOI timing of 60 deg ATDC, the fuel spray targets at squish regions for all three pistons. The high CO for HCC and OCC could be attributed to the much extended combustion duration (above 100 deg CA) as shown in Fig. 9. When the combustion duration is very long, the in-cylinder temperature at the end of the combustion process could be much lower which is unfavorable for the oxidation of CO emissions. Thus, compared to HCC and OCC, SCC can achieve low CO emissions with the relatively short combustion duration caused by the far-away squish flow effect in the squish region due to the wide throat diameter.
(a) OCC at CA50 Fig. 15. Phi–T map for B60 with SOI timings
(b) SCC at CA50 60 deg ATDC.
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3.3.2. NO emissions Fig. 14 shows the NO emissions for B80 and B60 with SOI timings of 11, 35, 60 deg ATDC. With B80, when the SOI timing is 60 deg ATDC, similar levels of NO emissions were found for the three piston bowls. Significant high NO can be seen for OCC piston bowl when the SOI timings are 11 and 35 deg ATDC. This high NO is resulted from the high temperature as implied from the peak HRR shown in Fig. 3. Similarly, with B60, the NO levels with SOI timings are associated with the peak pressure observed in Fig. 7. It is interesting to notice that when the combustion is reactivity-controlled, OCC generates relatively lower NO. The possible reason could be the relatively poorer combustion of OCC as implied from the pressure histories. Another reason could be the large surface area of OCC contributes to the heat losses which would subsequently lower the in-cylinder temperature. The incylinder temperature can be demonstrated in Fig. 15, which is the /–T (Equivalence ratio–Temperature) diagram [11] for OCC and SCC when the SOI timing is 60 deg ATDC with B60. The overall temperature distribution of OCC is lower compared to that of SCC.
(a) B80
3.3.3. Soot emissions Fig. 16 displays the soot emissions for B80 and B60 with SOI timings of 11, 35, 60 deg ATDC. It can be observed that there is not too much difference on the soot emissions for the three bowl geometries. Although the results are close to each other, OCC demonstrates a relatively lower soot under most of the conditions. The reason could be implied from Fig. 17. Fig. 17 shows the spatial temperature and equivalence ratio distribution for B60 with SOI timings of 60 deg ATDC at CA50. The vector in the cylinder is also presented along with the temperature distribution. It can be seen that OCC can generate strong flow near the cylinder liner compared to HCC and SCC, thus providing a better mixing process and subsequently resulting in local low equivalence ratio regions. Consequently, OCC generates relatively lower soot emissions compared to HCC and SCC. 4. Conclusion
(b) B60 Fig. 16. Soot emissions with SOI timings of
(a) HCC
11,
35,
60 deg ATDC.
In this study, three bowl geometries, namely HCC, SCC and OCC, were numerically examined in a gasoline/biodiesel fueled RCCI
(b) SCC
Fig. 17. Spatial temperature with vector (Length(cm)/magnitude = 0.0002) and equivalence ratio for B60 with SOI timings of
(c) OCC 60 deg ATDC at CA50.
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engine at low, medium and high loads with engine speed of 3600 rpm. While changing the gasoline and biodiesel ratios, the total energy input was kept constant. In addition, the SOI timing should be varied during simulation to achieve reactivitycontrolled combustion. Finally, engine simulations were conducted for each bowl geometry with 80% and 60% of direct-injected biodiesel while adopting SOI timings at 11, 35 and 60 deg ATDC, respectively. The results on combustion characteristics, engine performance and emissions were compared. The major conclusions can be drawn as follows: (1) Regarding the combustion characteristics, the original designed OCC for diesel engine is better for the mixingcontrolled combustion due to its strong intensity of squish flow in the piston bowl during compression stroke. However, it is not suitable for the reactivity-controlled combustion due to its narrowed squish flow effect and its higher heat losses caused by larger surface area. In contrast, SCC is better for RCCI combustion because of the relatively short combustion duration. Thus, the throat diameter and surface area of a piston bowl can play a crucial role in affecting the combustion process at high engine speed. (2) For RCCI combustion, in terms of emissions, SCC can achieve low CO emissions which indicates a higher combustion efficiency; however, the NO emissions of SCC are relatively higher than that of OCC because of the overall higher combustion temperature. Meanwhile, similar levels of soot emissions were observed for three bowl geometries, though a slightly lower soot can be found for OCC. (3) Overall, SCC is the best piston design for RCCI combustion among the three selected geometries under the investigated operating conditions of the Toyota engine in this study. To use SCC, better combustion process and performance can be achieved while maintaining relatively lower CO, NO and soot emissions. Acknowledgement This work is supported by the program of China Scholarship Council (CSC). Appendix A. Supplementary material Supplementary data associated with this article can be found, in the online version, at http://dx.doi.org/10.1016/j.enconman.2016. 01.041. References [1] Kokjohn SL, Hanson RM, Splitter DA, Reitz RD. Experiments and modeling of dual-fuel HCCI and PCCI combustion using in-cylinder fuel blending. SAE Int J Engine 2010;2:24–39.
[2] Hanson R, Curran S, Wagner R, Kokjohn S, Splitter D, Reitz R. Piston bowl optimization for RCCI combustion in a light-duty multi-cylinder engine. SAE Int J Engine 2012;5:286–99. [3] Kokjohn SL, Hanson RM, Splitter DA, Reitz RD. Fuel reactivity controlled compression ignition (RCCI): a pathway to controlled high-efficiency clean combustion. Int J Engine Res 2011;12:209–26. [4] Hanson RM, Kokjohn SL, Splitter DA, Reitz RD. An experimental investigation of fuel reactivity controlled PCCI combustion in a heavy-duty engine. SAE Int J Engine 2010;3:700–16. [5] Kokjohn SL, Reitz RD. Characterization of dual-fuel PCCI combustion in a lightduty engine. In: Proceedings of the international multi-dimensional engine modeling user’s group meeting; 2010. [6] Splitter D, Kokjohn S, Rein K, Hanson R, Sanders S, Reitz R. An optical investigation of ignition processes in fuel reactivity controlled PCCI combustion. SAE Int J Engine 2010;3:142–62. [7] Benajes J, Molina S, García A, Belarte E, Vanvolsem M. An investigation on RCCI combustion in a heavy duty diesel engine using in-cylinder blending of diesel and gasoline fuels. Appl Therm Eng 2014;63:66–76. [8] An H, Yang WM, Maghbouli A, Li J, Chou SK, Chua KJ. Performance, combustion and emission characteristics of biodiesel derived from waste cooking oils. Appl Energy 2013;112:493–9. [9] Vedharaj S, Vallinayagam R, Yang WM, Chou SK, Chua KJE, Lee PS. Experimental investigation of kapok (Ceiba pentandra) oil biodiesel as an alternate fuel for diesel engine. Energy Convers Manage 2013;75:773–9. [10] Vallinayagam R, Vedharaj S, Yang WM, Lee PS, Chua KJE, Chou SK. Combustion performance and emission characteristics study of pine oil in a diesel engine. Energy 2013;57:344–51. [11] Li J, Yang WM, An H, Zhao D. Effects of fuel ratio and injection timing on gasoline/biodiesel fueled RCCI engine: a modeling study. Appl Energy 2015;155:59–67. [12] Vedharaj S, Vallinayagam R, Yang WM, Saravanan CG, Lee PS. Optimization of combustion bowl geometry for the operation of kapok biodiesel – diesel blends in a stationary diesel engine. Fuel 2015;139:561–7. [13] Kokjohn SL, Reitz RD. Reactivity controlled compression ignition and conventional diesel combustion: a comparison of methods to meet lightduty NOx and fuel economy targets. Int J Engine Res 2013;14:452–68. [14] Splitter D, Wissink M, Kokjohn S, Reitz R. Effect of compression ratio and piston geometry on RCCI load limits and efficiency. SAE Technical Paper; 2012. [15] Benajes J, Pastor JV, García A, Monsalve-Serrano J. An experimental investigation on the influence of piston bowl geometry on RCCI performance and emissions in a heavy-duty engine. Energy Convers Manage 2015;103:1019–30. [16] Li J, Yang W, An H, Maghbouli A, Chou S. Effects of piston bowl geometry on combustion and emission characteristics of biodiesel fueled diesel engines. Fuel 2014;120:66–73. [17] Maghbouli A, Yang W, An H, Li J, Chou SK, Chua KJ. An advanced combustion model coupled with detailed chemical reaction mechanism for D.I diesel engine simulation. Appl Energy 2013;111:758–70. [18] Torres DJ, Trujillo MF. KIVA-4: an unstructured ALE code for compressible gas flow with sprays. J Comput Phys 2006;219:943–75. [19] Li J, Yang WM, An H, Zhou DZ, Yu WB, Wang JX, et al. Numerical investigation on the effect of reactivity gradient in an RCCI engine fueled with gasoline and diesel. Energy Convers Manage 2015;92:342–52. [20] Wang B, Wang Z, Shuai S, Yang H, Wang J. Combustion and emission characteristics of Multiple Premixed Compression Ignition (MPCI) fuelled with naphtha and gasoline in wide load range. Energy Convers Manage 2014;88:79–87. [21] Li Y, Jia M, Chang Y, Liu Y, Xie M, Wang T, et al. Parametric study and optimization of a RCCI (reactivity controlled compression ignition) engine fueled with methanol and diesel. Energy 2014;65:319–32. [22] Chao Y, Zhi W, Jianxin W. Sequenced combustion characteristics, emission and thermal efficiency in gasoline homogeneous charge induced ignition. Appl Energy 2014;124:343–53.
Contents lists available at ScienceDirect
Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman
Modeling study on the effect of piston bowl geometries in a gasoline/biodiesel fueled RCCI engine at high speed J. Li, W.M. Yang ⇑, D.Z. Zhou Department of Mechanical Engineering, Faculty of Engineering, National University of Singapore, Singapore 117575, Singapore
a r t i c l e
i n f o
Article history: Received 23 November 2015 Accepted 16 January 2016
Keywords: RCCI Biodiesel Gasoline Bowl geometry KIVA
a b s t r a c t This paper reports the numerical investigation on the effects of three bowl geometries on a gasoline/ biodiesel fueled RCCI engine operated at high engine speed. The three bowl geometries are HCC (Hemispherical Combustion Chamber), SCC (Shallow depth Combustion Chamber) and OCC (Omega Combustion Chamber). To simulate the combustion in an RCCI engine, coupled KIVA4–CHEMKIN code was used. One recently developed reaction mechanism, which contains 107 species and 425 reactions, was adopted in this study to mimic the combustion of gasoline and biodiesel. During the simulation, the engine speed was fixed at 3600 rpm. The low reactivity fuel gasoline was premixed with air with energy percentages of 20% and 40%; accordingly, to maintain the same energy input, the percentages of biodiesel were 80% and 60% (B80 and B60). In addition, the SOI timing was varied at three levels: 11, 35 and 60 deg ATDC for B80 and B60, respectively. With SOI timing of 11 deg ATDC, the combustion is mixing-controlled; in contrast, advancing SOI timing to 60 deg ATDC, the combustion turns into the reactivity-controlled. Comparing the results on combustion characteristics, engine performance and emissions among different bowl geometries, it is concluded that the original OCC design for Toyota diesel engine is better for mixing-controlled combustion; whereas, SCC is the most suitable piston design for RCCI combustion among the three selected geometries under the investigated operating conditions of the engine. With SCC, better combustion and performance can be achieved while maintaining relatively lower CO, NO and soot emissions. Ó 2016 Elsevier Ltd. All rights reserved.
1. Introduction Energy shortage and air pollution are the two major issues that intrigue the engine researchers to find advanced in-cylinder combustion with high efficiency and low emissions. RCCI (reactivity controlled compression ignition), which was proposed in 2009 by Kokjohn et al. [1], has been considered as an advanced combustion scenario. Compared to CDC (conventional diesel combustion), RCCI can reduce soot and NOx emissions to ultra-low levels. It has been investigated by adopting gasoline/diesel fuel combination by some researchers [2–7]. Recently, biodiesel, which is derived from plant or animal sources, and thus renewable, has been tested in CDC engines [8,9]. The oxygen-borne biodiesels can enhance the oxidation of fuels, and consequently bring more complete combustion to CDC engines. However, increased NOx emissions were usually observed in biodiesel fueled diesel engine due to the contained oxygen [9,10]. Therefore, RCCI technique can be applied to alter the combustion mode of CDC engines fueled with biodiesels, ⇑ Corresponding author. Tel.: +65 6516 6481. E-mail address: [email protected] (W.M. Yang). http://dx.doi.org/10.1016/j.enconman.2016.01.041 0196-8904/Ó 2016 Elsevier Ltd. All rights reserved.
consequently to further improve the thermal efficiency and meanwhile lower the NOx emissions. To realize the RCCI combustion in this study, biodiesel was chosen as the high reactivity fuel; whereas gasoline was selected as the low reactivity fuel. Our group has conducted one study on gasoline/biodiesel fueled RCCI combustion [11], where the effects of gasoline ratio and SOI timing were investigated. The results showed that less NOx emissions were generated in gasoline/biodiesel fueled RCCI combustion compared to CDC fueled with pure biodiesel. In this study, to further optimize the gasoline/biodiesel fueled RCCI combustion, the effect of piston bowl geometry was taken into consideration. The piston bowl plays an essential role in facilitating the mixing of air and fuel within the combustion chamber [12,13]. Splitter et al. [14] suggested to decrease the surface to volume ratio while designing the bowl geometry of RCCI engines for the purpose of reducing heat losses. Hanson et al. [2] compared the designed RCCI piston with original diesel piston. It was found that the RCCI piston, which has a lower surface to volume ratio, can provide 3% of increases in thermal efficiency with ultra-low levels of NOx and PM emissions at the engine speed of 2600 rpm and IMEP of 6.9 bar. Benajes et al. [15] also compared three bowls
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Table 1 Models in KIVA4–CHEMKIN. Name
Model
Turbulent model Spray model Breakup model Collision model Combustion model
RNG k—e model Discrete droplet model KH–RT model O’Rourke model CHEMKIN II solver
which are stock, stepped, and bathtub at the engine speed of 1200 rpm under low, medium and high loads. The stepped bowl was proved to be the most suitable piston for the tested RCCI engine. In the past investigation undertaken by our group [16], three types of bowl geometries, namely HCC (Hemispherical Combustion Chamber), SCC (Shallow depth Combustion Chamber) and OCC (Omega Combustion Chamber), were selected and tested at 1200, 2400 and 3600 rpm in a conventional diesel engine. In this study, the objective is to find the suitable piston for the RCCI engine which was modified from the diesel engine. It may be assumed that the piston bowl geometry designed for CDC may not be suitable for RCCI combustion. Moreover, the studies on bowl geometries of RCCI engines which have been carried out so far were operated at low engine speeds [2,14,15], at which the combustion process much relies on the piston surface to volume ratio [16]. The surface to volume ratio could significantly affect the heat transfer process and
(a) HCC
consequently play a dominant role in the combustion process. Thus, in this study, the gasoline/biodiesel fueled RCCI engine was operated at high engine speed to investigate whether the surface to volume ratio of a piston is still crucial for the combustion process. Conclusively, there is a need to study the effect of bowl geometry on a gasoline/biodiesel fueled RCCI engine at high engine speed. In this study, simulations with a newly developed mechanism were conducted based on the high speed (3600 rpm) operating condition in [16], 20% and 40% of gasoline was premixed while maintaining the same energy input. Moreover, the SOI (start of injection) timing was varied from 11 to 60 deg ATDC to transform the combustion mode from the mixing-controlled to the reactivity-controlled. Finally, the results on combustion characteristics including in-cylinder pressure, HRR (heat release rate), PRR (pressure rise rate) and combustion duration, engine performance and emissions formation on CO, NO and soot were compared.
2. Methodology 2.1. KIVA4–CHEMKIN A multi-component fuel combustion model [17], which coupled CHEMKIN II solver with KIVA-4 code, was used in this study. KIVA4 [18] is a computer code for the numerical calculation of the transient, two and three dimensional, multiphase, multicomponent chemically reactive flows with sprays. However, in the original
(b) SCC
(c) OCC
Fig. 1. Generated grids of bowl geometries.
Table 2 The engine specifications and operating conditions for validation cases. (a) Pure biodiesel Engine type Number of cylinders Bore, mm Stroke, mm Displacement, L Compression ratio Injector nozzle number Intake valve close, deg ATDC Exhaust valve open, deg ATDC
GW4D20 1 83.1 92 0.5 16.7 7 130 94
Engine speed, rpm Pressure at IVC, bar Temperature at IVC, K Biodiesel consumption, kg/h Gasoline consumption, kg/h SOI (ATDC)
1400 1.1 360 0.45 0.0 7.0
Piston bowl at TDC
(b) Gasoline/biodiesel
(c) Pure biodiesel Toyota 2KD-FTV 4 92 93.8 2.494 18.5 6 149 150
1400 1.10 362 0.26 0.24 9.0
3600 1.89 370 11.34 0.0 11.0
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361
(a) Pure biodiesel in GW4D20
Fig. 3. In-cylinder pressure and HRR of B80.
(b) Gasoline/biodiesel in GW4D20
Fig. 4. Temperature distribution and fuel spray of OCC with B80 and SOI timing of 11 deg at TDC.
Rayleigh–Taylor (KH–RT) hybrid model. RNG k–e model is selected for the consideration of turbulence.
(c) Pure biodiesel in Toyota 2KD-FTV Fig. 2. Validation of in-cylinder pressure and HRR of the pure biodiesel and gasoline/biodiesel combustion.
Table 3 Validation on NOx emissions. NOx (g/kW h)
Experiment Simulation
Case (a)
Case 2 (b)
2.29 3.29
1.61 1.86
KIVA-4, only a global reaction and 12 species were included. To realize the detailed chemistry calculation, CHEMKIN II was coupled with KIVA-4. Then a newly developed reaction mechanism by our group [11,19] was used to model the chemical kinetics of gasoline and biodiesel. In this mechanism, there are 107 species and 425 reactions. Among the species, iso-octane is considered as the surrogate of gasoline, and the mixture with 25% MD (methyldecenoate), 25% MD9D (methyl-9-decenoate) and 50% n-heptane in mole represents biodiesel. Table 1 lists the models used in this study. The original breakup model in KIVA4 was replaced with Kelvin–Helmholtz and
2.2. Piston bowl geometry The effect of three different pistons on a diesel engine fueled with pure biodiesel has been investigated in our previous study [16]. Same pistons were selected in this study to examine their effects on the gasoline/biodiesel fueled RCCI engine. Fig. 1 demonstrates the geometries and generated grids of the three pistons: HCC, SCC and OCC. The compression ratio is fixed at 18.5. From the previous study [16], it is known that during the compression stroke in the bowl, OCC can form the strongest squish flow, then followed by HCC; SCC generates the weakest among the three. The piston surface area of OCC is 10,456 mm2 which is the largest. HCC and SCC have similar surface areas which are 8118 and 8192 mm2, respectively. In addition, the throat diameter of SCC is the largest among the three, followed by HCC, and the shortest is OCC. 2.3. Validation To validate the models, experimental data were firstly compared with numerical results under two conditions, fueling
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(a) HCC
(b) SCC
(c) OCC
Fig. 5. Vector (Length(cm)/magnitude = 0.0002) for B80 with SOI timings of
gasoline/biodiesel and pure biodiesel, respectively. Table 2 lists the engine specifications and operating conditions for the validations. The experimental data was obtained from GW4D20 and Toyota 2KD-FTV. The experimental setup of GW4D20 is at Tsinghua University [20]. The in-cylinder pressure is measured by AVL GH14P sensor. The data is recorded in 150 consecutive cycles for combustion analysis. The NOx emission is measured using AVL CEB2 gas analyzer. The measurement on soot is conducted using AVL 439 opacimeter. The NOx and soot are averaged for 30s when the engine condition is stable. The Toyota 2KD-FTV engine setup is at National University of Singapore. The cylinder pressure was measured by AVL GH13P water-cooled pressure transducer with a resolution of 1 °CA. The measured data is obtained from averaged 50 cycles for each condition. The cooling water temperature is maintained at 70 °C. The NO and NO2 are measured by a portable gas analyzer (E instrument 4400 N). Fig. 2 compares the in-cylinder pressure and HRR between experiments and simulation for both pure biodiesel and gasoline/ biodiesel combustion. Though a slightly over-prediction on pressure and HRR can be observed, the discrepancies are still
(a) Temperature (Kelvin) iso-surface
60 deg ATDC at CA50.
acceptable. The abrupt HRR of the simulation on gasoline/biodiesel fueled combustion maybe due to the assumption on the homogeneously premixed gasoline during the simulation. In reality, the gasoline distribution cannot be homogeneous. Thus, the HRR of dual-fuel simulation is faster compared to experiment. Thus, the models and reaction mechanism is validated. The validation on emissions formation is conducted on GW4D20 engine (case (a) and case (b) in Table 3). As mentioned, the NOx and soot are measured by AVL CEB-II and AVL 439 opacimeter, respectively. Table 3 compares the NOx emissions under conditions of validation case (a) and (b). It can be seen that with increase in gasoline ratio, the NOx was reduced for both experiment and simulation cases. Thus, the model could predict the NOx formation trend by increasing the gasoline fraction. The experimental data on soot is not suitable to validate the simulation because the measured smoke by an opacimeter includes not only soot, but also solids and liquids etc. However, it is found that with the increase in gasoline, the predicted soot increases. This trend is consistent with that observed from experiment.
(b) Equivalence ratio iso-surface
Fig. 6. Iso-surfaces of (a) temperature and (b) equivalence ratio of three bowl geometries with B80 and SOI timing of
60 deg at CA50.
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2.4. Study cases As mentioned, the studies carried out so far were focused on the RCCI combustion at low engine speed where the surface area of a piston can play a crucial role on the combustion process [16]. To study the effect of piston bowl geometry on RCCI combustion at high engine speed, the validation case with pure biodiesel was selected as the base case. Thus, OCC is the original design piston for the Toyota 2KD-FTV engine. Then the total energy input remained constant, gasoline was premixed in the cylinder with the energy percentages of 20% and 40%, respectively; accordingly, the percentages of biodiesel decreased to 80% (B80) and 60% (B60). It should be noted that to further increase gasoline ratio, the abrupt peak pressure rise rate is unacceptable due to the absence of EGR (exhaust gas recirculation). Thus, the results with more premixed gasoline (40% above) will not be presented in this study. Moreover, the SOI timing of the base case is 11 deg ATDC, which results in the mixing-controlled combustion [21]. To form the reactivity-controlled combustion, the SOI timing was further advanced to 35 and 60 deg ATDC. Thus, for each bowl geometry, numerical simulations were conducted for B80 and B60, with SOI timings of 11, 35, and 60 deg ATDC, respectively.
(a) B80
3. Results and discussion 3.1. Combustion characteristics 3.1.1. In-cylinder pressure and heat release rate Fig. 3 shows the in-cylinder pressure and HRR of B80 with SOI timings of 11, 35 and 60 deg ATDC. As can be observed, with SOI timing of 11 deg, the peak pressures from the highest to the lowest are of OCC, HCC and SCC, which is associated with the intensity of squish flow formed in piston bowl. This is because the fuel spray targets at piston bowl area and subsequently the combustion happens in the bowl, as shown in Fig. 4. Thus, with late SOI timing, the combustion process is strongly dependent on the
(b) B60 Fig. 8. Maximum pressure rise rate.
Fig. 7. In-cylinder pressure and HRR of B60.
mixing process of fuel and premixed mixture, and the OCC can produce better combustion due to its strong squish formed in the bowl. By advancing SOI timing, the combustion mode gradually transits from the mixing-controlled to the reactivity-controlled [21]. With SOI at 35 deg ATDC, OCC still shows the highest peak pressure; however, HCC could not show superiority compared to SCC. Further advancing SOI timing to 60 deg ATDC, SCC can achieve a slightly higher peak pressure than HCC and OCC. This indicates that the squish flow in piston bowl cannot prevail the combustion process when the mode is reactivity-controlled. This is because for the RCCI combustion with SOI of 60 deg ATDC, the fuel spray will target at squish region and part of the fuel will deposit on the piston surface and cylinder liner. The slightly higher peak pressure of SCC may be due to its shallow depth and wider throat diameter by which the squish flow formed in the bowl can be pushed much far away from the centerline of the bowl, as shown in Fig. 5, which displays the vector of squish flow. Fig. 6 displays the iso-surfaces of temperatures at 1300, 1500, 2000 K and equivalence ratio of 0.14, 0.5 and 1.0 for three bowl geometries with B80 and SOI timing of 60 deg at CA50. CA50 is the crank angle where 50% of heat is released. It can be seen that due to the early injection, high equivalence ratio is observed near
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Fig. 10. In-cylinder average temperature profile with B80 and SOI timing of 60 deg ATDC.
(a) B80
reducing. When SOI timing is advanced to 35 deg ATDC, the peak pressure of OCC reduces. With SOI timing of 60 deg ATDC, same trend on peak pressure is observed with those of B80. SCC can achieve the highest peak pressure during RCCI combustion.
(b) B60 Fig. 9. Combustion duration with SOI timings of
11,
35,
60 deg ATDC.
the cylinder liner (red1 surface in Fig. 6(b)). SCC bowl has a shallow depth and larger diameter, as such, the far-away squish flow formed in the bowl could be more effectively directed to the near-wall region and entrain part of the biodiesel in the bowl (higher equivalence ratio in Fig. 6(b)), thereby enhancing the fuel mixing process in this region subsequently resulting in better combustion (Fig. 6(a)). This also explains the higher HRR of SCC compared to HCC and OCC. Moreover, the highest peak pressure of SCC may be also partly due to the reduced heat losses caused by the less surface area. Also, it is worth noting that by advancing SOI timing from 11 to 60 deg ATDC, the combustion mode transits from slow single-stage combustion to fast single-stage combustion and then two-stage combustion [22]. Fig. 7 displays the in-cylinder pressure and HRR of B60 with SOI timings of 11, 35 and 60 deg ATDC. Compared to B80, B60 can create relatively higher reactivity difference. With SOI timings of 11 deg ATDC, the peak pressures are still following the trend as shown in B80 which indicates the combustion is still mixingcontrolled; however, the gaps among the peak pressures are
1
For interpretation of color in Fig. 6, the reader is referred to the web version of this article.
(a) B80
(b) B60 Fig. 11. IMEP with SOI timings of
11,
35,
60 deg ATDC.
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(a) B80
(a) B80
(b) B60
(b) B60
Fig. 12. CO emissions with SOI timings of
11,
35,
60 deg ATDC.
3.1.2. Maximum peak pressure rise rate Fig. 8 compares the maximum PRR. The acceptable limit of maximum PRR is chosen as 10 bar/deg. As shown in the figure, only B80 with SOI timing of 11 and 60 deg ATDC can satisfy the criterion. The other operating conditions would suffer the unacceptable engine knocking. The high maximum PRR for B80 with SOI timing of 35 deg ATDC is caused by the fast single heat release combustion. When the premixed gasoline increases to 40%, i.e., B60, the maximum PRRs all exceed 10 bar/deg regardless of the SOI timings. The fast combustion with higher gasoline ratio is due to the fast
(a) HCC
Fig. 14. NO emissions with SOI timings of
11,
35,
60 deg ATDC.
combustion resulted from gasoline–air mixture. Also, it should be noted that EGR is not applied to all the cases in this study.
3.1.3. Combustion duration Fig. 9 shows the combustion duration for (a) B80 and (b) B60 with SOI timings of 11, 35, 60 deg ATDC. Combustion duration is defined as the interval between CA5 and CA90, which are the crank angles when 5% and 90% of heat release are completed, respectively. From Fig. 9(a) and (b), it can be seen that with SOI
(b) SCC
Fig. 13. Spatial CO distribution and vector (Length(cm)/magnitude = 0.0001) for B80 with SOI timings of
(c) OCC 35 deg ATDC at crank angle of 25 deg.
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timing of 11 deg ATDC, the combustion duration has a negative correlation with the squish strength during compression stroke in the bowl. SCC which has the weakest squish flow can achieve the longest combustion duration. However, when SOI timing is advanced to 60 deg ATDC, the combustion duration of HCC and OCC can be extended to more than 100 degree; in contrast, the duration of SCC is about 60 degree. The relatively short combustion duration with SOI timing of 11 deg ATDC is due to the mixingcontrolled combustion as mentioned. The stronger squish flow in OCC results in a better mixing process between biodiesel and gasoline–air mixture, subsequently completing combustion at a rapid rate. When the combustion mode transits to the reactivitycontrolled, SCC achieves relatively short combustion duration, though this duration is longer compared to that of all the pistons with SOI timings of 11 and 35 deg ATDC. This short combustion duration of SCC with SOI of 60 deg ATDC can be attributed to the stronger squish flow far away from the centerline caused by the wide throat diameter. This far-away strong flow can be effectively directed to the near wall region, providing better mixing as well as better combustion and high temperature. The higher temperature leads to the fast consumption of fuel and consequently brings short combustion duration. In addition, the increased in-cylinder temperature caused by less heat loss of SCC leads to shorter combustion duration. Fig. 10 shows the in-cylinder average temperature profile with B80 and SOI timing of 60 deg ATDC. It can be observed that the temperature of SCC is higher compared to HCC and OCC, which could consolidate the inference. 3.2. Engine performance Fig. 11 displays the IMEP (indicated mean effective pressure) for (a) B80 and (b) B60 with SOI timings of 11, 35 and 60 deg ATDC. As can be observed from Fig. 11(a) and (b), with B80, OCC and HCC, which have stronger squish flow, can generate the highest IMEP when SOI timing is 11 deg ATDC. This result can be inferred from the peak pressure observed in Fig. 3. Thus, the strong squish flow of a bowl geometry could achieve better performance for a mixing-controlled combustion. However, when the
SOI timing is advanced to 60 deg ATDC, the best performance is observed for SCC. The degraded performance for OCC and HCC may be due to the longer combustion duration.
3.3. Emissions formation 3.3.1. CO emissions Fig. 12 compares the CO emissions for B80 and B60 with SOI timings of 11, 35, 60 deg ATDC. Similar trends are observed for B80 and B60. With SOI timing of 11 deg ATDC, CO levels for all three pistons are almost the same. With advancing SOI timing to 35 deg ATDC, CO of HCC significantly increases whereas the other two remains at the low levels. Further advancing SOI to 60 deg ATDC, both HCC and OCC generates high CO emissions. Fig. 13 shows the spatial CO distribution and the vector for B80 with SOI timings of 35 deg ATDC at the crank angle of 25 deg. The CO of HCC and SCC are mainly generated in the bowl region; in contrast, the CO of OCC is mainly at squish region. This indicates that the combustion of HCC and SCC happens in the bowl and that of OCC happens in the squish region. This is because when SOI timing is 35 deg ATDC, the fuel spray still targets at bowl region of HCC and SCC due to their wider throat diameter compared to OCC. However, the CO of HCC generated in the bowl could not be further oxidized due to the weak mixture flow as implied from the length of the vector, as shown in Fig. 13(a). In contrast, the CO generated in SCC can be further oxidized because of the stronger flow during combustion stroke. With SOI timing of 60 deg ATDC, the fuel spray targets at squish regions for all three pistons. The high CO for HCC and OCC could be attributed to the much extended combustion duration (above 100 deg CA) as shown in Fig. 9. When the combustion duration is very long, the in-cylinder temperature at the end of the combustion process could be much lower which is unfavorable for the oxidation of CO emissions. Thus, compared to HCC and OCC, SCC can achieve low CO emissions with the relatively short combustion duration caused by the far-away squish flow effect in the squish region due to the wide throat diameter.
(a) OCC at CA50 Fig. 15. Phi–T map for B60 with SOI timings
(b) SCC at CA50 60 deg ATDC.
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3.3.2. NO emissions Fig. 14 shows the NO emissions for B80 and B60 with SOI timings of 11, 35, 60 deg ATDC. With B80, when the SOI timing is 60 deg ATDC, similar levels of NO emissions were found for the three piston bowls. Significant high NO can be seen for OCC piston bowl when the SOI timings are 11 and 35 deg ATDC. This high NO is resulted from the high temperature as implied from the peak HRR shown in Fig. 3. Similarly, with B60, the NO levels with SOI timings are associated with the peak pressure observed in Fig. 7. It is interesting to notice that when the combustion is reactivity-controlled, OCC generates relatively lower NO. The possible reason could be the relatively poorer combustion of OCC as implied from the pressure histories. Another reason could be the large surface area of OCC contributes to the heat losses which would subsequently lower the in-cylinder temperature. The incylinder temperature can be demonstrated in Fig. 15, which is the /–T (Equivalence ratio–Temperature) diagram [11] for OCC and SCC when the SOI timing is 60 deg ATDC with B60. The overall temperature distribution of OCC is lower compared to that of SCC.
(a) B80
3.3.3. Soot emissions Fig. 16 displays the soot emissions for B80 and B60 with SOI timings of 11, 35, 60 deg ATDC. It can be observed that there is not too much difference on the soot emissions for the three bowl geometries. Although the results are close to each other, OCC demonstrates a relatively lower soot under most of the conditions. The reason could be implied from Fig. 17. Fig. 17 shows the spatial temperature and equivalence ratio distribution for B60 with SOI timings of 60 deg ATDC at CA50. The vector in the cylinder is also presented along with the temperature distribution. It can be seen that OCC can generate strong flow near the cylinder liner compared to HCC and SCC, thus providing a better mixing process and subsequently resulting in local low equivalence ratio regions. Consequently, OCC generates relatively lower soot emissions compared to HCC and SCC. 4. Conclusion
(b) B60 Fig. 16. Soot emissions with SOI timings of
(a) HCC
11,
35,
60 deg ATDC.
In this study, three bowl geometries, namely HCC, SCC and OCC, were numerically examined in a gasoline/biodiesel fueled RCCI
(b) SCC
Fig. 17. Spatial temperature with vector (Length(cm)/magnitude = 0.0002) and equivalence ratio for B60 with SOI timings of
(c) OCC 60 deg ATDC at CA50.
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engine at low, medium and high loads with engine speed of 3600 rpm. While changing the gasoline and biodiesel ratios, the total energy input was kept constant. In addition, the SOI timing should be varied during simulation to achieve reactivitycontrolled combustion. Finally, engine simulations were conducted for each bowl geometry with 80% and 60% of direct-injected biodiesel while adopting SOI timings at 11, 35 and 60 deg ATDC, respectively. The results on combustion characteristics, engine performance and emissions were compared. The major conclusions can be drawn as follows: (1) Regarding the combustion characteristics, the original designed OCC for diesel engine is better for the mixingcontrolled combustion due to its strong intensity of squish flow in the piston bowl during compression stroke. However, it is not suitable for the reactivity-controlled combustion due to its narrowed squish flow effect and its higher heat losses caused by larger surface area. In contrast, SCC is better for RCCI combustion because of the relatively short combustion duration. Thus, the throat diameter and surface area of a piston bowl can play a crucial role in affecting the combustion process at high engine speed. (2) For RCCI combustion, in terms of emissions, SCC can achieve low CO emissions which indicates a higher combustion efficiency; however, the NO emissions of SCC are relatively higher than that of OCC because of the overall higher combustion temperature. Meanwhile, similar levels of soot emissions were observed for three bowl geometries, though a slightly lower soot can be found for OCC. (3) Overall, SCC is the best piston design for RCCI combustion among the three selected geometries under the investigated operating conditions of the Toyota engine in this study. To use SCC, better combustion process and performance can be achieved while maintaining relatively lower CO, NO and soot emissions. Acknowledgement This work is supported by the program of China Scholarship Council (CSC). Appendix A. Supplementary material Supplementary data associated with this article can be found, in the online version, at http://dx.doi.org/10.1016/j.enconman.2016. 01.041. References [1] Kokjohn SL, Hanson RM, Splitter DA, Reitz RD. Experiments and modeling of dual-fuel HCCI and PCCI combustion using in-cylinder fuel blending. SAE Int J Engine 2010;2:24–39.
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