An Investigation of the Effects of the Piston Bowl Geometries of a Heavy-Duty Engine on Performance and Emissions Using Direct Dual Fuel Stratification Strategy, and Proposing Two New Piston Profiles [13/3]

Table of contents :
10.4271/03-13-03-0021: An Investigation of the Effects of the Piston Bowl Geometries of a Heavy-Duty Engine on Performance and Emissions Using Direct Dual Fuel Stratification Strategy, and Proposing Two New Piston Profiles
10.4271/03-13-03-0021: Abstract
10.4271/03-13-03-0021: Keywords
Introduction
Material and Methods
Computational Model
Model Validation
Results and Discussion
A Comparative Study for Different Piston Profiles under DDFS Strategy
The Effects of Shifting the Gasoline Injector Position to the Center of Head Cylinder Using Stock Piston
Proposed Geometry Profiles
Summary/Conclusions
References

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ARTICLE INFO Article ID: 03-13-03-0021 © 2020 SAE International doi:10.4271/03-13-03-0021

An Investigation of the Effects of the Piston Bowl Geometries of a HeavyDuty Engine on Performance and Emissions Using Direct Dual Fuel Stratification Strategy, and Proposing Two New Piston Profiles Sasan Shirvani,1 Saeid Shirvani,1 Amir H. Shamekhi,1 and Rolf D. Reitz2 1

KN Toosi University of Technology, Iran University of Wisconsin Madison, USA

2

Abstract Direct dual fuel stratification (DDFS) strategy benefits the advantages of the RCCI and PPC strategies simultaneously. DDFS has improved control over the heat release rate, by injecting a considerable amount of fuel near TDC, compared to RCCI. In addition, the third injection (near TDC) is diffusion-limited. Consequently, piston bowl geometry directly affects the formation of emissions. The modified piston geometry was developed and optimized for RCCI by previous scholars. Since all DDFS experimental tests were performed with the modified piston profile, the other piston profiles need to be investigated for this strategy. In this article, first, a comparative study between the three conventional piston profiles, including the modified, stock, and scaled pistons, was performed. Afterward, the gasoline injector position was shifted to the head cylinder center for the stock piston. NOX emissions were improved; however, soot was increased slightly. The other emissions, in-cylinder pressure, and AHRR remained unchanged. Finally, the advantages of modified and stock pistons were combined, and two new piston profiles based on the effective geometrical parameters were proposed and investigated. The first-proposed piston profile offered better NOX and CO emissions compared to the other profiles. In addition, the gross thermal efficiency of this profile is at high levels.

History Received: 20 Dec 2019 Revised: 13 Feb 2020 Accepted: 25 Feb 2020 e-Available: 16 Mar 2020

Keywords Piston profile, Piston bowl geometry, Numerical study, CFD simulation, DDFS engine

Citation Shirvani, S., Shirvani, S., Shamekhi, A., and Reitz, R., “An Investigation of the Effects of the Piston Bowl Geometries of a Heavy-Duty Engine on Performance and Emissions Using Direct Dual Fuel Stratification Strategy, and Proposing Two New Piston Profiles,” SAE Int. J. Engines 13(3):311-332, 2020, doi:10.4271/03-13-03-0021. ISSN: 1946-3936 e-ISSN: 1946-3944



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Introduction

N

owadays, governments are imposing stringent regulations on engines’ emissions to tackle global warming effects. Conventional diesel engines have high NOX and soot emissions, and without aftertreatments, they cannot meet these regulations. Some injection strategies were introduced in order to decrease emissions without using aftertreatments [1]. Therefore, low-temperature combustion (LTC) was introduced in order to reach a cleaner and more efficient combustion. Some injection strategies, including homogeneous-charged compression ignition (HCCI), premixed charge compression ignition (PCCI), partially premixed combustion (PPC), and reactivity controlled compression ignition (RCCI) were presented in order to achieve LTC. These strategies avoid high local temperature and local rich zones, so NOX and soot emissions can be  reduced significantly. However, these strategies have higher CO and UHC emissions [2, 3, 4, 5, 6, 7, 8]. HCCI strategy delivers homogeneous combustion, but it has the same significant drawbacks, such as combustion controllability, high engine noise level, and high CO and UHC levels [9]. Aoyama et al. [10] accomplished higher control over HCCI combustion by injecting a considerable fuel amount in the intake port, and this strategy was called PCCI. NOX at the air to fuel ratio of 40 in the PCCI case decreased twenty times, when compared to the conventional diesel strategy, but UHC was increased 0.3%-0.4% and one order of magnitude higher than the conventional diesel engine. In 2006, Inagaki et al. [11] presented a new strategy for controlling the PCCI strategy by using two fuels with different reactivity, and today, it is known as RCCI strategy. It was demonstrated that by using this strategy, EPA 2010 is achievable without using aftertreatments. RCCI benefits from the reactivity gradients of two fuels, and it offers better combustion stability, and wide operating ranges compared to PCCI and HCCI. However, UHC and CO emissions remain high in RCCI. In 2013, Lim and Reitz [12] utilized two direct injectors in the combustion chamber to improve RCCI controllability, UHC, and CO emissions. Gasoline fuel was injected with a narrow spray angle in the combustion chamber to prevent the presence of trapped UHC in the crevice volume, and diesel fuel was injected with a wide spray angle to stratify the fuel mixture. Consequently, UHC and CO were improved by about 7% and 27%, respectively. In addition, in 2014, they adopted this method and achieved RCCI operating range to high levels, such as 21 bar gross IMEP [13]. Martin Wissink and Rolf D. Reitz utilized two direct injectors in the combustion chamber and introduced a novel strategy called direct dual fuel stratification strategy (DDFS) in 2015 [14, 15]. In this strategy, gasoline is early injected by the direct injector at about 340 °BTDC to prepare a premixed mixture. Diesel fuel is directly injected at 40-60 °BTDC to stratify the mixture. Near top dead center (TDC), after lowtemperature heat release starts, and before high-temperature heat release, gasoline injector injects considerable fuel

amount to control combustion. This strategy is more controllable compared to RCCI and other LTC strategies. The DDFS strategy was compared with PPC and RCCI at the nominal load of 9.0 bar IMEP. It was found that PPC and RCCI are limited by noise constraints, especially at higher loads, but they are not limiting factors for the DDFS. In addition, the DDFS can achieve a comparable thermal efficiency to the RCCI, and this efficiency is 15% higher than PPC. By adopting DDFS, cyclic variation, noise levels also can be decreased. In contrast to the RCCI, the DDFS has great potential for heat downstream recovery due to its higher exhaust temperature. Thus, some applications such as aftertreatment and turbocharger are more possible with the DDFS strategy. DDFS has a diffusion-limited flame near TDC, which is similar to conventional diesel combustion. Hence, the piston profile affects the formation of emissions significantly. Due to these facts, the modified piston (shown in Figure 1), which was designed and developed for RCCI, may not be effective and suitable for DDFS. Therefore, investigating the piston profile for this strategy is essential. Some previous researches on piston geometry are presented as follows. Poulos and Heywood [16] developed a comprehensive quasi-dimensional model in order to simulate an SI engine and combustion process. The model was validated against experimental data over a wide operating range. Ten different chamber geometries for combustion duration, heat transfer (HT) loss, and thermal efficiency were investigated. The importance of chamber geometry on combustion and the location of the spark plug on burn rate and efficiency were also reported. Zolver et al. [17] performed a numerical investigation on the effect of three different piston bowl shapes on combustion and emissions in a conventional diesel engine. The injection parameters such as mass flow rate were found to be  very important in the control of the combustion. Swirl ratio and turbulent kinetic energy were found to be two critical factors at TDC for combustion characteristics. Wickman et al. [18] performed a numerical investigation on the effect of the chamber geometry on emissions and fuel consumption in heavy-duty and high-speed, light-duty diesel

 FIGURE 1   Piston profiles for stock (open crater), modified,

and scaled.

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engines. Numerical results were validated against experimental data. Moreover, in order to find the best chamber geometry for the mentioned targets, a genetic algorithm with nine inputs was developed. The main objective was to achieve simultaneous optimization for fuel consumption, NOx, and soot emissions. Results showed that NOX was reduced up to one-seventh, and soot was slightly increased by about 25% compared to conventional diesel for the light-duty engine. Christensen et al. [19] performed an experimental investigation on the effect of crevice volumes on UHC emissions in a single-cylinder engine fueled with isooctane under the HCCI strategy. Furthermore, some information about the quenching distance for ultra-lean mixtures was presented. The piston top land was found to be the primary source of UHC and had a significant impact on combustion efficiency. By opening top land up to 2.8 mm, it was found that UHC can be reduced by over 50%. Shi and Reitz [20] presented optimization tools in order to decrease the emissions, swirl ratio levels, and spray targeting using different chamber geometries for a heavy-duty RCCI engine at high load operation. A genetic algorithm (GA) was coupled with the KIVA-CFD code to optimize emissions and fuel consumption. An optimal combination of bowl geometry, spray targeting, and swirl ratio was found in order to minimize the mentioned targets simultaneously. The best results showed a 26% reduction in NOX. Prasad et al. [21] carried out an experimental investigation on the effect of swirl induced due to the two different piston geometries on the emissions and cylinder pressure. Two piston profiles, a hemispherical piston bowl, and a stock one (conventional omega type) were used in a single-cylinder diesel engine. The results showed that the piston profile has a significant impact on the swirl ratio and in-cylinder turbulent kinetic energy (TKE) at the TDC. Seven different piston profiles are investigated numerically to find the best piston profile with respect to the lowest emissions. Moreover, injection timing was optimized, and it was reported that the injection timing of 8.6° BTDC has the best results, and it led to a 27% and 85% reduction in NOX and soot, respectively. Splitter et al. [22] conducted comprehensive experimental and numerical investigations on the effect of compression ratio, piston profile, crevice volume, and squish geometry design on performance and emissions. Experiments were done on two engines, including a light-duty (LD) and a heavy-duty (HD) engine using the RCCI strategy. Experiments demonstrated that CO was decreased by about 42% in the modified profile compared to the stock profile. A reduction in UHC and CO by about 53% and 50% was found in the modified profile compared to the scaled profile. Moreover, it was mentioned that squish volume has dominant effects on UHC emissions compared to crevice volume. It is also demonstrated that with the modified piston profile, the peak pressure rise rate decreased due to the lower compression ratio, but the brake thermal efficiency increased due to the reduction in mechanical and pumping losses. Moreover, a 5%-15% increase in the brake thermal efficiency could be possible in the optimized RCCI compared to the optimized conventional diesel combustion (CDC).

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Figure 1 shows the different conventional piston profiles, including stock (omega type or open crater), modified, and scaled for diesel engines. Dimensions are taken from [22, 23, 24]. Hanson et al. [23] performed experimental and numerical investigations on the effect of the piston profile optimization on emissions and engine performance. Experimental tests were performed on a LD, multicylinder RCCI engine under three operating points. At these points, the stock piston is compared to the original piston equipped by the engine manufacturer. A GA was adopted to optimize the piston profile, and a modified piston was manufactured. By using the modified piston, pumping and HT losses were decreased, and also thermal efficiency was increased by 3%, and a reduction by about 2% in UHC was reported compared to the stock piston design. Dempsey et al. [24] carried out an experimental investigation on the effect of piston geometry profile on emissions and thermal efficiency in a LD RCCI engine over high load ranges. The dual fuel engine was operated with gasoline/diesel and ethanol/diesel fuels. Results proved that the stock profile could be an appropriate choice for conventional diesel engines due to its induced mixing, but for premixed combustion, this character can be  detrimental because of higher HT loss. Furthermore, an increase of 3% in gross indicated thermal efficiency (GTE), and a maximum peak of 51% for GTE was reported. Li et al. [25] conducted a numerical investigation on the effect of three different piston bowl geometries on the combustion process and emissions in a biodiesel engine operated at medium load. The chambers were named as baseline omega combustion chamber (OCC), hemispherical combustion chamber (HCC), and shallow-depth combustion chamber (SCC). It was reported that at low load, the SSC piston had higher IMEP and NOx emissions by about 10% and 17%, respectively, compared to the others. However, the OCC geometry had higher indicated work by about 5% for high load operation. Benajes et al. [26] performed an experimental investigation on the effect of three different piston profiles on emissions and performance at low to high loads. A HD, singlecylinder RCCI engine running at 1,200 rev/min was used. All geometries had the same compression ratio of about 14.4:1 to be theoretically comparable. Single and double injection strategies were presented in all loads for all piston profiles. It was demonstrated that at low load, the stock piston had the highest NOX by about 3.2 gr/kW-hr, and the stepped piston had the lowest NOX by about 1.8 gr/kW-hr. Soot levels were at the same level at low load. At high load, the bathtub piston had the highest NOX and soot by about 3.9 and 0.014 gr/kW-hr, respectively, and the stepped piston had the lowest NOX by about 3.2 gr/kW-hr. Gafoor and Gupta [27] presented a numerical investigation of the effect of different piston bowl geometries on swirl ratio, engine performance, and emissions. The model was validated against experimental data provided from a singlecylinder diesel engine. Some parameters, such as compression

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ratio, engine speed, and total fuel injected mass, were considered constant. An optimization method was performed to find the best piston bowl depth with respect to emissions and performance. In this study, the ratio of bowl diameter to piston diameter (d/D) was varied for thirty-five cases. They found that a low d/D increases NO and decreases soot emissions. Additionally, they reported that the case with initial swirl ratio = 0.5 and d/D = 0.7 has the best results for emissions, and the case with initial swirl ratio = 2.5 and d/D = 0.55 has the best results for both emissions and performance. However, NO is slightly increased in the second case. All DDFS experiments were performed with modified piston, which was optimized and developed for the RCCI strategy [22]. The main objective of this article is to investigate and compare the effects of three conventional piston profiles: modified, stock, and scaled pistons, on the performance and emissions using the DDFS strategy. In the DDFS strategy, a large amount of fuel is directly injected near TDC, which results in a diffusion-limited flame. Thus, the piston profile plays a dominant role in the emission formations. Finally, the benefits of the modified piston and the stock piston were combined, and two new piston profiles are proposed and numerically investigated.

TABLE 1  Engine geometry and common rail injectors’

specifications for the experimental conditions (experimental data are taken from [14]). Engine type

Caterpillar 3401E singlecylinder oil test engine (SCOTE)

Piston type

Modified piston (wide shallow)

Displacement (L)

2.44

Bore (mm)

137.2

Stroke (mm)

165.1

Connecting rod length (mm)

261.6

Squish height (mm)

1.57

Number of valves per cylinder

4

IVO (°ATDC)

335

IVC (°ATDC)

−143

EVO (°ATDC)

130

EVC (°ATDC)

−355

Swirl ratio

0.7

Compression ratio

14.88:1

Common rail injectors name (CRI1) Body style

Bosch CRI2 series

Nozzle angle (°)

148

Hole diameter (𝜇m)

141

Material and Methods

Number of holes

7

A 3D-CFD model was used to simulate combustion and validated against experimental data of a 2.44 L single-cylinder engine at the laboratory at the University of Wisconsin Madison [14, 15]. The engine’s geometrical specifications are given in Table 1. Tests were performed at 1,300 rpm and the nominal IMEP of 9.0 bar. Operating conditions for the experiments are provided in Table 2.

Body style

Bosch CRI2 series

Nozzle angle (°)

143

Hole diameter (𝜇m)

117

Number of holes

10

Computational Model In this article, the 3D-CFD model was used to simulate DDFS combustion at the nominal IMEP of 9.0 bar. This model is prepared for the closed cycle from IVC (143 °BTDC) to EVO (130 °ATDC). In the DDFS strategy, the first injection starts at 340 °BTDC, so for the closed-cycle simulation, it is assumed that gasoline exists homogenously in the combustion chamber. Furthermore, mesh independency was studied for three mesh sizes (2 mm, 1.4 mm, 1 mm), and the appropriate mesh size was considered 1.4 mm for all simulations. In this simulation, the following physical models were applied. The standard discrete droplet model (DDM), a technique for simulating evaporating liquid parcels in the gaseous environment, was used [28]. The Kelvin-Helmholtz (KH) and Rayleigh-Taylor (RT) models, which was described by Beale and Reitz, were used to simulate spray atomization for the primary and secondary break-ups [29]. The no time counter (NTC) model, which is faster than the O’Rourke algorithm, was used to calculate droplet collision in Lagrangian spray simulation [30]. The wall film submodel was used to consider

Common rail injectors name (CRI2)

the interactions between wall and droplets [31]. To simulate turbulent transport of the flow, the renormalization group (RNG) k-𝜀 was applied. This model is appropriate for ICEs [32]. The reduced chemical kinetic mechanism, which was developed for the oxidation of primary reference fuels in ICEs, was applied [33]. In order to predict soot and NOX emissions, Hiroyasu and extended Zeldovich mechanism were used, respectively [34, 35]. Figure 2 depicts the combustion chamber at TDC for the modified piston profile and crevice volume. A  reduced chemical mechanism used for gasoline/diesel combustion was taken from [33].

Model Validation The 3D-CFD model for the DDFS combustion was validated against the experimental data in [14]. The CFD model was validated for the RCCI mode to examine the model’s accuracy, and the numerical results were compared to the experimental data. Mesh independency for the CFD model was studied. The results of the mesh study are presented in Table 3. A mesh with grid size 1.4 mm was chosen for further simulations owing to its efficient computational time and accuracy. Figure 3 shows the comparison of in-cylinder pressure and AHRR for between numerical results and experimental

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RCCI mode

DDFS mode

EGR (%)

49.8

39.5

TEGR (°C)

90.0

90.3

Tin (°C)

51.5

49.3

Pin (kPa)

186.0

186.2

Equivalence ratio

0.66

0.57

Gross IMEP (bar)

9.25

9.41

Qfuel (kJ/cyc)

4.62

4.71

Fuel1

Diesel

Diesel

Injector name

CRI1

CRI

Pressure injection (bar)

500

500

SOI1 (°BTDC)

60

60

Dur1 (ms)

0.6

0.6

Total energy ratio (%)

6.5

7.0

Fuel2

Gasoline

Gasoline

Injector name

PFI

CRI2

Pressure injection (bar)

2.8

1,000

SOI1 (°BTDC)

320

340

Dur1 (ms)

17.1

1.4

SOI2 (°BTDC)

-

4

Dur2 (ms)

-

0.8

data for the DDFS and RCCI modes. RCCI and DDFS experimental tests were performed by the modified piston profile. Table 4 presents numerical results and experimental data for gross IMEP and emissions for both modes. The relative error for the UHC emissions for both strategies is slightly high. UHC has different sources such as incomplete combustion, UHC trapped in crevice volumes, leak past the exhaust valve, deposits on combustion chamber walls, oil on combustion chamber walls, and so on [35, 36], although the CFD model considers UHC sources from incomplete combustion and the crevice volume.

Results and Discussion In the conventional diesel engines, piston profiles affect emissions and thermal efficiency due to controlling the injected fuel direction by the piston profile. In the DDFS strategy experiments performed by Martin Wissink and Rolf Reitz

[14], about 34% of the in-cylinder energy is supplied by the direct gasoline injection near TDC. Hence, the DDFS strategy, unlike the RCCI strategy, piston profile can play a dominant role in the formation of emissions and thermal efficiency. In this section, standard piston geometries for CI engines are presented and used for the DDFS strategy in order to make a comparative study and find the best piston geometry among available piston geometries with regard to engine emissions and performance. Two essential factors, including the squish region and piston surface, are investigated for these piston profiles. Then, the benefits of the geometries are combined, and two new piston profiles are proposed and studied. The compression ratio has significant effects on the indicated thermal efficiency, brake thermal efficiency, and pumping loss in internal combustion engines [22, 37, 38]. Hence, in order to study only the effects of piston geometry profile on emissions and performance, compression ratios of all cases are considered constant to 14.9:1. Crevice volume is considered constant for all piston profiles in the simulation to make sure that the comparative study of the piston profiles is accurate and shows just the effects of piston bowl geometries on emissions and performance. Gross indicated work and gross indicated power per cycle are calculated from Equations 1 and 2, and gross indicated mean effective pressure (gross IMEP) is calculated from Equation 3 [35]:

Workgross =

180

∫ pdV

Eq. (1)

−180



Pgross =

Workgross × nR Vd N



Eq. (2)

where p is pressure and V is volume, P is power, nR for the four-stroke engine equals 2, N is engine speed, and Vd is displacement volume.

Gross IMEP =

Pgross nR Vd N



Eq. (3)

Fuel energy and gross thermal efficiency (GTE) are calculated from Equations 4 and 5, as follows:

E fuel = mgasoline LHVgasoline + mdiesel LHVdiesel

η gross = GTE =

Work gross E fuel

 FIGURE 2   Piston geometry and the crevice volume for the modified piston’s combustion chamber at TDC.

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TABLE 2  Experimental operating conditions for the DDFS and RCCI modes (experimental data are taken from Ref. [14]).

315



Eq. (4) Eq. (5)

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TABLE 3  Mesh independency investigation for the CFD model using DDFS strategy (system configuration is Intel Core i7 6700k

Case

Base grid size (dx,dy,dz) (mm)

NOX (gr/kW-hr)

Soot (gr/kW-hr)

UHC (gr/kW-hr)

CO Computational (gr/kW-hr) time (h) Maximum cell

Coarse mesh

2

0.269

0.104

1.53

4.71

~13

~690,000 ~126,000

Medium mesh

1.4

0.265

0.107

1.57

4.95

~20

~960,000 ~252,000

Fine mesh

1

0.265

0.107

1.58

4.95

~41

~2,580,000 ~462,000

where Efuel is the fuel energy, m is the mass for each species, and LHV is the lower heating value. Exhaust loss (Ex), combustion efficiency (ηcomb), combustion loss (CL), and HT loss are calculated from Equations 6-9 as follows [15]: Ex =





ηcomb =

m exh hexh − m int hint E fuel

Eq. (6)

and intake species (at IVC), respectively. Ex, CL, and HT are the fractions of fuel energy attributed to exhaust, combustion, and HT losses. The adiabatic apparent heat release rate is calculated based on Equation 10 [15]:

∑ m i × LHVi − mCO × LHVCO − mUHC × LHVUHC ∑ m i × LHVi Eq. (7)



CL = 1 − ηcomb

Eq. (8)



HT = 1 − (η gross + Ex + CL )

Eq. (9)

where mi is the mass of each species, mCO is the mass of carbon monoxides, and mUHC is the mass of unburned hydrocarbon at EVO. hexh and hint are the enthalpy of the exhaust (at EVO)

dQad dp 1 γ dV = p + V γ − 1 dθ γ − 1 dθ dθ

Eq. (10)

where θ is the crank angle degree and 𝛾 is the ratio of specific heats.

A Comparative Study for Different Piston Profiles under DDFS Strategy According to Figure 1, three different piston profiles, which are commonly used in the CI engines, are shown. The

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 FIGURE 3   Comparison of the in-cylinder pressure and AHRR between numerical results and experimental data for the modified piston.

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with 32 GB RAM).

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TABLE 4  Numerical results vs. experimental data from [14] for emissions and performance.

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RCCI mode

DDFS mode

Exp. data

Sim. results

Relative error (%)

Exp. data

Sim. results

Gross IMEP (bar)

9.25

9.75

5.41

9.41

10.15

7.86

NOX (gr/kW-h)

0.015

0.013

−13.33

0.291

0.265

−8.93

Soot (gr/kW-h)

0.025

0.026

4.00

0.106

0.107

0.9

UHC (gr/kW-h)

2.25

1.63

−27.55

2.15

1.57

−26.97

CO (gr/kW-h)

2.5

2.36

5.06

4.95

−5.60

modified piston profile is developed and optimized for the RCCI strategy. In addition, this piston profile has higher brake thermal efficiency compared to stock and scaled profiles. Also, UHC and CO emissions were minimized by using this profile in RCCI [22]. Stock and scaled profiles are commonly used in conventional diesel engines. These profiles are developed for combustion with diffusionlimited flames, which is common in conventional diesel engines. Figure 4 specifies the crevice volume, squish volume, and piston crown surface for the stock piston. In this section, the effects of three piston profiles on emissions and the engine’s performance under the DDFS strategy are compared and studied. Two important factors, including combustion efficiency and HT loss, affect GTE. HT loss is proportional to the chamber surface area, and combustion efficiency can be attributed to the squish volume where the most UHC emissions concentrate owing to the flame impenetrability [22]. These geometrical factors are given for piston profiles in Table 5. The modified piston case has the lowest piston area and squish volume. On the other hand, the scaled piston case has the highest piston area and squish volume. All parameters and injection strategies are considered constant, and the 3D-CFD simulations were performed for the three cases. The piston geometry profile, intake port, and valve profiles can directly affect the in-cylinder swirl ratio. In this article, all simulations are considered closed cycle, so only piston profile affects swirl ratio (initial swirl ratios for all cases were considered 0.7 at IVC). Figure 5 shows the swirl ratio diagrams for all cases. The scaled piston case has the highest  FIGURE 4   The crevice volume, squish volume, piston

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surface, and piston top land for the stock piston.

Relative error (%)

−2.18

swirl ratio. On the other hand, modified piston offers the lowest swirl ratio. According to [27, 39], swirl ratio and piston bowl geometry influence combustion efficiency, engine’s emissions, and performance. An increase in the swirl ratio improves CO emissions. However, high swirl ratios lead to higher trapped UHC in the crevice volume. Figure 6 illustrates the in-cylinder pressure and AHRR diagrams for all cases. As it is shown, the highest swirl ratio in the scaled piston case advances the in-cylinder pressure and AHRR. However, the differences between in-cylinder pressure and AHRR for all cases are negligible. The modified piston combustion is retarded due to the lowest swirl ratio compared to other cases. Table 6 gives the data about the peak of pressure, the peak of AHRR, CA50, and gross IMEP for all cases. The scaled piston case has higher in-cylinder pressure by about 1.2 bar, and CA50 is advanced by about 0.4 CAD compared to the modified piston case. Figure 7 shows NOX and soot emission results from the simulation of the three cases. The modified piston case has the highest NOX emissions compared to the others. The stock piston has the lowest NOX and soot emissions. In this case, NOX and soot emissions are reduced by about 9.8% and 8.4%, respectively, compared to modified piston. Figure 8 depicts UHC and CO emissions. The stock case has lower CO emissions by about 23.6% compared to the modified case. However, UHC emissions are increased in the cases with a high swirl ratio owing to the trapped UHC in the crevice and squish regions. The scaled profiles have the highest UHC emissions. In the stock and scaled cases, UHC emissions are increased by about 2.5% and 7%, respectively. Figure 9 shows temperature cut planes for the three piston cases at TDC, 10 °ATDC, 20 °ATDC, and 40 °ATDC. As it is shown, the flame front cannot adequately penetrate into the squish volumes when the combustion chamber geometry has larger squish volume. These regions are appropriate places for UHC. The modified piston case has the lowest squish volume, and the stock and scaled cases have higher squish volumes. Consequently, in the stock piston case, flame penetration is greater (at 20 °ATDC) than the scaled case. However, it is not better than the modified piston case. Figure 10 shows the UHC mass fraction cut planes for all cases at 40 °ATDC. The scaled piston case has the highest UHC emissions, which are trapped in the crevice volume and squish regions. Squish volume does not allow the flame to penetrate,

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Case number

Piston area (mm2)

Chamber area at TDC (mm2)

Clearance volume (mm3) 175,600

(Squish volume/ clearance volume) × 10

Piston area/chamber area

Modified piston

15,441.4

40,982.6

0.14

0.38

Stock piston

19,208.7

44,224.0

1.61

0.43

Scaled piston

19,758.6

44,832.9

2.05

0.44

so the concentration of UHC emissions is high. Consequently, in the engines with an early injection such as the RCCI, HCCI, and PCCI strategies, it is not suitable to have large squish volumes. GTE is related to combustion efficiency and HT losses. Efficiency is related to the complete burning of CO and UHC, and an increase in unburned fuel denotes a decrease in combustion efficiency and GTE. HT loss is proportional to the piston surface area (or generally chamber area). Figure 11 shows the relationships between the piston area/chamber area (PA/CA) and squish volume/ clearance volume (SV/CV) with GTE for all cases. Figure 11 illustrates GTE and two geometrical factors, given in Table 5, for all cases. As previously mentioned, HT and combustion losses can negatively affect GTE. The squish volume/clearance volume (SV/CV) is a geometrical ratio

attributed to the combustion losses; UHC emissions are commonly concentrated in the squish volume with respect to the same crevice volume for all cases. The HT loss is proportional to the ratio of the piston area/chamber area (PA/CA). The scaled piston case has the lowest GTE due to the highest combustion and HT losses, which are associated with squish regions and large piston surface area. On the contrary, modified piston, which is optimized for RCCI, has the highest GTE compared to the other cases. Table 7 shows the comparison of all cases for emissions and GTE. Zero is a neutral feature, plus is a positive feature, and minus is a negative feature for each item. As shown in Table 7, the modified piston case offers the best results for UHC emissions and GTE. The stock piston case has the best results for NOX, soot, and CO emissions.

© SAE International

 FIGURE 5   Comparison of the swirl ratios for the three cases.

© SAE International

TABLE 5  The piston surface specification for the cases.

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 FIGURE 6   Comparison of the in-cylinder pressures and AHRRs for the three cases.

The Effects of Shifting the Gasoline Injector Position to the Center of Head Cylinder Using Stock Piston In this section, the gasoline injector is shifted to the center of the cylinder in the simulation, and the effects of injector position at the center are numerically studied for the stock  piston. Figure 12 shows the shifted injector position schematically. By shifting the gasoline injector position to the center of the cylinder, no considerable differences were found in the in-cylinder pressure, AHRR, CO, and UHC emissions. However, NOX and soot emissions were slightly changed. Figure 13 shows the comparison of NOX and soot emissions for the center-shifted case and the case before shifting. NOX

emissions are reduced by about 13%, and soot emissions are increased by about 6% in the center-shifted case compared to the stock case. It can be concluded that center-injector position decreases NOX and increases soot slightly. NOX and soot emissions depend on local equivalence ratio and local temperature. NOX emissions are formed approximately in the zones with temperatures of more than 2,000  K. In order to compare the high local temperature zones, which is associated with NOX formation, for stock and center-shifted cases, isothermal surfaces with 2,000 and 2,200 K are shown in Figure 14. Regarding these high local temperature regions, NOX is slightly higher in the stock case compared to the center-shifted case due to its larger high local temperature surfaces. To summarize this section, the direct gasoline-injector position impacts NOX and soot emissions, but no considerable effects on UHC, CO, in-cylinder pressure, and AHRR were

TABLE 6  The peak pressure, heat release peak, and gross IMEP for all cases.

Case

Gross IMEP (bar)

Peak Pressure (MPa@CAD)

HR peak (J/θ@CAD)

CA50

Modified piston

10.15

[email protected]

[email protected]

3.9

Stock piston

10.14

[email protected]

[email protected]

3.8

Scaled piston

10.08

[email protected]

[email protected]

3.5 © SAE International

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 FIGURE 7   NOX and soot emissions for all cases.

NOx (gr/ikW-h) 0.30

0.265

Soot (gr/ikW-h) 0.247

0.239

0.25 0.15

0.107

0.10

0.098

0.099

Stock-piston

Scaled-piston

0.05 0.00

Modified-piston

found. In addition, if the stock piston profile locates under the gasoline injector (the diffusion-limited injection), NOX improves significantly.

Proposed Geometry Profiles In this section, two proposed pistons are numerically studied and compared with the stock and modified pistons from the previous sections. Figure 15 shows the benefits of the modified and stock pistons. Figure 16 illustrates the proposed pistons. The proposed piston 1 is a combination of the modified and stock pistons. Modified piston has a low squish volume (benefit), and the stock piston has a higher swirl ratio, and it has a suitable profile for diffusion-limited f lames. Therefore, proposed piston 1 has these benefits simultaneously. However, it has a higher piston area, which

© SAE International

0.20

associates with HT loss compared to the modified piston. The proposed piston 2 was a copy of the modified piston at first, but a dent below the gasoline injector was created to increase its swirl ratio, which associates directly with emission formations. In addition, it has low squish volume (same as the modified piston) and smaller piston area compared to proposed piston 1. Figure 17 shows a 3D view of two proposed pistons for the DDFS strategy. These pistons are asymmetric due to the gasoline injector position. Table 8 presents geometrical features for the proposed pistons. The proposed pistons 1 and 2 are geometrically close to the stock and modified pistons, respectively. All geometries have constant compression ratios to be thermodynamically comparable. According to Figure 18, the swirl ratio of the proposed piston 2 is near to the modified piston, and the swirl ratio of the proposed piston 1 is close to the stock piston case.

 FIGURE 8   UHC and CO emissions for all cases.

UHC (gr/ikW-h) 6.00

CO (gr/ikW-h)

4.95

5.00

3.79

3.78

4.00 3.00

1.57

1.61

1.68

1.00

0.00

Modified-piston

Stock-piston

Scaled-piston

© SAE International

2.00

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© SAE International

 FIGURE 9   Temperature cut planes for all cases at different CADs.

Modified-piston

Stock-piston

Scaled-piston

TDC

TDC

TDC

10 °ATDC

10 °ATDC

10 °ATDC

20 °ATDC

20 °ATDC

20 °ATDC

40 °ATDC

40 °ATDC Temperature (K)

40 °ATDC

1,100 1,200 1,300 1,400 1,500 1,600 1,700 1,800

 FIGURE 10   UHC mass fraction cut planes for all cases at 40 °ATDC.

Modified-piston

Stock-piston

© SAE International

Scaled-piston



UHC mass fracon (–)

321

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 FIGURE 11   The effects of the piston area/chamber area (PA/CA) and squish volume/clearance volume (SV/CV) on gross thermal efficiency (GTE).

Figure 19 depicts the in-cylinder pressure and AHRR diagrams for the cases. The proposed piston 1 case’s in-cylinder pressure and AHRR are slightly advanced, but the proposed piston 2 retards the in-cylinder pressure and AHRR. As presented in Table 9, the proposed piston 2 retards the combustion phasing (CA50) by about 1 CAD, and the gross IMEP is reduced by about 0.6% compared to the modified piston. The proposed piston 1 advances CA50 by about 0.5 CAD, and its gross IMEP is increased by about 0.2% compared to the modified piston. Consequently, piston profiles of the proposed cases have little effects on gross IMEP, but it has considerable effects on emissions. Figure 20 illustrates NOX and soot emissions for the two new proposed pistons and the stock and modified cases. The proposed piston 1 reduces NOX emissions by about 16.6% compared to the modified case. This decrease can be attributed to the piston profile, which is suitable for diffusionlimited flames, and a higher swirl ratio compared to the modified piston case. In addition, this new piston case reduces NOX by about 7.5% compared to the stock case (conventional diesel piston). This case has lower soot emissions by about 1.9% compared to the modified case. The proposed piston 2 case has lower NOX emissions by about 2.6% compared to the modified case. However, it has a higher NOX than the stock

TABLE 7  The comparison of emissions and gross thermal efficiency (GTE) for all cases.

Case

NOX

Soot

UHC

CO

GTE

Modified piston

-

0

+

-

+

Stock piston

+

+

0

+

0

Scaled piston

0

+

-

+

© SAE International

and proposed piston 1 cases. This case has the highest soot emissions compared to the others. Figure 21 shows UHC and CO emissions for the cases. The proposed piston 1 case has the lowest CO emissions compared to the other cases. UHC emissions for all cases are slightly at the same level. In addition, proposed piston 1 reduces UHC by about 1.2% compared to the stock piston. The proposed piston 2 case has the highest CO emissions, which results in poor combustion efficiency and high CL. Consequently, this issue has adverse effects on GTE. NOX emissions are formed in the zones with a temperature higher than 2,000-2,200 K. According to Figure 22, the proposed piston 2 has larger isothermal surfaces (2,200 K) compared to the proposed piston 1, so it has higher NOX emissions. Figure 23 depicts the relation between squish volume/ clearance volume and piston area/chamber area parameters on GTE for the two proposed piston cases and modified and stock profiles. As SV/CV and PA/CA increase, GTE is decreased. However, there is an anomaly in the proposed piston 2 case for the GTE (U-shaped anomaly). This anomaly is attributed to the poor combustion in the proposed piston 2, which resulted from high CO emissions compared to the others. In addition, the in-cylinder pressure of this case is lower than the other cases owing to this poor combustion. To summarize this section, from the emissions standpoint proposed piston 1, which is a combination of the modified and stock pistons, is an acceptable option for the DDFS strategy that uses two direct injectors in the combustion chamber. However, it has about 0.5% lower GTE compared to the modified case. Figure 24 illustrates the fraction of fuel energy for the modified piston, stock piston, scaled piston, and proposed cases. In all cases, exhaust temperatures at EVO are

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 FIGURE 12   The gasoline injector position is shifted to the center of the cylinder, and the diesel-injector position remains constant.

 FIGURE 13   Comparison of NOX and soot emissions for the stock and center-shifted case.

NOx (gr/ikW-h) 0.30 0.25

0.239

0.208

0.20 0.15 © SAE International

0.10

Soot (gr/ikW-h)

0.098

0.104

0.05 0.00

Stock case

Center-shifted case

© SAE International

 FIGURE 14   2,000 and 2,200 K isothermal surfaces for the stock and center-shifted cases at 5 °ATDC.



Stock case (2,000 K)

Center-shi ed case (2,000 K)

Stock case (2,200 K)

Center-shi ed case (2,200 K) 323

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Modified-piston

Low UHCand high GTE (owing to small squish volume and piston surface area)

Stock-piston

Low NOX, soot, and CO (owing to higher swirl rao and the piston profile is suitable for diffusionlimited injecon)

Gain the advantages of both profiles simultaneously by combining two profiles together

© SAE International

 FIGURE 15   The benefits of the modified and stock pistons for the DDFS strategy.

 FIGURE 16   The piston profiles of the two proposed pistons with the same compression ratios with previous cases for the

© SAE International

DDFS strategy.

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 FIGURE 17   The 3D view of the two proposed pistons for the DDFS strategy (pistons are asymmetric).

Proposed-piston 1

Proposed-piston 2

approximately at the same level. Hence, exhaust enthalpies and the difference in the exhaust losses between all cases are negligible. The modified piston has the lowest piston surface area, which leads to the lowest HT loss compared to other cases. The proposed piston 1 has the lowest CL due to the lowest CO and UHC emissions compared to others. Although the proposed piston 2 has a low piston surface area, which associates with low HT loss, its profile is not appropriate for diffusion-limited flames. Consequently, it has the highest CL compared to others.

Summary/Conclusions In this article, a 3D-CFD model was used to simulate DDFS combustion. The model was validated against experimental data for a HD engine at 9 bar and 1,300 rpm, which was obtained by scholars at the University of Wisconsin Madison. These tests were performed with a modified piston profile,

which is developed and optimized for the RCCI strategy. In the DDFS strategy, a considerable amount of fuel is injected directly into the combustion chamber near TDC. In addition, the combustion flame is a diffusion-limited type that is common in conventional diesel engines. So, the piston geometry profile has a significant role in the formation of emissions. In this article, three piston profiles, including modified, stock, and scaled pistons, were numerically investigated and discussed. Finally, two new piston profiles were proposed based on the benefits of the previous pistons and numerically investigated. The main conclusions are as follows: 1. In the first section, three piston profiles, including modified, stock, and scaled pistons, were numerically investigated and compared with the same compression ratios. No considerable changes were found in the in-cylinder pressure and AHRR for all cases. However, the different piston profiles and swirl ratios affected emissions. The modified piston case had the highest GTE owing to smaller squish volume and piston surface. In addition, this case offered the

© SAE International

TABLE 8  The piston geometry specifications for the proposed pistons and previous cases.

Case number

Piston area (mm2)

Chamber area at TDC (mm2)

Clearance volume (mm3) 175,600

(Squish volume/ clearance volume) ×10

Piston area/ chamber area

Modified piston

15,441.4

40,982.6

0.14

0.38

Stock piston

19,208.7

44,224.0

1.61

0.43

Proposed piston 1

18,035.1

42,597.6

0.6

0.42

Proposed piston 2

15,964.3

41,229.4

0.3

0.39

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© SAE International

 FIGURE 18   Comparison of the swirl ratios for two proposed pistons with modified and stock pistons.

© SAE International

 FIGURE 19   Comparison of the in-cylinder pressures and AHRRs for the proposed pistons with the stock and modified cases.

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Case

Gross IMEP (bar)

Peak pressure (MPa@CAD)

HR peak (J/θ@CAD)

CA50

Modified piston

10.15

[email protected]

[email protected]

3.9

Stock piston

10.14

[email protected]

[email protected]

3.8

Proposed piston 1

10.17

[email protected]

[email protected]

3.4

Proposed piston 2

10.09

[email protected]

[email protected]

4.9

lowest UHC emissions. However, NOX, soot, and CO emissions were slightly higher than the other cases. The stock profile had the lowest NOX, soot, and CO emission. 2. In the second section, the gasoline injector position was shifted to the center of the head cylinder and numerically investigated for the stock piston. No considerable differences were found in the in-cylinder pressure, AHRR, CO, and UHC emissions of the center-shifted case compared to the stock case. However, NOX emissions were decreased, and soot was increased slightly. The isothermal surfaces revealed that the center-shifted case has smaller isothermal surfaces (2,200 K), which yield lower NOX emissions. 3. In the third section, two new piston profiles with the same compression ratios were proposed and numerically investigated. The advantage of the modified piston (smaller squish volume, which associated with UHC) and the benefits of the stock piston (high swirl ratio and spray-guided profile, which is suitable for diffusion-limited flames) were adopted and combined in the first-proposed piston profile. The second-proposed piston profile benefits from small squish volume (like modified piston), but

it has a dent to increase the swirl ratio. The incylinder pressures of all cases were compared. The proposed piston 1 advances the in-cylinder pressure and AHRR, and it results in slightly higher gross IMEP. On the contrary, the proposed piston 2 retards the in-cylinder pressure and AHRR. In addition, gross IMEP is decreased compared to other cases. 4. The emissions of the two proposed pistons, stock and modified, were numerically investigated and compared. The proposed piston 1 has the lowest NOX emissions, and soot emissions are decreased compared to the modified case. Furthermore, this new profile yields the lowest CO emissions, and the UHC emissions of all cases are almost at the same level. The proposed piston 2 has a rather high NOX and soot emissions. In addition, this case has the highest CO. 5. Combustion and HT losses affect GTE adversely. Regarding this point, the proposed piston 2 has the highest CL, so it leads to lower GTE compared to others. Although the proposed piston 1 has lower CL compared to the modified piston, it has higher HT loss. Hence, the proposed piston 1 has slightly lower GTE than the modified piston.

 FIGURE 20   Comparison of NOX and soot emissions of the proposed cases with modified and stock cases.

NOx (gr/ikW-h) 0.30

0.265

0.25

0.239

Soot (gr/ikW-h) 0.258 0.221

0.20 0.15 0.10 © SAE International

© SAE International

TABLE 9  The peak pressure, heat release peak, and gross IMEP for all cases.

0.107

0.098

0.105

0.112

0.05 0.00

Modified-piston

Stock-piston

Propsed-piston 1 Propsed-piston 2

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 FIGURE 21   Comparison of UHC and CO emissions of the proposed pistons with modified and stock cases.

UHC (gr/ikW-h)

CO (gr/ikW-h)

7.00

5.83

6.00

4.95 3.78

4.00 3.00 2.00

1.57

1.61

3.61 1.59

1.58

1.00 0.00

Modified-piston

Stock-piston

Propsed-piston 1 Propsed-piston 2

© SAE International

5.00

328

Proposed-piston 1 (2,000 K)

Proposed-piston 2 (2,000 K)

Proposed-piston 1 (2,200 K)

Proposed-piston 2 (2,200 K)

© SAE International

 FIGURE 22   2,000 and 2,200 K isothermal surfaces for the proposed cases at 5 °ATDC.

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© SAE International

 FIGURE 23   The effects of piston area/chamber area (PA/CA) and squish volume/clearance volume (SV/CV) on gross thermal efficiency (GTE) for proposed piston cases.



Fraction of fuel energy (%)

© SAE International

 FIGURE 24   The fraction of fuel energy for all piston profiles with similar compression ratios for all cases.

100% 90% 80% 70% 60% 50% 40% 30% 20% 10% 0%

1 13.9

0.8 15.1

0.9 15.5

0.6 14.8

3 14.3

31.3

31.4

31.6

31.3

31.2

CL HT Ex

53.8

52.7

Modified Stock piston piston

52

53.3

51.5

Scaled piston

Proposed piston1

Proposed piston2

GTE

329

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Contact Information Sasan Shirvani KN Toosi University of Technology [email protected] Saeid Shirvani KN Toosi University of Technology [email protected] Amir H. Shamekhi KN Toosi University of Technology [email protected] Rolf D. Reitz University of Wisconsin Madison [email protected]

Definitions/Abbreviations AHRR - Apparent Heat Release Rate ATDC - After Top Dead Center BDC - Bottom Dead Center BTDC - Before Top Dead Center CA - Chamber Area at TDC CA50 - Crank Angle where 50% of fuel burns CAD - Crank Angle Degree CDC - Conventional Diesel Combustion CFD - Computational Fluid Dynamics CL - Combustion Loss CO - Carbon Monoxide CRI - Common Rail Injector CV - Clearance Volume (Dead Volume) DDFS - Dual Direct Fuel Stratification DDM - Discrete Droplet Model Dur - Duration EGR - Exhaust Gas Recirculation EPA - Environmental Protection Agency EVC - Exhaust Valve Closing EVO - Exhaust Valve Opening Ex - Exhaust Loss GTE - Gross Thermal Efficiency HCCI - Homogenous Charge Compression Ignition HT - Heat Transfer Loss ICE - Internal Combustion Engine IMEP - Indicated Mean Effective Pressure IVC - Intake Valve Closing IVO - Intake Valve Opening KH-RT - Kelvin-Helmholtz-Raleigh-Taylor LHV - Lower Heating Value

LTC - Low-Temperature Combustion NOX - Nitrogen Oxides NTC - No Time Counter PA/CA - Piston Area/Chamber Area PCCI - Premixed Charge Compression Ignition PM - Particulate Matter PPC - Partially Premixed Combustion PPRR - Peak Pressure Rise Rate RCCI - Reactivity Controlled Compression Ignition RNG k-ε - Renormalization Group k-𝜀 SOI - Start of Injection SV/CV - Squish Volume/Clearance Volume (Dead Volume) TDC - Top Dead Center UHC - Unburned Hydro Carbon

References 1. Flynn, P.F., Durrett, R.P., Hunter, G.L., Loye, A.O.Z. et al., “Diesel Combustion: An Integrated View Combining Laser Diagnostics, Chemical Kinetics, and Empirical Validation,” SAE Technical Paper 1999-01-0509, 1999, https://doi. org/10.4271/1999-01-0509. 2. Kokjohn, S.L., Hanson, R.M., Splitter, D.A., and Reitz, R.D., “Experiments and Modeling of Dual-Fuel HCCI and PCCI Combustion Using In-Cylinder Fuel Blending,” SAE Int. J. Engines 2:24-39, 2010, https://doi.org/10.4271/2009-01-2647. 3. Hanson, R.M., Kokjohn, S.L., Splitter, D.A., and Reitz, R.D., “An Experimental Investigation of Fuel Reactivity Controlled PCCI Combustion in a Heavy-Duty Engine,” SAE Int. J. Engines 3:700-716, 2010, https://doi.org/10.4271/201001-0864. 4. Kokjohn, S.L., Reactivity Controlled Compression Ignition (RCCI) Combustion (USA: The University of WisconsinMadison, 2012). 5. Splitter, D., Hanson, R., Kokjohn, S., Reitz, R.D., Reactivity Controlled Compression Ignition (RCCI) Heavy-Duty Engine Operation at Mid-and High-Loads with Conventional and Alternative Fuels, SAE Technical Paper 2011-01-0363, USA, 2011, https://doi.org/10.4271/2011-010363. 6. Li, J., Yang, W., An, H., and Zhao, D., “Effects of Fuel Ratio and Injection Timing on Gasoline/Biodiesel Fueled RCCI Engine: A Modeling Study,” Applied Energy 155:59-67, 2015. 7. Benajes, J., Molina, S., García, A., Belarte, E., and Vanvolsem, M., “An Investigation on RCCI Combustion in a Heavy Duty Diesel Engine Using In-Cylinder Blending of Diesel and Gasoline Fuels,” Applied Thermal Engineering 63:66-76, 2014. 8. Olmeda, P., Garcia, A., Monsalve-Serrano, J., and Sari, R.L., “Experimental Investigation on RCCI Heat Transfer in a Light-Duty Diesel Engine with Different Fuels: Comparison

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