Effects of Injection Parameters and Injection Strategy on Emissions and Performance of a Two-Stroke Opposed-Piston Diesel Engine [2020-01-5064]

Table of contents :
10.4271/2020-01-5064: Abstract
10.4271/2020-01-5064: Keywords
Introduction
Materials and Methods
CFD Simulation
Mesh Study
Model Validation
Results and Discussion
Number of Nozzles
Spray Angle
Injectors’ Angle
Injection Pressure
Start of Injection
Post-Injection and Pre-Injection Strategies
Comparison of the OP2S Piston Engine with 4S Conventional Engines
Summary/Conclusion
Future Work
References

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2020-01-5064 Published 14 Jul 2020

Effects of Injection Parameters and Injection Strategy on Emissions and Performance of a Two-Stroke Opposed-Piston Diesel Engine Saeid Shirvani, Sasan Shirvani, and Amir H. Shamekhi K.N. Toosi University of Technology Citation: Shirvani, S., Shirvani, S., and Shamekhi, A.H., “Effects of Injection Parameters and Injection Strategy on Emissions and Performance of a Two-Stroke Opposed-Piston Diesel Engine,” SAE Technical Paper 2020-01-5064, 2020, doi:10.4271/2020-01-5064.

Abstract

M

odern two-stroke opposed-piston (OP2S) engines offer high efficiency and high power-to-weight ratio compared to conventional four-stroke (4S) engines. Oil consumption and high emissions are usually attributed to 2S engines. However, it is claimed that these drawbacks are well addressed in the modern OP2S engines. In this paper, a three-dimensional computational fluid dynamics (3D-CFD) model for combustion in an OP2S diesel engine was developed and validated against experimental data. The effects of injection parameters, such as the number of nozzles, spray angle, injector’s angle, injection pressure, and the start of injection (SOI), on emissions and performance, were investigated. By rotating one of the injectors by 45 degrees around its axis, it was found that NOX emission decreased by 20% and gross thermal efficiency (GTE) improved by 0.9%. In the case of injection

pressure studies, by increasing the injection pressure up to 1600 bar, soot emission reduced by 7.7%. Advancing the SOI by two crank angle degrees (CAD) can reduce soot emission by up to 16%. To address NOX and soot emissions as primary drawbacks of conventional diesel engines, post- and pre-injection strategies were also studied. In the post-injection study, soot reduced by 30%, and the results of the pre-injection case showed a 37% reduction in NOX. A combination of pre- and post-injection strategies was studied, and this case resulted in 31.2% and 38% reductions in NOX and soot emissions, respectively. Finally, the emissions and performance of the OP2S diesel engine were compared with the experimental results of two conventional 4S diesel engines. The OP2S produces 3.5% and 6.5% higher GTE than heavy-duty (HD) and light-duty (LD) 4S engines. The OP2S also has lower fuel consumption and CO2 by about 15.5% and 10.5%, respectively, compared to the HD4S.

Keywords Opposed-piston engine, Injection parameters, Injection strategy, CFD simulation

Introduction

G

enerally, it is accepted that the first two-stroke (2S) engine was invented by Sir Dugald Clerk in the late nineteenth century in England. The induction process, intake-exhaust port, timing, and piston geometry was patented by Joseph Day in England in 1891. The first engine production was done by Edward Butler in 1887 and J.D. Roots in 1892. At that time, these types of engines were widely used in motorcycles [1]. It is mentioned that two-stroke opposed-piston (OP2S) engines date back to the late 1800s in Europe. The development process of these engines was subsequently done in some other countries [2]. Nowadays, due to the global warming effect, meeting the stringent emission regulations for internal combustion engines has become more challenging. These regulations can be achieved more easily by four-stroke (4S) engines. It is claimed that oil consumption and high emissions, which are considered as two significant drawbacks of OP2S engines, can be reduced in modern ones.

These types of engines can be applied to different applications such as trucks, ships, locomotives, public transportation, SUVs, stationary power generator sets [3, 4, 5], and even unmanned aerial vehicles [6]. Sharma and Redon [7] conducted an on-road transient test cycle (hot-start HD Federal Test Procedure [FTP] cycle) on a multicylinder OP engine. Their results showed that the brake-specific fuel consumption during the hot-start FTP cycle has a deviation of 1.2% from the steady-state torque-tofuel map. Moreover, the emissions of this engine can meet EPA 2010 by adopting conventional aftertreatment systems. Morton et al. [8] carried out an analytical study on a classic OP2S engine. They found a disconnect between the thermodynamic process and converted mechanical work. In the classic OP2S engine, a portion of engine torque was delivered by the leading crank rather than from the tailing crank. This resulted in a fundamental mechanical loss. The analysis provided a fundamental basis for designing OP2S engines.

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EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

Naik et al. [9] presented a paper about performance and emissions results of a 2S 4.9L multicylinder OP diesel engine, which was specifically configured to meet Bharat Stage VI (BS-VI) emission standards in India. The paper incorporates performance and emissions data for test cycles, aftertreatment details, indicated thermal efficiency, friction, and pumping losses of the engine. Results showed great performance is achievable for the engine on both steady-state and transient emission cycles, and conventional aftertreatments could be used to achieve the required emission mandates. Moreover, the engine provides a 10% to 21% fuel economy over conventional 4S engines. Mattarelli et al. [10] performed a numerical investigation focused on the assembly set up of scavenge ports, cylinder, and manifold. Results showed that the optimum value for vane angles was 15°, and the scavenging process was efficient enough at any operating condition. Moreover, the offset angle between the crankshafts should be 10-15° in order to achieve the best fuel efficiency. Salvi et al. [11] presented an overview of an OP2S gasoline compression ignition (GCI) engine. Tests on a prototype demonstrated a considerable improvement in fuel consumption and thermal efficiency and the capability of mass production. Moreover, results showed that engine emissions could comply with the stringent emission regulations. Patil et al. [12] presented the results of a cold-start HD FTP testing of a 4.9L OP2S engine. It was aimed to estimate the necessary warm-up time for the engine to activate the aftertreatment system. The cold-start results showed the capability of this engine to control emissions by providing a rapid heat release at the tailpipe within the first 40 seconds. This advantage will eliminate the need for electrically heat catalysts or diesel fuel mini burners to activate the aftertreatment system. Chown et  al. [13] performed an investigation on the refinements of the cylinder form, hone texture, oil retention, ring design, port sealing, and oil control ring design of an OP2S engine. The aim of the study was to address the mentioned parameters in order to solve the oil consumption, which is one of the most significant drawbacks of 2S engines. Test results showed that the oil consumption level would be competitive to modern 4S engines, and speed versus load maps of the oil consumption diagram for the engine was presented. Durability tests for over 100 hours were done on the engine, and as a result, no performance loss or oil consumption increase was reported. In the study performed by Beatrice et al. [14], the effects of spray parameters and injection pressure were investigated in order to achieve a high power/weight ratio in a 0.5-liter single-cylinder engine. The high injection pressure was a practical way to tackle the trade-off between performance and fuel economy. Utilizing a turbocharger was an important factor in maintaining a high power density; however, a very high fuel injection system and robust mechanical component were fundamental for achieving the high power density. In another study performed by Blasio et al. [15, 16], they found that, by reaching an injection pressure up to 3000 bar, it is possible to decouple the trade-off between emissions and engine performance. They also assessed the system sensitivity to the boundary condition and developed a set of targets for the entire system. They chose three injector nozzles with different

hydraulic flow rates in order to optimize the combustion process. The results in partial load conditions showed significant improvements in noise and engine smoke reduction and fuel economy. In a comparison of OP2S engines with 4S counterparts, it was found that OP2S engines are good alternatives to the conventional 4S engines because of some underlying reasons regarding thermal efficiency, fuel efficiency, and emissions. Some salient reasons are discussed in the following sentences. The OP engines can significantly reduce heat transfer losses due to lower surface area compared to the 4S conventional engines. For instance, a 6-liter OP engine has the area-tovolume ratio as a 15-liter conventional 4S engine. The reduction in heat transfer loss will reduce radiator size and fuel consumption. Faster combustion of OP engines enables a faster heat addition into the combustion chamber and approaches the Otto cycle, which is the ideal cycle for internal combustion engines; thus, thermal efficiency will increase [17]. The OP2S engine is capable of delivering 55% brake thermal efficiency under optimized conditions. From the fuel efficiency perspective, it was found that OP2S has 9% lower brake-specific fuel consumption than the 4S engine. From the studies performed on the OP engine, it was found that the brake thermal efficiency of these engines can reach 15-30% higher than conventional 4S engines [18]. In addition, these engines have great potential in using downstream recovery for selective catalytic reduction. In conventional engines, the catalyst needs 400 to 500 seconds to reach the desirable NOX conversion efficiency in the state of cold-start, but it was reported that this time in OP engines could be reduced to 60-100 seconds [19]. A new generation of GCI OP engine was developed in 2018. This engine demonstrated 30% improvements in the cycle efficiency than conventional 4S engines. NOX and soot were at acceptable levels, and by increasing compression ratio, unburned hydrocarbons (UHC) and carbon monoxide (CO) decreased. By comparing between GCI and diesel combustion, it was found that GCI can operate at the same load as diesel engines, and it has lower NOX and soot emissions. A significant soot reduction by about 3-10 times was achieved by using gasoline, and combustion efficiency improved up to 98% in GCI compared to OP diesel engines [20, 21, 22]. The main objective of this paper was to study the effects of injection parameters, including the number of nozzles, spray angle, injector’s angle, injection pressure, and SOI on the emissions and performance of an OP2S diesel engine. In order to tackle NOX and soot emissions, as the main drawbacks of conventional diesel engines, pre-injection and postinjection strategies were also studied. Finally, the emissions and performance of the OP2S engine were compared with experimental data of conventional 4S diesel engines.

Materials and Methods A numerical CFD model was developed and validated against experimental data for an OP diesel engine operating at 8.8 bar gross indicated mean effective pressure (IMEP) and 1200 rpm. Engine specifications and experimental operating conditions are given in Table 1.

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EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

TABLE 1  Engine specifications and experimental operating

TABLE 2  Boundary conditions for the simulation (data are taken from [4, 17]).

conditions (data are taken from [17, 23, 24]). Bore (mm)

98.4

Stroke (mm)

215.9 (half both pistons)

Connecting rod length (mm)

197.5

Number of cylinders

1

Swept volume (liters)

1.64

Trapped compression ratio

Wall temperature

550 K

Wall motion

Translating

Velocity and temperature condition

Law of the wall

Absolute roughness

0 m

Wall temperature

430 K

17.4

Wall motion

Stationary

Engine speed (rpm)

1200

Law of the wall

Air-fuel ratio

28.4

Velocity and temperature condition

EGR rate (%)

30.4

Absolute roughness

0 m

Intake manifold pressure (bar)

2.1

Intake-port opening

122° BTDC

Exhaust-port opening

110° ATDC

Fuel mass (mg/rev)

62.7

Fuel rail pressure (bar)

1200

Number of injectors (narrow-spray)

2

Number of holes

4

Nozzle diameter (𝜇m)

140

Start of injection (° BTDC)

6

Injection duration (CAD)

10

CFD Simulation The simulation was closed cycle from 122° BTDC to 110° ATDC, and the model generated the desired mesh automatically during each run time. To simulate the exchange of mass, momentum, and energy terms for parcels, the standard Discrete Droplet Model (DDM) was adopted [25]. KelvinHelmholtz and the Rayleigh-Taylor (KH-RT) breakup models were used to capture primary and secondary breakup regimes [26]. To model the spray collisions in the diesel engine, the No Time Counter (NTC) model was applied because it is faster than the standard O’Rourke model [27]. The drop drag model was used to simulate variation in the shape of droplets [28]. For the splash model of the wall films, the O’Rourke model was employed [29]. To simulate the f low turbulence, Renormalization Group k-𝜀 was used. Detailed chemical kinetics using multizone modeling was used to simulate combustion, and it was solved by SAGE solver, which computes reaction rates in chemical reactions [30]. For thermal NOX and soot emissions, the Zeldovich procedure and Hiroyasu model were used, respectively [31, 32]. A reduced chemical mechanism was used for diesel combustion [33]. The boundary conditions for the simulation are given in Table 2.

Cylinder wall

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Fuel system specifications

350

Turbulent kinetic energy at intake-port closing (m2/s2)

60

Swirl ratio at intake-port closing

1.5

Model Validation Figure 2 (A and B) show the injectors’ positions and oval shape of the chamber geometry at TDC. Figure 2(C) illustrates a schematic of the airflow in the cylinder. Figure 3 illustrates the comparison of cylinder pressure and AHRR traces between the numerical results and experimental data for the OP diesel engine.  FIGURE 1   Comparison of cylinder pressure and AHRR

traces for all three mesh sizes.

Mesh Study Mesh independency was studied for three different mesh sizes of 2.8, 2, and 1.6 mm. The model was designed such that it could refine meshes in some critical regions such as spray, cylinder wall, and piston surfaces. In the regions near nozzles, mesh sizes were refined to one-fourth the size of the base mesh size. Moreover, for the cylinder wall and piston surfaces,

The initial temperature in the combustion chamber (K)

meshes were refined to one-half with three sublayers. Figure 1 shows the cylinder pressure and apparent heat release rate (AHRR) for all three mesh sizes. Table 3 presents emissions, computational time, and cell numbers for all mesh sizes. According to Figure 1 and Table 3, the variations of the emissions of the mesh sizes below 2 mm are almost negligible. Regarding computational time and accuracy, a 2 mm mesh size was used for further simulations. All simulations were performed by a computer configured by Intel Core i7-6700K and 32 GB RAM.

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Piston

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EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

Case

Base grid size (dx, dy, dz) NOX (g/kW-h) Soot (g/kW-h) UHC (g/kW-h)

CO (g/kW-h)

Computational time (h)

Min. cells Max. cells

Coarse mesh

2.8 mm

4.112

0.00415

0.175

0.065

~13

7,000

Medium mesh

2 mm

3.893

0.00491

0.198

0.074

~28

Fine mesh

1.6 mm

3.893

0.00491

0.198

0.074

~38

120,000 19,000 300,000 39,000 850,000

Table 4 presents the numerical results and experimental data and relative errors for emissions. According to Table 4, maximum relative errors belong to UHC and CO emissions. UHC and CO emissions are of no significant concern in conventional diesel engines. In addition, UHC emission has different sources, including UHC trapped in crevice volumes, incomplete combustion, deposits on combustion chamber walls, oil on combustion chamber walls, etc. [31, 34]. The UHC sources for this simulation just cover the crevice volume and incomplete combustion sources. Figure 4 shows temperature cut planes for the diffusion flame combustion of the OP diesel engine. Fuel was injected at 6° BTDC, and temperature cut-planes in the oval chamber of the engine are shown in Figure 4 for different CADs. Also, Figure 5 illustrates isothermal surfaces for 1500 K to 2500 K of the combustion simulation. This figure depicts the flame front from the start of combustion to the end of injection for different temperatures. Thermal NOX formation is associated with high local temperature and high equivalence ratio zones.

In diffusion flame combustion, NOX formation usually occurs at the flame front and temperatures over 2000 K [35]. The isothermal surfaces (2100, 2300, and 2500 K) can show these regions.

Results and Discussion The effects of different parameters, including the number of nozzles, spray angle, injectors’ angle, injection pressure, SOI, and post- and pre-injections, are presented and discussed. All formulas for calculating gross work, GTE, and gross IMEP are provided in the Appendix.

Number of Nozzles The effects of the number of nozzles, 2, 4, and 6, for each injector, on emissions and performance, were investigated.

 FIGURE 2   Piston profile and airflow schematic of the OP diesel engine’s cylinder.

(B) 3D chamber view at TDC

(C) Schemac of air-flow in the cylinder

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(A) Piston view

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TABLE 3  Investigation of the mesh independency.

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 FIGURE 3   Cylinder pressure and AHRR traces for numerical

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results and experimental data.

TABLE 4  IMEP, GTE, and emissions for numerical results and

Exp. data Num. results

Relative error (%)

Figure 6 shows the cylinder pressure and AHRR traces for all three cases. The 2-nozzle case had a longer duration of heat release owing to longer injection duration, and as shown in Figure 7, the mean temperature of this case peaked between 0° and 20° ATDC. The NOX emission increased owing to longer injection duration and higher mean temperature, while soot, UHC, and CO emissions decreased and oxidized better. Figure 8 illustrates the differences between flame shapes or isothermal surfaces (2100, 2300, and 2500 K), which contributed to the flame propagation and high thermal NOX regions. The 6-nozzle case had the shortest injection duration and delivered constant fuel mass in a shorter time, so the AHRR peaked and the cylinder pressure advanced compared to the 4-nozzle case (base case). The UHC and CO emissions of the 6-nozzle case were higher than the others due to its shorter time for oxidization. Soot, UHC, and CO emissions of the 6-nozzle case increased by about 3.4 times, 6.6%, and 7.12 times, respectively, compared to the base case. As it is given in Table 5, the GTE for the 6-nozzle case was slightly higher than the other cases which can be attributed to higher gross IMEP and lower convective heat flow shown in Figure 7. Convective heat flow is proportional to the cylinder mean temperature, and the 2-nozzle case had the highest heat f low rate, which affected GTE negatively. Convective heat flow was calculated according to the Woschni correlation given in the Appendix.

Gross IMEP (bar)

8.8

9.2

4.5

Gross indicated power (kW)

28.9

30.2

4.5

Spray Angle

NOX (g/kW-h)

3.772

3.893

3.2

Soot (g/kW-h)

0.005

0.00491

-1.8

UHC (g/kW-h)

0.211

0.198

-6.2

CO (g/kW-h)

0.08

0.074

-7.5

The effects of spray angles on emissions and performance were investigated. The spray angle (𝛼) is the angle between each nozzle and the central axis of the injector, as shown in Figure 9. The study investigated the effects of different values for 𝛼, including 20°, 30°, and 40°. For the amounts above 40°, the

 FIGURE 4   Temperature cut planes at different CADs for the OP diesel engine.

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experimental data (experimental data are taken from [23]).

5

-4° ATDC

-2° ATDC

0° ATDC

2° ATDC

4° ATDC Temperature (K)

6° ATDC

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EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

-4° ATDC (1500 K)

-2° ATDC (1700 K)

0° ATDC (1900 K)

2° ATDC (2100 K)

4° ATDC (2300 K)

6° ATDC (2500 K)

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 FIGURE 5   Isothermal surfaces for the combustion simulation at different CADs.

injected fuel may collide directly with the piston surfaces, so 𝛼 was increased up to 40°. Figure 10 illustrates the cylinder pressure and AHRR traces for different spray angles. The 40° case had the highest cylinder pressure and AHRR, and the 20° case (base case) had the lowest pressure and AHRR compared to the others. Gross IMEP in the 40° case increased by about 3.3% compared to the base case due to higher cylinder pressure, presented in Table 6. NOX increased in the 40° case by a factor of 2 in magnitude due to increased cylinder pressure, when compared to the base case. Soot, UHC, and CO emission also decreased by about 40%, 87%, and 61%, respectively, in this case. As it is given in Table 6, CO for the 40° case was slightly higher than the 30° case; it may account for the fact that the spray collision with the piston surfaces decreased the flame temperature and led to incomplete combustion. Therefore, a high spray angle for the 40° case resulted in higher CO emissions compared to the 30° case.

The 40° case had the highest GTE, and by about 0.9% higher than the base case. It accounts for this fact that output work and IMEP of the 40° case increased, and UHC and CO emissions burned and decreased significantly. Figure 11 shows different isothermal surfaces for each case. The 40° case completely shows that the combustion flames were expanded, and this led to fewer rich zones and more oxygen penetration. So UHC, CO, and soot can be better oxidized in the 40° case. The other isothermal surfaces show the differences in the flame distribution for all cases at different CADs.

 FIGURE 6   Cylinder pressure and AHRR traces for different

 FIGURE 7   Mean temperature and convective heat flow for 2-, 4-, and 6-nozzle cases.

Injectors’ Angle

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nozzle numbers.

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The effects of injectors’ angles on emissions and performance were investigated. As it is shown in Figure 12, in Case 1, the central axis of the injectors in the combustion chamber was rotated 10° around their perpendicular axis, and it may lead to more fluid flow circulation and higher oxygen penetration.

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 FIGURE 8   Isothermal surfaces for the 2-, 4-, and 6-nozzle cases at different CADs. 2-nozzle case

4-nozzle case (base case)

6-nozzle case

2° ATDC (2100 K)

2° ATDC (2100 K)

2° ATDC (2100 K)

4° ATDC (2300 K)

4° ATDC (2300 K)

4° ATDC (2300 K)

6° ATDC (2500 K)

6° ATDC (2500 K)

6° ATDC (2500 K)

 FIGURE 10   Cylinder pressure and AHRR traces for different

TABLE 5  IMEP, GTE, and emissions for different

spray angles.

Case

2-nozzle case

4-nozzle case (base case)

6-nozzle case

Gross IMEP (bar)

9.00

9.2

9.3

GTE (%)

50.8

51.7

52.00

NOX (g/kW-h)

7.138

3.893

3.580

Soot (g/kW-h) 0.00039

0.00491

0.0220

UHC (g/kW-h) 0.00246

0.198

0.211

CO (g/kW-h)

0.074

0.601

0.0023

 FIGURE 9   Schematic of the injector spray angle (𝛼).

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nozzle numbers.

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TABLE 6  IMEP, GTE, and emissions for different spray angles.

Case

20° case (base case)

30° case

40° case

Gross IMEP (bar)

9.2

9.3

9.5

GTE (%)

51.7

52.2

52.6

NOX (g/kW-h)

3.893

6.406

8.162

Soot (g/kW-h)

0.00491

0.00113

0.00292

UHC (g/kW-h)

0.198

0.058

0.026

CO (g/kW-h)

0.074

0.017

0.029

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EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

20-degree case (base case)

30-degree case

40-degree case

2° ATDC (2100 K)

2° ATDC (2100 K)

2° ATDC (2100 K)

4° ATDC (2300 K)

4° ATDC (2300 K)

4° ATDC (2300 K)

6° ATDC (2500 K)

6° ATDC (2500 K)

6° ATDC (2500 K)

In Case 2, one injector was rotated 45° around its axis in order to prohibit direct droplet collisions with another injector, and it may result in higher air and fuel combination than the base case. Figure 13 depicts the cylinder pressure and AHRR traces for different injectors’ angle. In Case 1, peaks of pressure and AHRR increased dramatically. NOX increased by about 26.6% due to the peak of cylinder pressure and AHHR. According to Table 7, the combustion duration of Case 1 decreased by 49.7%. This may account for the collision of the flame with the piston surfaces, and it had negative effects on the soot,

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 FIGURE 11   Isothermal surfaces for 20°, 30°, and 40° cases at different CADs.

UHC, and CO oxidization, owing to incomplete combustion. As it is shown in Figure 14, isothermal surface (2500 K) indicates that the flame vanished faster, unlike other cases, due to the incomplete combustion. A significant reduction in the gross IMEP was found in Case 1, and GTE decreased by about 5.6% compared to the base case. Although Case 1 may lead to a more fluid flow circulation, numerical investigations of the performance and emissions for this case do not recommend Case 1. In Case 2, cylinder pressure and AHRR slightly decreased compared to the base case, and also this type of injector angle

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 FIGURE 12   Two different injector’s angle for the OP diesel engine.

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 FIGURE 13   Cylinder pressure and AHRR traces for different

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injector’s angle.

Case

Case 1

Base case

Case 2

Gross IMEP (bar)

8.1

9.2

9.4

GTE (%)

46.1

51.7

52.6

NOX (g/kW-h)

4.93

3.893

3.116

Soot (g/kW-h)

0.01610

0.00491

0.00487

UHC (g/kW-h)

0.320

0.198

0.185

CO (g/kW-h)

1.138

0.074

0.063

Combustion duration (CAD)

9.5

18.9

22.5

increased combustion duration from 18.9 CAD (base case) to 22.5 CAD. Soot, UHC, and CO emissions reduced by about 1%, 7%, and 15% due to longer combustion duration. A reduction in the peaks of cylinder pressure and AHRR yielded lower NOX emissions by about 20% compared to the base case. Gross IMEP and GTE increased by about 2.2% and 0.9% compared to the base case. From the emissions and performance perspective, a rotation of 45° in the injector’s axial angle (Case 2) will improve emissions and performance. Figure 14 shows the isothermal surfaces for all cases at different CADs. This figure illustrates that in Case 1, a 10° rotation in the injectors’ angles disperses the flame fronts and fluid flow. However, numerical simulation revealed incomplete combustion owing to high UHC and CO emissions. Isothermal surfaces of Case 2 shows that the chance of the collision between two flame fronts of two injectors was reduced, and flames penetrated better in the combustion chamber. This fact may account for the reduction in the UHC, CO, and soot emissions. On the other hand, in the base case, two flame fronts collided at the center of the combustion chamber.

Injection Pressure The effects of injection pressure, 800, 1200 (base case), and 1600 bar, on emissions and performance, were studied. Figure 15 shows the cylinder pressure and AHRR traces for all cases. Reducing the injection pressure to 800 bar resulted in a reduction in the peak of cylinder pressure and AHRR. On the other hand, by increasing the injection pressure, the combustion advanced.

 FIGURE 14   Isothermal surfaces for different injector’s angle cases.

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TABLE 7  IMEP, GTE, and emissions for different injectors’ angles.

9

Case 1

Base Case

Case 2

2° ATDC (2100 K)

2° ATDC (2100 K)

2° ATDC (2100 K)

4° ATDC (2300 K)

4° ATDC (2300 K)

4° ATDC (2300 K)

6° ATDC (2500 K)

6° ATDC (2500 K)

6° ATDC (2500 K)

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EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

 FIGURE 15   Cylinder pressure and AHRR traces for different injection pressures.

 FIGURE 16   Cylinder pressure and AHRR traces for

Table 8 shows the effects of injection pressure on emissions and performance results. In the 800 bar case, gross IMEP and GTE decreased by about 1.1% and 0.5% compared to the base case, respectively. In the 1600 bar case, soot, UHC, and CO emissions decreased by about 7.7%, 11%, and 16%, indicating that injection pressure directly impacted fuel atomization and droplet sizes. As injection pressure increased, soot, UHC, and CO emissions decreased owing to better fuel oxidization.

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different SOIs.

emissions will be improved, but at the expense of a slight increase in NOX.

Post-Injection and PreInjection Strategies It is determined that the post-injection strategy is a promising method to reduce soot in conventional diesel engines. An increase in the mean temperature of the combustion due to post-injection can assist the soot oxidization, and also it is a practical way to activate aftertreatment in these types of engines. However, post-injection usually increases fuel consumption. In the case of post-injection, 10% of the total fuel mass was injected at 8°, 12°, and 16° ATDC. Figure 17 illustrates the cylinder pressure and AHRR traces for the post-injection cases. By retarding the post-injection, the peak of cylinder pressure dropped compared to the base case. Thus, the 16° ATDC had the lowest gross IMEP and GTE, as given in Table 10. NOX increased slightly owing to a sudden increase in the AHRR for the post-injection cases. As it was anticipated, soot decreased dramatically in all cases and reduced up to 30% in the 16° ATDC case. UHC and CO decreased owing to the

Start of Injection The effects of the SOI on performance and emissions of the OP diesel engine were investigated. SOI was swept from -8° ATDC to -4° ATDC by 2 CADs interval. Figure 16 shows the cylinder pressure and AHRR traces for all cases. Advancing SOI by 2 CADs yielded an increase in the peak of AHRR and extended the cylinder pressure; as a result, gross IMEP and GTE increased by about 1% and 0.1%, as given in Table 9. In the -8° ATDC case, NOX increased by 5.7%, while soot, UHC, and CO decreased by 16.7%, 6.6%, and 10.8%, respectively. On the other hand, by retarding SOI, NOX decreased by about 30% in the -4° ATDC case, while soot, UHC and CO emissions increased by 45.2%, 38.8%, and 52.7%. According to the results, by advancing 1 CAD to 2 CAD, GTE and

injection pressures.

Case

SOI = -4° ATDC

base case (SOI = -6° ATDC)

SOI = -8° ATDC 9.3

9.2

1600 bar

GTE (%)

51.4

51.7

51.8

Gross IMEP (bar)

9.1

9.2

9.2

NOX (g/kW-h)

2.732

3.893

4.117

Soot (g/kW-h)

0.00713

0.00491

0.00409

UHC (g/kW-h)

0.275

0.198

0.185

CO (g/kW-h)

0.113

0.074

0.066

Combustion duration (CAD)

26.2

18.9

18.4

GTE (%)

51.2

51.7

51.9

NOX (g/kW-h)

2.750

3.893

3.909

Soot (g/kW-h)

0.00731

0.00491

0.00453

UHC (g/kW-h)

0.212

0.198

0.176

CO (g/kW-h)

0.135

0.074

0.062

© SAE International.

9.1

800 bar

1200 bar (base case)

Gross IMEP (bar)

Case

© SAE International.

TABLE 9  IMEP, GTE, and emissions for different SOI. TABLE 8  IMEP, GTE, and emissions for different

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 FIGURE 17   Cylinder pressure and AHRR traces for different post-injection cases.

11

© SAE International.

© SAE International.

 FIGURE 18   Cylinder pressure, AHRR, and LTHR traces for different pre-injection cases.

sudden increase in the heat release rate and mean temperature of the combustion chamber and better oxidization. NOX and soot are the primary emissions in conventional diesel combustion (CDC); thus, the extra aftertreatment costs are necessary to reduce these emissions. One practical way to reduce NOX and soot simultaneously and reduce the trade-off between these emissions is a more homogenous combustion mode like partially premixed combustion (PPC). An early injection in the OP diesel engine can lead to a significant reduction in these emissions. According to the engine’s load (partial or full load) and engine’s rpm, the amount of preinjected fuel is considered between 5% and 15% of the total fuel. In this study, 10% of the total fuel mass was injected at -60° and -40° ATDC, and the results were discussed. Figure 18 illustrates the cylinder pressure, AHRR, and low-temperature heat release (LTHR) for different cases. Pre-injection caused advanced cylinder pressure and AHRR traces compared to the base case, and the peak of the cylinder pressure decreased, as shown in Figure 18. As it is given in Table 11, gross IMEP and GTE for pre-injection cases decreased slightly, but the pre-injection strategy had significant effects on NOX and soot reduction. According to Table 11, NOX and soot reduced by 36.8% and 3.7% in the -60° ATDC case compared to the base case.

UHC and CO of the pre-injection cases increased. It may account for that the pre-injection cases allow UHC to be trapped in the crevice and squish volumes and incomplete combustion. Early injection led to more homogeneity and higher trapped UHC, so the -60° ATDC case had the highest UHC. The combination case considered pre-injection and postinjection strategies simultaneously. Since the -60° ATDC case (pre-injection) had the most impact on NOX reduction, and the 16° ATDC case (post-injection) had the most effect on soot reduction, the combination case considered 10% of the total fuel energy at -60° ATDC and 10% of total fuel energy at 16° ATDC to target both NOX and soot simultaneously. In the Combination case, NOX increased by about 8.9% compared to the -60° ATDC pre-injection case, while it managed to decrease NOX by 31.2% compared to the base case. As in the -60° ATDC (pre-injection case), soot reduced by 0.7% compared to the base case, and the combination case managed to reduce soot by 38% compared to the base case. As a result, the pre-injection and post-injection strategies are recommended to tackle NOX and soot, as the most primary emissions in conventional diesel engines. The combination case proved that the OP engine could benefit pre-injection and TABLE 11  The effects of pre-injection on IMEP, GTE, and emissions.

Base case (no pre- -60° injection) ATDC

-40° ATDC

Combination of pre- and postinjections

9.1

9.1

Case

Base case 8° ATDC

12° ATDC

16° ATDC

Case

Gross IMEP (bar)

9.2

9.0

9.0

Gross IMEP (bar)

9.2

GTE (%)

51.7

9.1

GTE (%)

51.7

51.5

51.2

51.1

NOX (g/kW-h)

3.893

4.095

4.123

4.159

Soot (g/kW-h)

0.00491

0.00418

0.00382

0.00348

UHC (g/kW-h)

0.198

0.185

0.176

0.165

CO (g/kW-h)

0.074

0.073

0.069

0.066

© SAE International.

© SAE International.

TABLE 10  The effects of post-injection on IMEP, GTE, and emissions.

9.0 51.0

51.2

51.3

NOX (g/kW-h) 3.893

2.46

2.53

2.68

Soot (g/kW-h) 0.00491

0.00473

0.00464 0.00303

UHC (g/kW-h) 0.198

2.12

1.72

1.87

CO (g/kW-h)

4.125

2.451

3.112

0.074

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 12

EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

post-injection strategies, and it is a practical pathway to reduce both NOX and soot simultaneously.

Comparison of the OP2S Piston Engine with 4S Conventional Engines The OP2S engine was compared with the experimental results of two 4S engines: an HD single-cylinder research engine and an LD four-cylinder engine. The geometrical specifications of these engines are available in the Appendix. Since the OP2S had higher GTE than conventional engines, other low temperature combustion (LTC) strategies like reactivity controlled compression ignition (RCCI) and direct dual fuel stratification (DDFS) at almost the same IMEP were compared with the OP2S engine. Table 12 shows the results of emissions, GTE, and fuel consumption for the OP2S and conventional 4S engines under the CDC, RCCI, and DDFS modes. As it is given in Table 12, OP2S had higher GTE by about 3.5% and 6.5% compared to the HD4S and LD4S under CDC at the same load. The OP2S also showed an improvement in the indicated specific fuel consumption (ISFC) by about 15.5% and 19.43% compared to the HD4S-CDC and LD4SCDC, respectively. CO2 is a greenhouse gas and a global concern, from the results of the table; CO2 reduction can be improved by about 10.6% and 14.7% in the OP2S when compared to the HD4S-CDC and LD4S-CDC. From the comparison of experimental results, NOX and soot reduction improved in the OP2S. The OP2S engine is comparable to advanced LTC strategies such as RCCI and DDFS from the GTE and ISFC perspectives. According to the results of the OP2S engine and other conventional 4S engines, it can be concluded that the OP2S is a suitable alternative to the conventional 4S engines.

Summary/Conclusion OP2S engines have a higher power-to-weight ratio owing to their inherent thermodynamic advantages. In addition, they have fewer mechanical parts, and they are smaller than conventional 4S engines. The OP2S engines might be a future alternative to 4S engines owing to their lower costs in mass production and acceptable emissions and performance compared to conventional engines. In this study, a 3D-CFD model was validated against experimental data for an OP2S diesel engine. The effects of injection parameters and strategies on emissions and performance in the OP diesel engine were investigated. The main results are as follows: •• Three nozzle number cases, including 2, 4, and 6, were studied. The 6-nozzle case yielded lower NOX, but soot emission increased. GTE slightly decreased in the 2-nozzle case owing to higher heat transfer loss and longer injection duration. •• Spray angles were studied in three cases, including 20°, 30°, and 40°. In the 30° and 40° cases, NOX increased, but soot decreased. Since the gross IMEP in the 40° case increased, the GTE slightly enhanced. •• The injectors’ angles were varied to study their effects on emissions and performance. Two cases were proposed, including Case 1, which injectors were rotated around their perpendicular axis by 10°, and Case 2, in which one of the injectors was rotated around its axis by 45°. In Case 1, NOX, soot, CO, and UHC emissions increased dramatically. In addition, the GTE and gross IMEP decreased in this case. Thus, this case was not recommended. In Case 2, NOX and soot decreased simultaneously, and a slight increase in the GTE was observed. Consequently, a rotation of 45° like Case 2 could improve emissions and performance in the OP diesel engine. •• In the injection pressure study, higher injection pressure yielded higher NOX and GTE, and soot decreased due to better fuel atomization. In addition, UHC and CO

Case

OP2S

HD4S-DDFS [36, 37, 38, 39, 40]

HD4S-RCCI [36, 38] HD4S-CDC [41]

LD4S-CDC [42]

Gross IMEP (bar)

9.2

9.41

9.25

9.9

10.3

GTE (%)

51.7

50.5

51.1

48.2

45.2

NOX (g/kW-h)

3.893

0.291

0.015

10.0

5.2

Soot (g/kW-h)

0.00491

0.047

0.0011

0.076

0.16

UHC (g/kW-h)

0.198

2.15

2.25

NA

NA

CO (g/kW-h)

0.074

5.06

2.50

NA

NA

CO2 (g/kW-h)

693.75 (30% EGR)

810.45 (40% EGR)

805.92 (50% EGR)

542.95 (No EGR)

569.27 (No EGR)

ISFC (g/kW-h)

149.54

158.4

156.7

177.03

185.61

© SAE International.

TABLE 12  Emissions, GTE, and fuel consumption results of OP2S and 4S engines (NA stands for not available).

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

reduction were improved by increasing the injection pressure. •• The effects of the SOI were studied in three cases, including SOI = -8°, -6°, -4° ATDC. It was found that by advancing the SOI, the GTE and gross IMEP increased, while soot, UHC, and CO emissions decreased. However, a slight increase in NOX was observed. It was recommended to advance the SOI from -6° ATDC to -8° ATDC. •• As soot production is the primary drawback of conventional diesel engines, post-injection is a promising pathway to tackle this problem. The effects of postinjection on emissions and performance were studied. It was found that by 10% of the total fuel injection at 16° ATDC, soot reduced by 30% compared to the base case. The pre-injection strategy was aimed to reduce NOX emission in this study. By injecting 10% of the total fuel at -60° ATDC, NOX emission decreased by 36.8%. •• To benefit pre-injection and post-injection strategies simultaneously, a combination of both strategies was studied. In the combination case, by injecting 10% of the total fuel mass at -60° ATDC and 10% at 16° ATDC, NOX, and soot emissions reduced by about 31.2% and 38%. •• Finally, the emissions and performance results of the OP2S diesel engine were compared with the experimental results of two conventional 4S engines: a heavy-duty engine (HD4S) and a light-duty engine (LD4S). The OP2S produced 3.5% and 6.5% higher GTE than HD4S and LD4S, respectively. The OP2S engine also had lower ISFC and CO2 emission by about 15.5% and 10.5%, respectively, compared to the 4SHD.

Future Work It is intriguing to study LTC strategies like RCCI and DDFS in the OP2S engine as it has two direct injectors in the combustion chamber. One injector can be used specifically for the low-reactivity fuel (gasoline) and the other injector can be used for the high-reactivity fuel (diesel). With these strategies, it may be possible to reach GTEs higher than 52% and meet stringent regulations for NOX and soot emissions without using aftertreatments.

Contact Information [email protected]

Definitions/Abbreviations 2S - Two-stroke 4S - Four-stroke

13

AHRR - Apparent heat release rate ATDC - After top dead center CAD - Crank angle degree CFD - Computational fluid dynamics DDFS - Direct dual fuel stratification DDM - Discrete droplet model FTP - Federal Test Procedure GTE - Gross thermal efficiency HD - Heavy-duty IMEP - Indicated mean effective pressure KH-RT - Kelvin-Helmholtz/Rayleigh-Taylor LTHR - Low-temperature heat release NTC - No time counter OP - Opposed-piston OP2S - Two-stroke opposed-piston OPFC - Opposed-piston folded-cranktrain OPOC - Opposed-piston opposed-cylinder RCCI - Reactivity controlled compressions ignition SMR - Sauter mean radius SOI - Start of injection TDC - Top dead center

References 1. Blair, G.P., Design and Simulation of Two-Stroke Engines (Warrendale, PA: Society of Automotive Engineers, 1996). 2. Pirault, J.-P. and Flint, M., Opposed Piston Engines: Evolution, Use, and Future Applications (Warrendale, PA: SAE International, 2010). 3. Regner, G., Herold, R.E., Wahl, M.H., Dion, E. et al., “The Achates Power Opposed-Piston Two-Stroke Engine: Performance and Emissions Results in a Medium-Duty Application,” SAE International Journal of Engines 4(3):2726-2735, 2011, https://doi.org/10.4271/2011-012221. 4. Redon, F., Kalebjian, C., Kessler, J., Rakovec, N. et al., “Meeting Stringent 2025 Emissions and Fuel Efficiency Regulations with an Opposed-Piston, Light-Duty Diesel Engine,” SAE Technical Paper 2014-01-1187, 2014, https:// doi.org/10.4271/2014-01-1187. 5. Regner, G., Johnson, D., Koszewnik, J., Dion, E. et al., “Modernizing the Opposed Piston, Two Stroke Engine for Clean, Efficient Transportation,” SAE Technical Paper 201326-0114, 2013, https://doi.org/10.4271/2013-26-0114. 6. Zhou, L., Zhang, H., Zhao, Z., and Zhang, F., “Research on Opposed Piston Two-Stroke Engine for Unmanned Aerial Vehicle by Thermodynamic Simulation,” SAE Technical Paper 2017-01-2408, 2017, https://doi. org/10.4271/2017-01-2408.

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7. Sharma, A. and Redon, F., “Multi-Cylinder Opposed-Piston Engine Results on Transient Test Cycle,” SAE Technical Paper 2016-01-1019, 2016, https://doi.org/10.4271/2016-011019. 8. Morton, R., Riviere, R., and Geyer, S., “Understanding Limits to the Mechanical Efficiency of Opposed Piston Engines,” SAE Technical Paper 2017-01-1026, 2017, https:// doi.org/10.4271/2017-01-1026. 9. Naik, S., Johnson, D., Fromm, L., Koszewnik, J. et al., “Achieving Bharat Stage VI Emissions Regulations While Improving Fuel Economy with the Opposed-Piston Engine,” SAE Int. J. Engines 10(1):17-26, 2017, https://doi. org/10.4271/2017-26-0056. 10. Mattarelli, E., Rinaldini, C., Savioli, T., Cantore, G. et al., “Scavenge Ports Ooptimization of a 2-Stroke Opposed Piston Diesel Engine,” SAE Technical Paper 2017-24-0167, 2017, https://doi.org/10.4271/2017-24-0167. 11. Salvi, A., Hanson, R., Zermeno, R., Regner, G., Sellnau, M., and Redon, F., “Initial Results on a New Light-Duty 2.7L Opposed-Piston Gasoline Compression Ignition MultiCylinder Engine,” ASME 2018 Internal Combustion Engine Division Fall Technical Conference, Volume 1: Large Bore Engines; Fuels; Advanced Combustion, 2018. 12. Patil, S., Ghazi, A., Redon, F., Sharp, C. et al., “Cold Start HD FTP Test Results on Multi-Cylinder Opposed-Piston Engine Demonstrating Rapid Exhaust Enthalpy Rise to Achieve Ultra Low NOx,” SAE Technical Paper 2018-01-1378, 2018, https://doi.org/10.4271/2018-01-1378. 13. Chown, D., Koszewnik, J., MacKenzie, R., Pfeifer, D. et al., “Achieving Ultra-Low Oil Consumption in Opposed Piston Two-Stroke Engines,” SAE Technical Paper 2019-01-0068, 2019, https://doi.org/10.4271/2019-01-0068. 14. Beatrice, C., Di Blasio, G., and Belgiorno, G., “Experimental Analysis of Functional Requirements to Exceed the 100 kw/l in High-Speed Light-Duty Diesel Engines,” Fuel 207:591601, 2017. 15. Di Blasio, G., Beatrice, C., Belgiorno, G., Pesce, F. et al., “Functional Requirements to Exceed the 100 kW/l Milestone for High Power Density Automotive Diesel Engines,” SAE Int. J. Engines 10(5):2342-2353, 2017, https://doi. org/10.4271/2017-24-0072. 16. Di Blasio, G., Beatrice, C., Ianniello, R., Pesce, F. et al., “Balancing Hydraulic Flow and Fuel Injection Parameters for Low-Emission and High-Efficiency Automotive Diesel Engines,” SAE Int. J. Adv. & Curr. Prac. in Mobility 2(2):638652, 2020, https://doi.org/10.4271/2019-24-0111. 17. Herold, R., Wahl, M., Regner, G., Lemke, J. et al., “Thermodynamic Benefits of Opposed-Piston Two-Stroke Engines,” SAE Technical Paper 2011-01-2216, 2011, https:// doi.org/10.4271/2011-01-2216. 18. Abani, N., Chiang, M., Thomas, I., Nagar, N. et al., “Developing a 55+ BTE Commercial Heavy-Duty OpposedPiston Engine without a Waste Heat Recovery System,” in Siebenpfeiffer, W., (Ed.), Heavy-Duty-, On- und Off-HighwayMotoren 2016, Proceedings (Wiesbaden: Springer Vieweg, 2017), https://doi.org/10.1007/978-3-658-19012-5_17. 19. Patil, S., Sahasrabudhe, A., Youngren, D., Redon, F. et al., “Cold-Start WHTC and WHSC Testing Results on MultiCylinder Opposed-Piston Engine Demonstrating Low CO2

Emissions while Meeting BS-VI Emissions and Enabling Aftertreatment Downsizing,” SAE Int. J. Adv. & Curr. Prac. in Mobility 1(1):23-37, 2019, https://doi.org/10.4271/2019-260029. 20. Hanson, R., Salvi, A., Redon, F., and Regner, G., “Experimental Comparison of GCI and Diesel Combustion in a Medium-Duty Opposed-Piston Engine,” in ASME 2018 Internal Combustion Engine Division Fall Technical Conference, Volume 1: Large Bore Engines; Fuels; Advanced Combustion, 2018. 21. Sellnau, M., Hoyer, K., Petot, J., Kahraman, E. et al., “Fuel Injection System for Opposed-Piston Gasoline CompressionIgnited (OP-GCI) Engines,” SAE Technical Paper 2019-010287, 2019, https://doi.org/10.4271/2019-01-0287. 22. Hanson, R., Salvi, A., Redon, F., and Regner, G., “Experimental Comparison of Gasoline Compression Ignition and Diesel Combustion in a Medium-Duty Opposed-Piston Engine,” Journal of Energy Resources Technology 141(12), 2019. 23. Naik, S., Johnson, D., Koszewnik, J., Fromm, L. et al., “Practical Applications of Opposed-Piston Engine Technology to Reduce Fuel Consumption and Emissions,” SAE Technical Paper 2013-01-2754, 2013, https://doi. org/10.4271/2013-01-2754. 24. Venugopal, R., Abani, N., and MacKenzie, R., “Effects of Injection Pattern Design on Piston Thermal Management in an Opposed-Piston Two-Stroke Engine,” SAE Technical Paper 2013-01-2423, 2013, https://doi.org/10.4271/2013-012423. 25. Dukowicz, J.K., “A Particle-Fluid Numerical Model for Liquid Sprays,” Journal of Computational Physics 35(2):229253, 1980. 26. Beale, J.C. and Reitz, R.D., “Modeling Spray Atomization with the Kelvin-Helmholtz/Rayleigh-Taylor Hybrid Model,” Atomization and Sprays 9(6), 1999. 27. Schmidt, D.P. and Rutland, C., “A New Droplet Collision Algorithm,” Journal of Computational Physics 164(1):6280, 2000. 28. Liu, A.B., Mather, D., and Reitz, R.D., “Modeling the Effects of Drop Drag and Breakup on Fuel Sprays,” SAE Transactions 83-95, 1993. 29. O'Rourke, P.J. and Amsden, A., “A Spray/Wall Interaction Submodel for the KIVA-3 Wall Film Model,” SAE Transactions 281-298, 2000. 30. Raju, M., Wang, M., Dai, M., Piggott, W. et al., “Acceleration of Detailed Chemical Kinetics Using Multi-zone Modeling for CFD in Internal Combustion Engine Simulations,” SAE Technical Paper 2012-01-0135, 2012, https://doi. org/10.4271/2012-01-0135. 31. Heywood, J., Internal Combustion Engine Fundamentals McGraw-Hill Series in Mechanical Engineering (1988). 32. Hiroyasu, H. and Kadota, T., “Models for Combustion and Formation of Nitric Oxide and Soot in Direct Injection Diesel Engines,” SAE Transactions 85:513–526, 1976. JSTOR, www.jstor.org/stable/44644056. 33. Ra, Y. and Reitz, R.D., “A Reduced Chemical Kinetic Model for IC Engine Combustion Simulations with Primary Reference Fuels,” Combustion and Flame 155(4):713738, 2008.

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EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

34. Pulkrabek, W., “Engineering Fundamentals of the Internal Combustion Engine,” 2nd Ed., Journal of Engineering for Gas Turbines and Power - Transactions of the ASME 126, 2004, https://doi.org/10.1115/1.1669459. 35. Neely, G.D., Sasaki, S., Huang, Y., Leet, J.A., and Stewart, D.W., “New Diesel Emission Control Strategy to Meet US Tier 2 Emissions Regulations,” SAE Transactions 114:512–524, 2005, JSTOR, www.jstor.org/stable/44720976. 36. Wissink, M. and Reitz, R., “The role of the Diffusion-Limited Injection in Direct Dual Fuel Stratification,” International Journal of Engine Research 18(4):351-365, 2017. 37. Wissink, M.L., "Direct Injection for Dual Fuel Stratification (DDFS): Improving the Control of Heat Release in Advanced IC Engine Combustion Strategies," The University of Wisconsin-Madison, 2015. 38. Wissink, M. and Reitz, R., “Direct Dual Fuel Stratification, a Path to Combine the Benefits of RCCI and PPC,” SAE Int. J. Engines 8(2):878-889, 2015, https://doi.org/10.4271/2015-010856.

Appendix

40. Shirvani, S., Shirvani, S., Shamekhi, A.H., and Reitz, R.D., “A Study of Using E10 and E85 under Direct Dual Fuel Stratification (DDFS) Strategy: Exploring the Effects of the Reactivity-Stratification and Diffusion-Limited Injection on Emissions and Performance in an E10/Diesel DDFS Engine,” Fuel 275:117870, 2020. 41. Kokjohn, S.L., Hanson, R.M., Splitter, D., and Reitz, R., “Fuel Reactivity Controlled Compression Ignition (RCCI): A Pathway to Controlled High-Efficiency Clean Combustion,” International Journal of Engine Research 12(3):209-226, 2011. 42. Dempsey, A.B., "Dual-Fuel Reactivity Controlled Compression Ignition (RCCI) with Alternative Fuels," The University of Wisconsin-Madison, 2013.

180

Wgross =

ò pdV

Eq. (A.1)

Pgross =

Wgross N nR

Eq. (A.2)

where p is pressure, V is volume, P is power, N is engine speed, and nR for the 2S engine equals 1. Gross indicated mean effective pressure equals

IMEPgross =

Pgrossn R Vd N

Eq. (A.3)

Input cylinder energy, or fuel energy (Ein), and gross thermal efficiency (GTE) are defined as follows:

Ein = m diesel ´ LHVdiesel GTE =

Gross work Ein

Eq. (A.4) Eq. (A.5)

where mdiesel is the mass of the injected fuels in kg and LHV is the lower heating value in kJ/kg for diesel. The adiabatic apparent heat release rate is calculated based on Equation A.6. [37]:

Eq. (A.6)

where θ is the crank angle degree and 𝛾 is the ratio of specific heats. Convective heat transfer is calculated based on the Woschni correlation as follows [31]:

-180



dQad g dV 1 dp = p + V g - 1 dq g - 1 dq dq



Gross indicated work and gross power per cycle are defined as follows [31]:

39. 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), 2020, https:// doi.org/10.4271/03-13-03-0021.

(

)

2 h c W / m K = 3.26B ( m )







-0.2

´ p ( kPa ) T ( K ) 0.8

-0.55

Vd Tr æmö w ç ÷ = C1 S p + C 2 ( p - pm ) pr Vr è s ø

w (m / s) Eq. (A.7) 0.8

Eq. (A.8)

vs C 2 = 3.24 ´ 10-3 Sp Eq. (A.9) ( for combustion and expansion period )

C1 = 2.28 + 0.308

(

) ( )

Qc ( J / q ) = h c W / m 2K ´ A m 2 ´ ( T - Tw )

Eq. (A.10)

where hc is the heat-transfer coefficient, B is the bore of the engine, p is the in-cylinder pressure, pm is motoring pressure, w is the average cylinder gas velocity, T is the in-cylinder mean temperature, S p is mean piston velocity, and pr, Tr, and Vr are the pressure, temperature, and volume of the working fluid at reference states (IVC). Tw is the wall temperature and Qc is heat flow or convective heat transfer. Table A.1 shows the geometrical engine specifications of two HD and LD engines.

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EFFECTS OF INJECTION PARAMETERS AND INJECTION STRATEGY ON EMISSIONS

TABLE A.1  Engine specifications for HD and LD engines [37, 41, 42].

Engine type

Heavy duty

Light duty

Caterpillar 3401E single-cylinder oil test engine (SCOTE)

GM 1.9 L Diesel (4 cylinder)

Piston type

Stock (CDC) Modified (RCCI and DDFS)

Stock

Displacement (L)

2.44

0.477

Bore (mm) × Stroke (mm)

137.2 × 165.1

82.0 × 90.4

Connecting rod length (mm)

261.6

145.4

Squish height (mm)

1.57

NA

Number of valves per cylinder

4

4

IVC (° ATDC)

-143

-132

EVO (° ATDC)

130

112

Swirl ratio

0.7

Compression ratio

16.1:1 (Stock piston)

Engine speed (rpm)

1200

1.5 to 5 14.88:1 Modified piston

17.4:1 Stock piston 1900

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Engine specification