diesel RCCI engine

Accepted Manuscript Effects of piston bowl geometry on combustion and emissions characteristics of a natural gas/diesel rcci engine Amir-Hasan Kakaee, Babak Partovi, Amin Paykani, Ali Nasiri-Toosi PII: DOI: Reference:

S1359-4311(16)30463-X http://dx.doi.org/10.1016/j.applthermaleng.2016.03.162 ATE 8030

To appear in:

Applied Thermal Engineering

Received Date: Accepted Date:

31 December 2015 30 March 2016

Please cite this article as: A-H. Kakaee, B. Partovi, A. Paykani, A. Nasiri-Toosi, Effects of piston bowl geometry on combustion and emissions characteristics of a natural gas/diesel rcci engine, Applied Thermal Engineering (2016), doi: http://dx.doi.org/10.1016/j.applthermaleng.2016.03.162

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Effects of Piston Bowl Geometry on Combustion and Emissions Characteristics of a Natural Gas/Diesel RCCI Engine Amir-Hasan Kakaee, Babak Partovi, Amin Paykani*, Ali Nasiri-Toosi Vehicle Powertrain System Research Laboratory, School of Automotive Engineering, Iran University of Science and Technology, Narmak, Tehran, Iran *

Corresponding author: [email protected]

Abstract Piston bowl geometries are crucial to the combustion and emission characteristics of reactivity controlled compression ignition (RCCI) engines. The present numerical study explores the effects of piston bowl geometry on natural gas/diesel RCCI performance and emissions at medium engine load. Three different piston bowl geometries including stock, bathtub and cylindrical with constant compression ratio 16.1:1 are selected using double injection strategy and Influences of engine speed, piston bowl depth and chamfered ring-land are investigated. It is found that the bowl profile does not affect combustion of RCCI engine at low engine speeds, but it has much considerable effect at higher engine speeds. The results obtained also show that bathtub design yields the best performance and emissions at higher speeds. It is also reported that both piston bowl depth and chamfered ring-land can also affect engine-out emissions specially UHC and CO emissions.

Keywords: Reactivity controlled compression ignition (RCCI); natural gas; efficiency; emissions; piston bowl geometry

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1. Introduction In last decades, engine researchers and manufacturers have focused their efforts on development of new combustion strategies and aftertreatment systems to abate the pollutant emissions associated to internal combustion engines. It has been shown that simultaneous reduction in NOx and PM can be achieved with the modern combustion strategies such as low temperature combustion (LTC) which is characterized by significantly lower combustion temperatures than conventional strategies [1]. There are various LTC strategies, including homogeneous charge compression ignition (HCCI), premixed charge compression ignition (PCCI) and reactivity controlled compression ignition (RCCI) [2-11]. In the HCCI and PCCI strategies, by premixing a significant portion of the fuel and relying on long ignition delay times, the high temperature flame fronts can be avoided which inhibits NOx formation, reduces heat transfer losses, and the higher ratio of specific heats can further increase efficiency. However, new challenges regarding combustion phasing control and NVH problems were identified [7]. Bessonette et al. [12] found that the best fuel for HCCI operation has autoignition qualities between that of diesel fuel and gasoline, depending on the operating condition. A strategy was also introduced by Inagaki et al. [13] for controlling PCCI combustion by means of varying fuel reactivity. It was found that the combustion phasing can be controlled by the ratio of the two fuels. Recently, reactivity controlled compression ignition (RCCI) engine aiming at controlling the combustion phasing and engine knock by forming stratified combustion has been proposed by Kokjohn et al. [14] to solve the difficulties of combustion timing control over wide range of engine loads faced by HCCI and PCCI engines. RCCI relies on the stratification versus homogenization by direct in-cylinder blending different reactivity fuels, hence controlling the combustion phasing. To generate the reactivity in cylinder, fuel with low reactivity, such as

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gasoline or natural gas, is injected from the port, whereas a fuel with high reactivity, such as diesel or biodiesel, is injected into the cylinder through an injector [15-22]. Combustion chamber geometry plays an important role in boosting the in-cylinder air motion which in turn helps in air-fuel mixing and combustion processes. Since optimum combustion chamber design can suppress pollutant emissions formation, without affecting seriously the RCCI engine performance, the effect of combustion chamber geometry on RCCI engine performance and exhaust emissions has been investigated in literature. Splitter et al. [23] studied effects of compression ratio and piston design in a heavy-duty diesel engine operated with RCCI combustion at various loads. They found that the bathtub style piston offers the highest brake efficiency and to enable low emissions and low pressure rise rates with practical intake and EGR temperatures. Moreover, a piston design with a short squish distance and large bowl volume reduced engine out hydrocarbon emissions. Hanson et al. [24] investigated effects of optimized piston geometry in RCCI operation in a light duty multi-cylinder engine at different conditions. They reported that the higher HC emissions from the optimized RCCI piston compared to experimental results from the equivalent singlecylinder engine and heavy-duty RCCI optimized piston suggests that piston squish area may not be the primary source of UHC. Instead, the piston-to-liner crevice volume may be a significant source. Effect of modified piston featuring a shallow, flat piston bowl with nearly no squish land in RCCI operation with gasoline/diesel and methanol/diesel fuels was studied by Dempsey et al. [25]. It was found that using the modified piston, the GIE of RCCI combustion was significantly improved at light loads due to increases in combustion efficiency and decreases in heat transfer losses. At higher loads the modified piston also performed better than the stock piston, but the improvements were not as significant. Lim and Reitz [26] used chamfered piston crown design in order to reduce UHC emissions from the ring-pack crevice of RCCI engine. They found that UHC and CO emissions were reduced by

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79% and 36%, respectively, achieving 99.5% combustion efficiency. Benajes et al. [27] experimentally investigated the effects of piston bowl geometry on RCCI performance and emissions at low, medium and high engine loads. Results suggested that the considered geometries enable ultra-low NOx and soot emissions at low and medium load when using double injection strategies. But, unacceptable emissions were measured at high load taking into account EURO VI limitations. They later studied effects of piston bowl geometry on heat transfer and combustion losses in a RCCI engine at different engine loads. They found that heat flux through the piston surface represent the major portion of the heat transfer energy. They also showed that reduced piston surface area and reduced charge motion, are the key factors to improve the gross indicated efficiency (GIE) over the different engine loads [28]. Li et al. [29] numerically investigated the effects of three bowl geometries on a gasoline/ biodiesel fueled RCCI engine operated at high engine speed. It was demonstrated that SCC (Shallow depth Combustion Chamber) is the best piston design for RCCI combustion among the three selected geometries under the investigated operating conditions. Natural gas is an interesting alternative to conventional liquid fuels (i.e., gasoline and diesel) for use in automotive engines [30-33]. Due to the higher octane number of natural gas compared to gasoline, it can be considered as a better alternative port fuel for RCCI combustion. There have been a few research works published regarding application of natural gas in RCCI combustion. Nieman et al. [34] used the KIVA-3V CFD code in conjunction with the CHEMKIN chemistry tool and the NSGA II algorithm to perform optimization for natural gas/diesel RCCI combustion from low load to high load, adopting a diesel double injection strategy. They found that the double injection strategy was more favorable for an RCCI engine to realize high efficiency and low emissions, while sweeping from low to high load.

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As can be seen from above, rare studies have been carried out to disclose the effect of piston bowl geometry in RCCI engine. In addition, no study has been conducted on the effects of bowl geometry on the combustion and emission characteristics of a RCCI engine fueled with natural gas/diesel. To bridge the gap, the present paper is studying effect of piston bowl geometry on natural gas/diesel RCCI combustion and emissions at medium engine load. Three different piston bowl geometries are selected and the effects of engine speed, piston bowl depth and chamfered ring-land on the combustion and emission characteristics of RCCI engine are studied.

2. Models and validation 2.1. Computational model In this work, the CONVERGE CFD tool [35] is used as the computational framework for natural gas/diesel RCCI engine simulations. Closed-cycle computations on sector grids with periodic boundaries were carried out to improve computational efficiency. Since the direct injector has six evenly distributed holes, a 60-degree sector geometry was created, as shown in Fig. 1. The CFD code uses a structured cartesian grid with base cell size of 1.4 mm, an adaptive mesh refinement (AMR), as well as a fixed refinement within the spray region. The port fuel injected natural gas is considered to be homogeneously mixed with air at IVC and the diesel injection process is simulated by the standard Droplet Discrete Model (DDM) [36]. Spray atomization and break-up are modeled using the hybrid KH-RT model [37]. The No Time Counter (NTC) collision method [38] is used in the present simulations to consider droplets collision. In general, the NTC sub-model has a computational cost that is linear with the number of spray parcels, while conventional models have a cost that increases with the square of the number of parcels. Turbulent flow is modeled by means of the RNG (ReNormalization Group) k   model with wall-functions [39] in order to account for wall heat

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transfer. Concerning combustion modeling, a direct integration of detailed chemistry approach was used by means of the CONVERGE code and the SAGE solver. The SAGE model uses the CVODES solver package. SAGE calculates the reaction rates for each elementary reaction while the CFD solves the transport equations. The physical properties of diesel fuel are represented by tetradecane (C14H30) as a fuel surrogate for the spray and mixing processes. To speed up the runtime for all cases, the simulations are performed using the multi-zone chemistry solver [40]. The multi-zone solver considerably speeds up combustion calculations by grouping cells with similar properties into zones in the chemistry calculations. A mechanism composed of 76 species and 464 reactions is used [41]. Soot is predicted using a phenomenological soot model [42] based on the approach of Hiroyasu [43]. The NOx formation is modeled using an extended Zeldovich mechanism. This mechanism includes three reactions and seven species and is able to predict NO formation with high accuracy over a fairly wide range of equivalence ratios [42]. 2.2. CFD model validation The numerical results have been validated against the results of Nieman et al. [34]. Table 1 shows the relevant parameters for the single-cylinder Cat® 3401E SCOTE test engine being simulated. Validations were conducted at 4, 9 and 23 bar indicated mean effective pressure (IMEP) from inlet valve closing (IVC = -143 ATDC) until the exhaust valve opens (EVO =130 ATDC) with 5%, 0% and 48% exhaust gas recirculation (EGR), respectively. The operating conditions for validation of simulation results are given in Table 2. Fig. 2 shows predicted cylinder pressure histories at different loads compared to Ref. [34]. As shown, the pressure traces and the combustion phasing were well-predicted for all three cases. Additionally, as shown in Fig. 3, the model successfully captured the soot, NOx, CO, and HC emission trends at engine low and medium loads. As can be seen, low and mid-load

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engine operations have almost acceptable values, however the prediction of engine-out emissions at high load operation (High EGR rates~48%) is significantly differs from the results of Ref. [34] which is likely due to simplifications in EGR modeling. 3. Results and Discussion 3.1. Effect of engine speed In this section, three different piston bowl profiles are selected and the effect of engine speed on combustion and emissions characteristics of natural/gas RCCI engine is studied. Fig. 4 shows the selected bowl profiles including stock, bathtub and cylindrical for heavy-duty Caterpillar engine. The parametric study is carried out at engine speeds of 800, 1300 and 1800 rpm. Fig. 5 illustrates the in-cylinder pressure, heat release rate (HRR) and temperature histories for three piston bowl profiles at different engine speeds. Efficiency and emissions comparisons are provided in Fig. 6. The results obtained showed that by increasing engine speed, pressure and temperature are reduced and combustion phasing is delayed. Higher engine speeds not only allows less time for chemical reactions to occur which leads to more incomplete combustion, but also results in less time for heat transfer, which may cause higher in-cylinder temperatures. Generally, the combination effect of these two factors results in reduced efficency and ringing intensity (RI), and increased UHC and CO emissions [6]. In additon, it was found that the bowl profile does not affect pressure and HRR of RCCI engine at low speeds, but it has considerable effect at higher engine speeds. As can be seen in Fig. 5, the bathtub profile yields highest pressure and temperature with advanced combustion phasing (CA50) at engine higher speeds. As can be seen in Fig. 6, the bowl profile greatly affects NOx emissions but it has less effect on UHC and CO emissions. In addition, combustion chamber geometry has significant influence on squish flow formation, evaporation and mixing processes. It is obvious that the stock and cylinderical bowl profiles

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have narrower entrance area to the deeper part of the bowl compared to bathtub design which consequently enhances the velocity of inlet air to the bowl and forms a robust squish flow (see Fig. 7). On the other hand, bathtub bowl has lower bowl depth and piston area compared to two other designs, having lower heat transfer loss and eventually higher in-cylinder pressure, temperature and efficiency. As the piston bowl depth reduces, heat losses lowers, so the mean temperature elevates which results in higher NOx emissions. Overally, it can be concluded that the bathtub design has the best performance and emissions at higher speeds compared to two other designs. Fig. 8 depicts the temperature cut planes at three crank angles for different piston bowl profiles and various engine speeds. It is evident that by increasing engine speed, temperature is higher in the squish and ring pack crevice of bathtub profile. 3.2. Effect of piston bowl depth In this section, the cylindrical piston is selected for the bowl depth study. Four different bowl depths are considered ranging from 0.5 cm to 1.9 cm. By increasing bowl depth, squish height is decreased to keep compression ratio constant at 16.1. The operating condition corresponds to Table 2. Fig. 9 shows the in-cylinder pressure, HRR and temperature histories for different piston bowl depths. Efficiency and emissions comparisons are depicted in Fig. 10. As can be seen, by increasing bowl depth, in-cylinder pressure and temperature are increased. Gross indicated efficiency (GIE) is also increased up to the 1 mm bowl, but then it is decreased. The main factors involving in this behavior are combustion efficiency, heat transfer, piston bowl area and squish region. By increasing bowl depth, heat losses are increased, since the area to volume ratio increases. On the other hand, combustion efficiency is also improved. Generally, the combination effects of these factors result in slight increase of GIE in bowl depth of 1 mm.

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Since, almost all the UHC emissions are located in the squish region, it is reasonable that there are function of squish volume, therefore UHC emissions should be reduced by decreasing squish volume. But as can be seen in Fig. 8, this trend is valid up to the bowl depth of 1 mm, and after that UHC emissions are increased. To justify this, Fig. 11 shows cylinder temperature contours at 3 crank angles for different piston bowl depths. It is clear that, the high temperature region has penetrated more to the squish region in bowl depth of 1 mm. By decreasing bowl depth (i.e., increasing squish height), more amount of high reactivity fuel enters the squish region, but the squish flow tends to push away the fuel from this region. On the other hand, the fuel reactivity distribution is different for various bowl depths in piston liner, and this difference can cause UHC emissions to vary. By increasing squish height, area to volume ratio of squish is reduced (i.e., volume is increased), and heat losses to the piston top locally decreases, and allows high temperature gas to reach the liner, and oxidize much of the fuel, and thus reduces UHC emissions. Therefore, in order to minimize it, the high temperature gas in the squish region should ideally oxidize the trapped fuel at regions near liner or ring pack crevice. Finally, it can be concluded that reduction of UHC emissions depend on existence of reactive fuel and local heat transfer. In addition, the best piston bowl depth in terms of UHC and CO emissions is the bowl with 1 mm depth. 3.3. Effect of chamfered ring-land Here, the stock piston is taken into account for exploring the chamfer size in ring-land. Three chamfer sizes including 3, 6 and 9 mm are selected. By changing chamfer size, squish height is also changed to keep compression ratio constant at 16.1. The operating condition corresponds to Table 2. Fig. 12 illustrates the in-cylinder pressure, HRR and temperature histories for different chamfer size. Efficiency and emissions comparisons are provided in Fig. 13. 9

It is obvious that by increasing chamfer size, in-cylinder pressure and temperature are increased, GIE slightly rises and NOx emissions enhances. But, CO emissions slightly reduce which can be attributed to the varying reactivity of the fuel in squish region. As can be seen in Fig. 14, by using chamfered ring-land the total volume of squish region is increased and piston with 6 mm chamfer approximately has 35% higher squish volume compared to stock piston at TDC. In addition, using chamfered ring-land can cause reductions in UHC emissions particularly at chamfer sizes higher than 3 mm. The reason for this behavior can be explained by temperature contours (see Fig. 15). As can be seen, temperature is higher in ring top land in the case of chamfered piston. It is due to the reduction in area to volume ratio of this region compared to stock piston design, which consequently reduces heat transfer losses. Higher temperature in this region expedites chemical reactions and ultimately decreases UHC emissions in crevice [26]. 4. Conclusion In the present study, the influence of piston bowl geometry on natural gas/diesel RCCI combustion was investigated by means of computational modeling. Three different piston bowl geometries including stock, bathtub and cylindrical were selected to explore the effects of engine speed, piston bowl depth and chamfered ring-land on RCCI engine combustion. Main conclusions can be drawn as follows: 

By increasing engine speed, in-cylinder pressure and temperature were reduced and combustion phasing was delayed which led to reduced GIE and RI, and increased UHC and CO emissions.



It was found that the bowl profile did not affect combustion of RCCI engine at low enigne speeds, but it greatly affected it at higher engine speeds. Bowl profile considerabley affected NOx emissions, but it had negiligble effect on UHC and CO emissions. 10



By increasing bowl depth, in-cylinder pressure and temperature were increased. GIE was also increased up to the 1 mm bowl, but then it was reduced. In addition UHC and CO emissions were minimum at bowl depth of 1 mm.



By increasing chamfer size, in-cylinder pressure and temperature are increased, GIE slightly rises and NOx emissions enhances. Using chamfered ring-land can cause reductions in UHC emission particularly at chamfer sizes higher than 3 mm.

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Figures

Fig. 1 Piston bowl geometry of the Cat® engine

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Fig. 2 The validation case pressure traces and HRR

18

Fig. 3 Validation case emissions

19

Fig. 4 Selected piston bowl profiles

20

a

b

c

Fig. 5 Comparison of pressure, HRR and temperature histories for thee piston bowl profiles at engine speeds of: a) 800 rpm, b) 1300 rpm, c) 1800 rpm

21

Fig. 6 Comparison of performance and emissions for thee piston bowl profiles at different engine speeds

22

cylindrical

bathtub

stock

-200 ATDC

00 ATDC

200 ATDC

Fig. 7 Cylinder velocity cut planes at 3 crank angles for different piston bowl profiles at 9 bar IMEP and 1300 rpm

23

cylindrical

bathtub -200 ATDC

00 ATDC

100 ATDC

800 RPM -200 ATDC

00 ATDC

100 ATDC

1300 RPM -200 ATDC

24

stock

00 ATDC

100 ATDC

1800 RPM

Fig. 8 Cylinder temperature cut planes at 3 crank angles for different piston bowl profiles and various engine speeds

25

Fig. 9 Comparison of pressure, HRR and temperature histories for different piston bowl depths

26

Fig. 10 Effect of piston bowl depth on GIE, RI, HC and CO emissions

27

bowl depth 0.5cm

bowl depth 1 cm bowl depth 1.55 cm -200 ATDC

bowl depth 1.9 cm

00 ATDC

200 ATD

Fig. 11 Cylinder temperature cut planes at 3 crank angles for different piston bowl depths

28

Fig. 12 Comparison of pressure, HRR and temperature histories for different chamfer size

29

Fig. 13 Effect of chamfer size on GIE, RI, HC and CO emissions

30

Fig. 14 Combustion chamber profile at TDC with the squish space highlighted (right: stock piston; left: piston with 6 mm chamfer)

31

stock piston

piston with 6 mm chamfer -200 ATDC

00 ATDC

200 ATDC

Fig. 15 Cylinder temperature cut planes at 3 crank angles for stock piston and piston with 6 mm chamfer

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Tables

Table 1 Cat® 3401E SCOTE engine geometry [34] Displacement 2.44 L Bore × Stroke (cm) 13.72×16.51 cm Connecting rod length (cm) 26.16 cm Compression ratio 16.1:1 Swirl ratio 0.7 Bowl type Mexican Hat Number of valves 4 0 Intake valve opening 335 ATDC Intake valve closing -1430 ATDC Exhaust valve opening 1300 ATDC Exhaust valve closing -3350 ATDC Common rail diesel fuel injector Number of holes 6 Hole diameter ( m ) 250 Included spray angle (deg) 145

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Table 2 Validation case summary for RCCI engine [34] Case IMEP (bar) Engine speed (RPM) Intake pressure (bar)

1 4 800 1

2 9 1300 1.45

3 23 1800 3

Intake temperature ( C )

60

60

60

Total fuel mass (mg) Premix fuel equivalence ratio (-) Methane mass (%) Diesel SOI 1 (deg ATDC) Diesel SOI 2 (deg ATDC) Fraction of diesel in 1 st injection (-) Diesel Injection pressure (bar) EGR (%)

40 0.35 73 -52.9 -22.5 0.52 1300 5

89 0.35 85 -87.3 -38.3 0.4 954 0

228 0.35 85 -92.7 -20.4 0.7 742 48

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Highlights 

Effect of piston bowl geometry on natural gas/diesel RCCI combustion is studied.



Bowl profile did not affect combustion of RCCI engine at low enigne speeds.



Bathtub design yields the best performance and emissions at higher speeds.



Piston bowl depth and chamfered ring-land affected RCCI’s UHC emissions.

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