An Investigation of the Replacement of E10, E85, and Methane with Gasoline in Reactivity Controlled Compression Ignition Combustion: A Comparison of Alternative Fuels Using Reactivity Controlled Compression Ignition Strategy [2020-01-5061]

Table of contents :
10.4271/2020-01-5061: Abstract
10.4271/2020-01-5061: Keywords
1 Introduction
2 Material and Methods
2.1 Computational Model
2.2 Model Validation
3 Results and Discussion
3.1 Effects of Methane, E10, and E85 on Cylinder Pressure, Heat Release, Emissions, and Performance
4 Summary/Conclusion
References

Citation preview

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

2020-01-5061 Published 15 Jun 2020

An Investigation of the Replacement of E10, E85, and Methane with Gasoline in Reactivity Controlled Compression Ignition Combustion: A Comparison of Alternative Fuels Using Reactivity Controlled Compression Ignition Strategy Sasan Shirvani, Saeid Shirvani, and Amir H. Shamekhi KN Toosi University of Technology Citation: Shirvani, S., Shirvani, S., and Shamekhi, A.H., “An Investigation of the Replacement of E10, E85, and Methane with Gasoline in Reactivity Controlled Compression Ignition Combustion: A Comparison of Alternative Fuels Using Reactivity Controlled Compression Ignition Strategy,” SAE Technical Paper 2020-01-5061, 2020, doi:10.4271/2020-01-5061.

Abstract

T

he Reactivity Controlled Compression Ignition (RCCI) strategy is a novel Low-Temperature Combustion (LTC) strategy that is used to minimize nitrogen oxides (NOX) and soot emissions to near zero. Methane and ethanol blends are the most conventional alternatives to gasoline used in the RCCI strategy. In this paper, a three-dimensional Computational Fluid Dynamics (3D-CFD) model was developed and validated against the experimental data at the load of 6.5 bar Indicated Mean Effective Pressure (IMEP) and 1300 rpm. E10, E85, and methane were replaced with gasoline to investigate their effects on emissions and performance. In order to make a fair comparison between cases, combustion phasing (CA50) and cylinder energy and boundary conditions were considered constant. Conventional Diesel Combustion (CDC) was also investigated with the constant cylinder energy to make a comparison between all RCCI cases and CDC. Adding ethanol in gasoline showed that it can increase the gross thermal efficiency (GTE) up to 1.3%, but at the expense

of higher fuel consumption. Ethanol fraction resulted in better fuel oxidization, and lower unburned hydrocarbon (UHC) and carbon monoxide (CO). The methane/diesel RCCI case had lower GTE compared to other RCCI cases, but it still showed higher GTE than CDC. One of the main drawbacks of the methane/diesel RCCI case was the high levels of UHC and CO, but fuel consumption was lower than gasoline/diesel RCCI. From the emissions standpoint, all RCCI cases managed to meet EURO6 and EPA2010 emission mandates for NOX, soot, and CO. However, it is impossible for the CDC strategy to meet these regulations without adopting aftertreatments. The CDC strategy had about 98% of the total produced NOX and soot in all cases. On the other hand, 63% and 51% of the total produced UHC and CO belonged to the methane/diesel case. From the global warming concerns and carbon dioxide (CO2) perspective, the CDC case produced 547.1 g/kW-h CO2 while the methane case produced 350.9 g/kW-h CO2. Other cases (gasoline, E10, and E85) showed middle levels of CO2 by about 460 g/kW-h.

Keywords Alternative fuels, E10/diesel RCCI, E85/diesel RCCI, Methane/ diesel RCCI, Conventional diesel combustion

1. Introduction

T

ypical compression ignition engines usually operate with a high reactivity fuel like diesel, and combustion phasing can be controlled directly by injection timing. Conventional Diesel Combustion (CDC) has high-temperature regions, mixing-controlled regimes, and diffusionlimited nature, which cause high levels of nitrogen oxides (NOX) and particulate matter (PM) emissions. These two emissions are hazardous to human life and can cause serious disease. Regarding these issues, stringent emission regulations mandate automotive industries to implement aftertreatments

in the engines using the CDC strategy. Fuel economy, aftertreatments’ costs, and reaching higher performance incentivized scholars to study advanced combustion strategies like Low-Temperature Combustion (LTC) [1, 2]. In LTC strategies, combustion occurs with the high lean mixture, and by eliminating high-temperature zones, NOX and soot will decrease. Homogeneous Charge Compression Ignition (HCCI), Partially Premixed Compression Ignition (PPCI), and Reactivity Controlled Compression Ignition (RCCI) are some examples of the LTC strategy. One of the most significant drawbacks of LTC strategies is being

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 2

AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

kinetically controlled; thus, they are subject to small changes in the boundary conditions [3, 4]. In the HCCI strategy, one of the firstly developed LTC strategies, a fully premixed charge is inducted in the combustion chamber and combustion is completely kinetically controlled. Combustion in HCCI occurs at a short duration, so the heat addition to the control volume is similar to the Otto cycle, the ideal cycle for engines. This fact will lead to rapid combustion with high noise levels. Many researchers studied HCCI to achieve more authority over this strategy and extend combustion duration [5, 6, 7]. In the PPCI strategy, by injecting a portion of the fuel in the combustion chamber, fuel stratification is created that is capable of extending the combustion duration and ignition delay of the combustion when compared with HCCI. PPCI seeks to increase the operating range of the engine and combustion stability. PPCI strategy needs an optimized fuel selection for each operating condition; combustion phasing and duration cannot be controlled with significant independence. Consequently, this strategy still has a trade-off between noise and engine efficiency [8, 9, 10]. RCCI strategy is a promising approach to extend the combustion duration, which is kinetically controlled, and provide more control via making a reactivity gradient. In RCCI engines, two fuels with different reactivities, such as isooctane (ON=100) and n-heptane (ON=0), are used to make a wide range of different reactivities. Fuel with low reactivity (i.e., isooctane, gasoline, ethanol, or methane) is injected in the intake port, and during the compression stroke of the engine, the fuel with high reactivity (i.e., n-heptane, diesel, biodiesel) is injected with single or double injection strategies. One single strategy is used for low and mid loads of operating conditions, while the double injection strategy is usually used for targeting the squish volume and piston bowl separately. Combustion starts in the zones with high reactivity and propagates into the regions with low reactivity. This concept can extend the combustion duration and engine operating range. RCCI has ultralow levels of NOX and soot, with high thermal efficiency. It was reported that by using RCCI under optimized conditions, the thermal efficiency can reach 60% [11, 12, 13]. Gross et al. [14] performed an experimental investigation on the effects of the mixture of gasoline and diesel fuel “dieseline” as the high-reactivity fuel in a light-duty RCCI engine. The tests were performed over broad loads and speeds to examine the effects of the mixture on emissions and performance. It was found that brake thermal efficiency increased from 3.25% (at 2 bar) up to 18.7% (1 bar) by using dieseline blends. Moreover, combustion efficiency was slightly improved compared to conventional RCCI. CO and UHC slightly reduced. DelVescovo et al. [15] performed an experimental investigation on the effects of fuel stratification, charge preparation, and premixed fuel chemistry on the performance and emissions of a heavy-duty single-cylinder RCCI engine. Fuel stratification was varied with the change in start of injection (SOI) timing and injection pressure. It was found that the peak gross thermal efficiency (GTE) is attainable between −60° ATDC and −45° ATDC regardless of intake conditions. Furthermore, the sensitivity of combustion phasing to intake conditions was decreased by retarding SOI. Secondly, tests were performed by syngas and on methane as the

low-reactivity fuel. It was found that GTE reduced and NOX emissions increased. Direct Dual-Fuel Stratification (DDFS) strategy is a promising pathway that benefits RCCI and PPCI simultaneously. This strategy can control the rate of heat release directly by the near-TDC injection. To have more control over the injection timing and reactivity stratification, two direct injectors are implemented in the combustion chamber. This strategy has comparable thermal efficiency to RCCI, but it has higher levels of soot emissions. Moreover, the DDFS strategy has a higher exhaust temperature due to its near-TDC injection that is suitable for the downstream recovery and aftertreatments [16, 17, 18, 19, 20]. Alcohol fuels are renewable sources of energy, and they are used in many countries such as Brazil, America, and China in the form of blends in gasoline to increase thermal efficiency, reduce emissions, and help better fuel oxidization. In addition, blending alcohol fuels with gasoline or diesel can keep the fuel cost constant during global oscillations in the fuel prices or global energy crisis [21]. By choosing a suitable ratio of ethanol and diesel fuel, it is possible to reduce the trade-off between NOX and soot in dual-fuel and conventional diesel engines. In addition, by adopting ethanol in dual-fuel engines, it is possible to reduce the need for EGR, especially at higher loads [22]. Chuepeng et al. [23] performed an experimental investigation on the effects of ethanol in an RCCI engine at low and mid loads. They reported that by using the ethanol component, fuel consumption increased and NOX and soot decreased. The oxygen molecules in ethanol contribute to better fuel oxidization and a reduction in smoke. The smoke concentration at higher loads was lower than low loads. The RCCI aerosols with the higher ethanol/diesel ratios yielded smaller sizes and greater numbers of particles. Other scholars have studied other alcohol fuels such as methanol, which can be produced from different sources, including natural gas, hydrogen, biomass, coal, and coke-oven gas. Methanol has the capability to decrease NOX and soot emissions, owing to this fact that methanol contains nearly 50% of oxygen content and higher enthalpy of vaporization [24]. Duraisamy et al. [25] conducted an experimental investigation on the effects of hot and cooled EGR on emissions, performance, and cyclic variation of a light-duty turbocharged diesel RCCI engine operated with methanol/diesel fuel. The tests were performed at 3.4 bar and 5.1 bar brake mean effective pressure (BMEP) at 1500 rpm. It was found that cooled EGR can decrease cyclic variation by about 26% and NOX, PM, and thermal efficiency improved with 76% and 81% methanol mass at 3.4 and 5.1 bar, respectively. At hot EGR, cyclic variation, NOX and PM increased compared to the cooled EGR case. Natural gas is a suitable alternative to gasoline, especially when used in dual-fuel RCCI engines. Natural gas is cheaper than gasoline, and it makes a higher reactivity gradient with the high-reactivity fuel (diesel) than gasoline. In addition, it has a higher resistance to auto-ignition than gasoline, which means reaching higher compression ratios is attainable with natural gas than gasoline. Natural gas is also sulfur free, and in the dual-fuel engines, using natural gas/diesel can eradicate soot. Additionally, natural gas makes RCCI engines to reduce

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

their need for EGR, especially at high-load operating conditions. Making combustion duration longer in RCCI engines to extend their operating ranges is more possible with natural gas compared to gasoline. One of the most significant drawbacks of using methane or natural gas is the higher amounts of methanic UHC. High UHC causes combustion loss or poor combustion efficiency; in the methane/diesel dual-fuel engines, combustion loss can vary between 1% and 7% with regard to the engine load. Methane combustion has a high exhaust gas temperature, which is suitable for the aftertreatment activation and downstream recovery purposes. Although UHC in methane/diesel RCCI combustion is higher, under optimized conditions, it is possible to keep UHC and CO at acceptable levels [26, 27, 28]. Syngas is another promising alternative fuel to gasoline in RCCI engines to reduce emissions. Syngas mainly consists of hydrogen (H2), methane (CH4), carbon monoxide (CO), nitrogen (N2), and carbon dioxide (CO2). A lean mixture of syngas in the RCCI engine can lead to reducing NOX and PM. Syngas is attractive because it is produced by urban and agriculture residues and for its combustion properties (high resistance to auto-ignition and flame speed) [29, 30]. Xu et al. [29] performed a numerical investigation on two types of engine fueled by syngas. Engine A is a single-cylinder engine with a shallow-wide piston bowl, and engine B is a single-cylinder engine with the conventional omega-type piston. They optimized some operating parameters, including syngas composition, fuel-supply strategies, and intake conditions using nondominated sort genetic algorithm II. The optimization showed that the shape of the piston bowl determines the injection methods, while the bore of the cylinder controls injection pressure. The optimal EGR rate for both engines was about 25%, and initial pressure and temperature for engine A was relatively high. In this study, a 3D-CFD model was validated against experimental data of a gasoline/diesel RCCI engine operating at the load of 6.5 bar gross IMEP and 1300 rpm. Ethanol blends (E10 and E85) and methane were replaced with gasoline with constant cylinder energy and constant combustion phasing in all cases. A comparative study was performed between the alternative fuels to gasoline using the RCCI strategy. In addition, the CDC was also simulated and compared with RCCI cases. The effects of each alternative fuels on emissions, performance, thermal efficiency, and fuel consumption were studied.

3

homogenous at IVC. The standard Discrete Droplet Model (DDM) was used to exchange mass, momentum, and energy terms of parcels with fluid-phase values [31]. Primary and secondary breakup regimes were simulated by using a combination of the Kelvin-Helmholtz and the Rayleigh-Taylor (KH-RT) breakup models [32]. Droplet collisions always occur in spray modeling, especially in diesel engines. The No Time Counter (NTC) collision approach was used, which is faster than O’Rourke’s model [33]. Droplet drag was employed based on the dynamic drag model to simulate variation in the shape of droplets [34]. To simulate the interaction of spray droplets with the wall surface, three regimes due to the droplets Webber number will occur. These regimes are rebound, slide and stick. It is necessary to simulate the wall film because it affects emissions, especially UHC and thermal efficiency. O’Rourke model was utilized to simulate the splash model for the wall film [35]. In order to simulate flow turbulence, Re-Normalization Group (RNG) k-𝜀 was employed [36]. Detailed chemical kinetics was used to simulate combustion modeling for the SAGE solver. SAGE model was utilized to calculate the concentration of each species during combustion. To simulate the soot formation, a model, according to Hiroyasu [37], was utilized, and for simulating the formation of nitric oxides, the extended Zeldovich mechanism was employed [38]. The piston surface was considered as a translating wall. To set a correct boundary for the walls, the law of the wall was employed for all wall boundary conditions for fully turbulent flow [39]. Physical models and boundary conditions are presented in Table 1 and Table 2, respectively. The reaction mechanism for gasoline/diesel, E10/diesel, and E85/ diesel RCCI was used from [40, 41]. Moreover, the chemical species and reaction mechanism used for the simulation of methane/diesel RCCI combustion was taken from [42]. To examine the reliability of the CFD model for different chemical mechanisms of all fuels, the model was validated against experimental data for other loads, provided in the Appendix.

2.2. Model Validation The experiments were performed on a heavy-duty RCCI engine at the laboratory of the University of Wisconsin-Madison. The engine is a 2.44 L Caterpillar 3401 single-cylinder oil test engine (SCOTE). Engine geometry specifications are given in

TABLE 1  Physical models were utilized in the simulation.

2. Material and Methods

Kelvin-Helmholtz and Rayleigh-Taylor (KH-RT)

In this portion, a numerical 3D-CFD model was validated against the experimental data, mesh independency was studied, and the appropriate mesh size was chosen for further numerical results.

Droplet collision

No Time Counter (NTC)

Spray-wall interaction model

Wall film (O’Rourke model)

Drop drag model

Dynamic drop drag

Splash model

O’Rourke

2.1. Computational Model

Combustion modeling

SAGE chemistry detailed solver

Emissions model

Extended Zeldovich model for NOX emissions

The simulation considers the closed cycle from the intake valve closing (IVC) at −143° ATDC to the exhaust valve opening (EVO) at 130° ATDC. Port injection was considered fully

© SAE International.

Breakup model

Hiroyasu soot model for soot emissions Turbulent modeling RNG k-𝜀

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 4

AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

TABLE 2  Boundary and initial conditions.

Cylinder head

Front face

Back face

470

Wall motion

Translating

IMEP (bar)

6.5

Velocity and temperature condition

Law of the wall

Engine speed (rpm)

1300

EGR (%)

0

Absolute roughness

0 m

Intake temperature (°C)

40

TCylinder wall (K)

425

Intake pressure (bar)

1.5

Wall motion

Stationary

Total fuel (mg/cycle)

75.0

Velocity and temperature condition

Law of wall

Premixed fuel mass of gasoline (%)

86.8

Absolute roughness

0 m

Diesel injection pressure (bar)

600

TCylinder head (K)

425

Diesel SOI#1 (CAD)

−58

Wall motion

Stationary

Velocity and temperature

Law of wall

Absolute roughness

0 m

Boundary type

Periodic

Periodic type

Stationary

Periodic shape

Sector (pie shape) with an angle of 60°

Matching boundary

Back face

Matching boundary

Front face

TIVC (K)

368

PIVC (bar)

1.5

Initial gasoline mass (mg)

65.1

13.72 × 16.51 cm

Connecting rod length

26.16 cm

Piston pin offset

None

Displacement

2.44 L

Geometric comp. ratio

14.88:1

Swirl ratio

0.7

Bowl type

Modified (bathtub)

Number of valves

4

IVC

143° BTDC

EVO

130° ATDC

Fuel injection type for direct injection

Direct injection common rail (6 holes)

© SAE International.

data taken from [43]).

Caterpillar SCOTE

67

Diesel injection duration#1 (CAD)

4.1

Diesel injection duration#2 (CAD)

2.0

3. Results and Discussion In this section, alternative fuels, including ethanol blends with gasoline (E10 and E85) and methane, were replaced with gasoline as the low-reactivity fuel for the RCCI strategy. In addition, all alternative cases were compared with the CDC strategy. In order to make a fair comparison between all cases, cylinder energy and combustion phasing for all cases were considered constant. Table 5 shows the simulation parameters for all cases. Initial boundary conditions were considered constant, and by changing the diesel SOI, combustion phasing was kept constant.

3.1. Effects of Methane, E10, and E85 on Cylinder Pressure, Heat Release, Emissions, and Performance

TABLE 3  HD engine geometry specification (experimental

Bore × stroke

−37

Fraction of diesel fuel in pulse#1 (%)

Figure 3 shows the comparison of the emissions, gross IMEP, and GTE for the numerical results and experimental data.

Table 3, and the tests were performed at the load of 6.5 bar gross IMEP and 1300 rpm [43]. In order to decrease computation time, one-sixth of the chamber geometry was modeled due to the number of the injector nozzles. Figure 1 shows the one-sixth of the combustion chamber with crevice volume and computational grid at TDC. Mesh independency study was performed for the CFD model, and an appropriate mesh size of 1.4 mm was considered for further simulations, available in the Appendix. Mesh independency was studied for three mesh sizes, including 2 mm (coarse mesh), 1.4 mm (medium mesh), and 1 mm (fine mesh). Figure 2 illustrates a comparison between experimental data and numerical simulation results.

Base engine type

Diesel SOI#2 (CAD)

Figure 4 shows a comparison between simulated cylinder pressures for all the alternative fuels and CDC with the gasoline/diesel (G/D) RCCI case. In the E10 and E85 cases, the peak cylinder pressure and AHRR increased, which resulted in higher NOX, presented in Table 6. The methane/ diesel (M/D) RCCI case has the lowest peak cylinder pressure due to its extended AHRR. The E85/diesel (E85/D) RCCI case has the highest peak cylinder pressure as well as AHHR, which accounts for faster ethanol combustion and flame propagation compared to other cases. Table 6 gives the results of emissions, gross IMEP, and peak pressure rise rate (PPRR) for all cases. Figure 5 shows a comparison of the thermal efficiency, which indicated specific fuel consumption (ISFC) and CO2 emissions between all cases.

© SAE International.

Cylinders wall

(experimental data taken from [43]).

TPiston (K)

© SAE International.

Piston

TABLE 4  Experimental engine operating conditions

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

5

© SAE International.

 FIGURE 1   The combustion chamber and computational grid with crevice volume at TDC.

By adding ethanol, NOX emissions increased 1.27 and 3.27 times in the E10 and E85 cases compared to the G/D RCCI case. On the other hand, soot decreased by 15.4% and 69%, respectively, in the E10/D and E85/D cases. Alcohol fuels have oxygen content, and this feature interferes with the chemical production of soot emissions. Moreover, the premixing properties of these fuels reduce soot emissions [44]. UHC and CO decreased by adding ethanol, and IMEP increased due to high oxidation in soot, UHC, and CO. Adding ethanol caused a sudden increase in cylinder pressure and PPRR, which accounts for the higher cylinder pressure. Figure 5 shows an increase in fuel consumption in E10 and E85 cases by about 3% and 40%, respectively. The M/D case has lower cylinder pressure and gross IMEP, but NOX increased, which may account for longer combustion duration of the methane case, indicated in Figure 4 and Table 6. NOX increased by about 3.27 times than the G/D case, while soot is near zero, indicating that methane combustion devastates carbon clusters of soot due to its

© SAE International.

 FIGURE 2   Comparison of numerical results with the experimental data.

longer combustion duration. Among alternative fuels, the M/D case has the highest levels of UHC and CO by about 4.72 and 2.03 orders of magnitude when compared to the G/D case. One explanation for having high amounts of UHC and CO is that methane combustion has slower flame speed and lower flame propagation than gasoline. In addition, methane combustion has a larger quenching distance which impedes better f lame penetration. Thus, UHC and CO concentration will increase, especially for the near-wall zones. Most of the UHC in the M/D RCCI case consists of methane gas, so-called methanic UHC. Fuel consumption in the M/D case is slightly lower than the G/D case that accounts for the high amount of lower heating value (LHV) of methane compared to gasoline. GTE of the M/D case decreased by 4.2% due to lower gross IMEP (lower output work) as well as higher UHC and CO, which indicates poorer combustion efficiency. The CDC case has diffusion-limited combustion, and by single fuel injection at −7.9° ATDC, CA50 was kept constant. In the CDC case, the peak cylinder pressure decreased, which resulted in lower gross IMEP, but the combustion duration of this case increased about three times in magnitude. NOX increased dramatically in the CDC case due to the nature of diffusion-limited combustion and longer injection duration when compared with RCCI combustion. Soot emission also increased in the CDC case. On the other hand, UHC and CO had negligible amounts compared to the RCCI cases. The CDC case has the lowest GTE compared to other cases, and it showed a 5% lower GTE compared to the G/D case. But the CDC case showed a comparable level of GTE to the M/D case but at the expense of 13% higher ISFC. EURO6 and EPA2010 emissions mandates consider 0.4 and 0.27 g/kW-h, respectively, for NOX and 0.013 and 0.01, respectively, for soot emission. Regarding these emission regulations, all of the RCCI cases can satisfy EURO6 and EPA2010, while CDC is not capable of meeting these standards. Figure 5 illustrates the comparison of thermal efficiency, fuel consumption, and CO2 emissions for all cases. The E85 case had the highest GTE but at the expense of higher fuel consumption. The E85 case also produced lower CO2 per kW-h.

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 6

AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

 FIGURE 3   Comparison of numerical results and experimental data.

Experimental data

Numerical results

1E+03 49.1 49.3

1E+02 1E+01

6.5 6.53

7.32 7.19

6.2 6.11

1E+00

1E-02

0.01

1E-03

0.011 0.0014 0.0013

1E-04 IMEP (bar)

NOx (g/kW-h) soot (g/kW-h) UHC (g/kW-h)

However, the variation between ethanol cases and gasoline case in CO2 emissions is negligible. As previously mentioned, the M/D case had the lowest thermal efficiency and fuel consumption, but the CO2 emission of this case is at the lowest level compared to other cases. In the M/D case, CO2 and GTE decreased by 23.1% and 4%, respectively. One of the big concerns of the M/D case is the high levels of methanic HC, which significantly can escalate global warming. The CDC strategy had the lowest thermal efficiency and highest CO2 emissions, but the CDC strategy is capable of providing great control over combustion, unlike the RCCI strategy. Figure 6 visualizes a comparison of emissions for all cases in the percentage index. One hundred percent is the sum of each emission that all cases produced, and the share of each case of the emissions is illustrated. The CDC case produced NOX and soot by about 97% of them, while the majority of produced UHC and CO belonged to the M/D case. The shares of the RCCI cases of all produced NOX and soot are less than 1.5%, indicating how the RCCI strategy produces ultralow NOX and soot emissions compared to CDC. According to Figure 6, one of the most significant drawbacks of the RCCI strategy is the high amounts of UHC and CO; however, these cases can meet EPA2010 for heavy-duty engines for CO emission (less than 21 g/kW-h). That being said, if UHC can

CO (g/kW-h)

GTE (%)

© SAE International.

1E-01

be  oxidized by using aftertreatments, this drawback can be tackled. From that study performed by Prikhodko et al. [45], it was reported that if the exhaust gas temperature exceeds 522 K and a load of 4.2 bar brake thermal efficiency, UHC and CO can be reduced by about 88% and 100% by using diesel oxidation catalyst (DOC). According to Table 6, the exhaust temperature of each case at EVO is about 800 K, which is suitable for the activation of the DOC. Thus, by adopting the DOC, it is possible to meet the EPA2010 standards for UHC for all alternative fuels. Figure 7 shows the energy distribution for all cases. As previously mentioned, GTE is gross thermal efficiency; HTL is the heat transfer loss; EL is the exhaust loss, which is the ratio of the differences between intake and exhaust enthalpies to the cylinder energy; and CL is the combustion loss, which is attributed to UHC and CO emissions. The equations for the calculation of these parameters are available in the Appendix. The M/D RCCI case has the highest level of combustion loss by about 7.6% owing to high UHC and CO emissions, indicating that 7.6% of the cylinder energy cannot be oxidized well or participate in the combustion process. The CDC case has the lowest amount of CL by less than 0.1%; however, in this case, 55.6% of the cylinder energy was lost by heat transfer and exhaust losses through the combustion chamber walls

G/D RCCI

E10/D RCCI

E85/D RCCI

M/D RCCI

CDC

Engine speed (rpm)

1300

1300

1300

1300

1300

Cylinder energy (kJ)

2.97

2.97

2.97

2.97

2.97

Diesel energy fraction of total fuel (%)

16.7

16.7

16.7

16.7

100

Diesel energy fraction in the first injection (%)

67

67

67

67

100

Diesel SOI #1 (CAD)

−58

−58

−58

−7.9

Diesel SOI #2 (CAD)

−37

−58

Diesel inj. press. (bar)

600

600

600

600

600

EGR (%)

0

0

0

0

0

−34.5

−27.1

−22

PIVC (kPa)

145

145

145

145

145

TIVC (K)

370

370

370

370

370

© SAE International.

TABLE 5  Simulation parameters for all cases.

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

 FIGURE 4   Comparison of simulated cylinder pressures for

7

TABLE 6  Emissions and performance parameters of all cases.

© SAE International.

© SAE International.

all alternative fuel cases with the G/D case.

G/D RCCI

E10/D E85/D RCCI RCCI

M/D RCCI

CDC

Gross IMEP (bar)

6.53

6.54

6.27

5.50

0.093

6.39

6.65

NOX (g/kW-h)

0.011

0.025 0.047

Soot (g/kW-h)

0.0013

0.0011 0.0004 0.00002 0.091

UHC (g/kW-h)

6.101

5.877

4.483

28.819

CO (g/kW-h)

7.187

4.594 1.066

14.616

0.780

PPRR (bar/deg)

5.09

5.11

8.14

2.13

5.66

CA50 (°ATDC)

5.63

5.63

5.63

5.63

5.63

Combustion duration (deg)

12.17

11.91

9.12

22.18

36.27

TEVO (K)

795.1

799.3

804.8

832.7

814.1

0.101

started to burn, but the reactivity stratification is still available, and RCCI benefits this feature to extend the combustion duration.

4. Summary/Conclusion

and exhaust ports. As it is shown in Figure 7, by adding 10% ethanol in gasoline, CL can be reduced by 0.6%, and GTE can be increased to 0.4%, indicating the role of -OH radical in ethanol blends on the fuel oxidization and increasing GTE. Figure 8 shows the temperature cut-planes and isothermal surfaces for all cases at similar CA50 (5.63° ATDC). The CDC case has higher local temperature zones, which result in high levels of NOX among all cases. Adding ethanol resulted in higher reactivity stratification than the G/D case. The isothermal surfaces of E10/D and E85/D show a higher stratification compared to the G/D case. The M/D case has the highest reactivity gradient due to the high octane number of methane, which is about 120. The isothermal surfaces of the M/D case show that the higher octane number makes an improvement in reactivity stratification. The CDC isothermal is also at constant CA50 and shows the differences in the combustion between CDC and RCCI. All the regions in RCCI

In this paper, a 3D-CFD model was used to simulate dual-fuel RCCI combustion for gasoline, ethanol, and methane fuels. The model was validated against the experimental data at the load of 6.5 bar gross IMEP and 1300 rpm. The low-reactivity fuel (gasoline) was replaced with E10, E85, and methane in the RCCI combustion to investigate their effects on the combustion characteristics, performance, and emissions. In order to make a fair comparison between all cases, combustion phasing and cylinder energy, as well as boundary conditions for all cases, were kept constant. The main results are as follows: •• The E85/D RCCI case had the highest peak cylinder pressure and PPRR. Thus, the gross IMEP of the E85/D RCCI increased, and it resulted in the highest GTE of 1.3% compared to the G/D RCCI case. The results

 FIGURE 5   Comparison of thermal efficiency, fuel consumption, and CO2 emissions of all cases. GTE (%)

ISFC (g/kW-h)

CO2 (g/kW-h)

600

547.1

500

459

457.4

455.6

400

350.9

300 218.8 © SAE International.

200 100

156 49.3

160.8

153

49.7

50.6

E10/D RCCI

E85/D RCCI

45.1

173.3 44.3

0 G/D RCCI

M/D RCCI

CDC

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 8

AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

 FIGURE 6   Comparison of the shares of each case in emission production.

E10/D RCCI 0.38% G/D RCCI 0.17%

E85/D RCCI 0.72%

E10/D RCCI 1.19%

M/D RCCI 1.42%

E85/D RCCI 0.43%

G/D RCCI 1.40%

M/D RCCI 0.02%

CDC 96.96%

CDC 97.32%

NOx (g/kW-h)

Soot (g/kW-h) CDC 2.83%

CDC 0.22% G/D RCCI 13.44%

G/D RCCI 25.44%

E10/D RCCI 13.00%

M/D RCCI 63.44%

M/D RCCI 51.59%

UHC (g/kW-h)

E10/D RCCI 16.25%

E85/D RCCI 3.89%

CO (g/kW-h)

 FIGURE 7   The comparison of the energy distribution of all cases.

GTE 100%

EL+HTL

0.1

4.2

3.6

1.3

46.5

46.7

48.1

49.3

49.7

50.6

45.1

44.3

G/D RCCI

E10/D RCCI

E85/D RCCI

M/D RCCI

CDC

90% 80% 70%

CL 7.6

47.3

55.6

60% 50% 30% 20% 10% 0%

© SAE International.

40%

© SAE International.

E85/D RCCI 9.91%

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

 FIGURE 8   Temperature cut-planes and isothermal surfaces

for all cases at similar CA50.

9

showed high levels of methanic UHC and CO than other cases. •• The CDC case resulted in the lowest gross IMEP and GTE. Fuel consumption of this case was at the highest level compared to RCCI cases. The CDC case had higher levels of CO2 emissions by about 19.2%, and the M/D case had lower CO2 by about 23.6% compared to the G/D RCCI case.

G/D RCCI

•• All of the RCCI cases can meet EURO6 and EPA2010 emission mandates for NOX, soot, and CO emissions. However, the CDC case cannot meet these regulations for NOX and soot without using aftertreatments.

Definitions/Abbreviations E10/D RCCI

E85/D RCCI

© SAE International.

M/D RCCI

CDC Temperature (K)

indicated that adding ethanol (-OH radical) yielded higher oxidization in soot, UHC, and CO; consequently, combustion loss can be decreased by adding a percentage of ethanol in gasoline. However, fuel consumption in the E10 and E85 cases increased by 3% and 40%, respectively, compared to the G/D case. •• Substituting gasoline with methane resulted in the lower cylinder pressure and gross IMEP, and it affected GTE negatively by 4.2%. But the fuel consumption of the methane case was at the lowest level when compared to other cases, owing to higher LHV of methane. NOX of the methane case increased by a factor of 4.3 compared to the G/D case, and soot was eradicated. The M/D case

AHRR - Apparent Heat Release Rate BMEP - Brake Mean Effective Pressure BTDC - Before Top Dead Center CA50 - The Crank Angle where 50% of fuel burn CAD - Crank Angle Degree CDC - Conventional Diesel Combustion CFD - Computational Fluid Dynamics CL - Combustion Loss CO - Carbon monoxide CO2 - Carbon dioxide DDM - Discrete Droplet Model DOC - Diesel Oxidation Catalysts E10 - Ethanol-gasoline blend with 10% ethanol by volume E10/D RCCI - E10/Diesel RCCI E85 - Ethanol-gasoline blend with 85% ethanol by volume E85/D RCCI - E20/Diesel RCCI EL - Exhaust Loss EVC - Exhaust Valve Closing EVO - Exhaust Valve Opening G/D RCCI - Gasoline/Diesel RCCI GHG - Greenhouse Gas GTE - Gross Thermal Efficiency HCCI - Homogeneous Charge Compression Ignition HTHR - High-Temperature Heat Release HTL - Heat Transfer Loss IARC - International Agency for Research on Cancer IMEP - Indicated Mean Effective Pressure ISFC - Indicated Specific Fuel Consumption IVC - Intake Valve Closing IVO - Intake Valve Opening KH-RT - Kelvin-Helmholtz-Rayleigh-Taylor LHV - Lower Heating Value LTHR - Low-Temperature Heat Release M/D RCCI - Methane/Diesel RCCI

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 10

AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

NOX - Nitrogen Oxides NTC - No Time Counter PCCI - Premixed Charge Compression Ignition PCI - Premixed Compression Ignition PFI - Port Fuel Injector PM - Particulate Matter PPRR - Peak Pressure Rise Rate RCCI - Reactivity Controlled Compression Ignition RI - Ringing Intensity RNG k-𝜀 - Re-Normalization Group k-𝜀 RON - Research Octane Number TDC - Top Dead Center UHC - Unburned Hydrocarbon

References

10. Sellnau, M., Moore, W., Sinnamon, J., Hoyer, K. et al., “GDCI Multi-Cylinder Engine for High Fuel Efficiency and Low Emissions,” SAE Int. J. Engines 8(2):775-790, 2015, https:// doi.org/10.4271/2015-01-0834. 11. 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. 12. Splitter, D., Wissink, M., DelVescovo, D., and Reitz, R.D., “RCCI Engine Operation Towards 60% Thermal Efficiency,” SAE Technical Paper 2013-01-0279, 2013, https://doi. org/10.4271/2013-01-0279. 13. Reitz, R.D. and Duraisamy, G., “Review of High Efficiency and Clean Reactivity Controlled Compression Ignition (RCCI) Combustion in Internal Combustion Engines,” Progress in Energy and Combustion Science 46:12-71, 2015. 14. Gross, C.W. and Reitz, R., “Investigation of Steady-State RCCI Operation in a Light-Duty Multi-Cylinder Engine Using “Dieseline”,” SAE Technical Paper 2017-01-0761, 2017, https://doi.org/10.4271/2017-01-0761.

1. Theis, J.R., Ura, J.A., Li, J.J., Surnilla, G.G. et al., “NOx Release Characteristics of Lean NOx Traps during Rich Purges,” SAE Technical Paper 2003-01-1159, 2003, https:// doi.org/10.4271/2003-01-1159.

15. DelVescovo, D., Kokjohn, S., and Reitz, R., “The Effects of Charge Preparation, Fuel Stratification, and Premixed Fuel Chemistry on Reactivity Controlled Compression Ignition (RCCI) Combustion,” SAE Int. J. Engines 10(4):1491-1505, 2017, https://doi.org/10.4271/2017-01-0773.

2. Johnson, T.V., “Diesel Emissions in Review,” SAE Technical Paper 2011-01-0304, 2011, https://doi.org/10.4271/2011-010304.

16. Wissink, M. and Reitz, R.D., “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/10.4271/2015-01-0856.

3. Sjöberg, M. and Dec, J.E., “EGR and Intake Boost for Managing HCCI Low-Temperature Heat Release over Wide Ranges of Engine Speed,” SAE Technical Paper 2007-010051, 2007, https://doi.org/10.4271/2007-01-0051.

17. Wissink, M. and Reitz, R., “Exploring the Role of Reactivity Gradients in Direct Dual Fuel Stratification,” SAE Int. J. Engines 9(2):1036-1048, 2016, https://doi.org/10.4271/201601-0774.

4. Dec, J.E., Yang, Y., and Dronniou, N., “Boosted HCCIControlling Pressure-Rise Rates for Performance Improvements Using Partial Fuel Stratification with Conventional Gasoline,” SAE Int. J. Engines 4(1):1169-1189, 2011, https://doi.org/10.4271/2011-01-0897.

18. 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.

5. Shibata, G. and Urushihara, T., “Auto-Ignition Characteristics of Hydrocarbons and Development of HCCI Fuel Index,” SAE Technical Paper 2007-01-0220, 2007, https://doi.org/10.4271/2007-01-0220. 6. Shibata, G. and Urushihara, T., “Realization of Dual Phase High Temperature Heat Release Combustion of Base Gasoline Blends from Oil Refineries and a Study of HCCI Combustion Processes,” SAE Technical Paper 2009-01-0298, 2009, https://doi.org/2009-01-0298.

19. Long, W., Li, B., Cao, J., Meng, X. et al., “Effects of DualDirect Injection Parameters on Performance of Fuel Jet Controlled Compression Ignition Mode on a High-Speed Light Duty Engine,” Fuel 235:658-669, 2019. 20. Shirvani, S., Shirvani, S., Shamekhi, A., and Reitz, R., “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(C):117870, 2020.

7. Dec, J.E. and Yang, Y., “Boosted HCCI for high power without engine knock and with ultra-low NOx emissions-using conventional gasoline,” SAE International Journal of Engines 3(1):750-767, 2010, https://doi.org/10.4271/2010-01-1086.

21. Çelebi, Y. and Aydın, H., “An Overview on the Light Alcohol Fuels in Diesel Engines,” Fuel 236:890-911, 2019.

8. Hanson, R., Splitter, D., and Reitz, R.D., “Operating a HeavyDuty Direct-Injection Compression-Ignition Engine with Gasoline for Low Emissions,” SAE Technical Paper 2009-011442, 2009, https://doi.org/10.4271/2009-01-1442.

22. Yu, S., Dev, S., Yang, Z., Leblanc, S. et al., “Early Pilot Injection Strategies for Reactivity Control in Diesel-ethanol Dual Fuel Combustion,” SAE Technical Paper 2018-01-0265, 2018, https://doi.org/10.4271/2018-01-0265.

9. Shen, M., Tuner, M., Johansson, B., and Cannella, W., “Effects of EGR and Intake Pressure on PPC of Conventional Diesel, Gasoline and Ethanol in a Heavy Duty Diesel Engine,” SAE Technical Paper 2013-01-2702, 2013, https:// doi.org/10.4271/2013-01-2702.

23. Chuepeng, S., Theinnoi, K., and Tongroon, M., “Combustion Characteristics and Particulate Matter Number Size Study of Ethanol and Diesel Reactivity Controlled Compression Ignition Engine,” SAE Technical Paper 2017-24-0143, 2017, https://doi.org/10.4271/2017-24-0143.

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

24. Jia, Z. and Denbratt, I., “Experimental Investigation into the Combustion Characteristics of a Methanol-Diesel Heavy Duty Engine Operated in RCCI Mode,” Fuel 226:745753, 2018. 25. Duraisamy, G., Rangasamy, M., and Nagarajan, G., “Effect of EGR and Premixed Mass Percentage on Cycle to Cycle Variation of Methanol/Diesel Dual Fuel RCCI Combustion,” SAE Technical Paper 2019-26-0090, 2019, https://doi. org/10.4271/2019-26-0090. 26. Nieman, D.E., Dempsey, A.B., and Reitz, R.D., “Heavy-Duty RCCI Operation Using Natural Gas and Diesel,” SAE Int. J. Engines 5(2):270-285, 2012, https://doi.org/10.4271/2012-01-0379. 27. Mattson, J.M., Langness, C., and Depcik, C., “An Analysis of Dual-Fuel Combustion of Diesel with Compressed Natural Gas in a Single-Cylinder Engine,” SAE Technical Paper 2018-01-0248, 2018, https://doi.org/10.4271/2018-01-0248. 28. Nieman, D.E., Morris, A.P., Miwa, J.T., and Denton, B.D., “Methods of Improving Combustion Efficiency in a HighEfficiency, Lean Burn Dual-Fuel Heavy-Duty Engine,” SAE Technical Paper 2019-01-0032, 2019, https://doi. org/10.4271/2019-01-0032. 29. Xu, Z., Jia, M., Xu, G., and Chang, Y., “Computational Optimization of Syngas/Diesel RCCI Combustion at Low Load in Different Engine Size,” SAE Technical Paper 201901-0573, 2019, https://doi.org/10.4271/2019-01-0573. 30. Pessina, V., D'Adamo, A., Iacovano, C., Fontanesi, S. et al., “Numerical Simulation of Syngas Blends Combustion in a Research Single-Cylinder Engine,” SAE Technical Paper 2019-24-0094, 2019, https://doi.org/10.4271/2019-24-0094. 31. Dukowicz, J.K., “A Particle-Fluid Numerical Model for Liquid Sprays,” Journal of Computational Physics 35(2):229253, 1980. 32. Beale, J.C. and Reitz, R.D., “Modeling Spray Atomization with the Kelvin-Helmholtz/Rayleigh-Taylor Hybrid Model,” Atomization and Sprays 9(6):623-650, 1999. 33. Schmidt, D.P. and Rutland, C., “A New Droplet Collision Algorithm,” Journal of Computational Physics 164(1):6280, 2000. 34. Liu, A.B., Mather, D., and Reitz, R.D., “Modeling the Effects of Drop Drag and Breakup on Fuel Sprays,” SAE Technical Paper 930072, 1993, https://doi.org/10.4271/930072. 35. O'Rourke, P.J. and Amsden, A., “A Spray/Wall Interaction Submodel for the KIVA-3 Wall Film Model,” SAE Technical Paper 2000-01-0271, 2000, https://doi.org/10.4271/2000-01-0271.

11

36. Han, Z. and Reitz, R.D., “Turbulence Modeling of Internal Combustion Engines Using RNG κ-ε Models,” Combustion Science and Technology 106(4-6):267-295, 1995. 37. Hiroyasu, H. and Kadota, T., “Models for Combustion and Formation of Nitric Oxide and Soot in Direct Injection Diesel Engines,” SAE Technical Paper 760129, 1976, https:// doi.org/10.4271/760129. 38. Heywood, J., Internal Combustion Engine Fundamentals, McGraw-Hill Series In Mechanical Engineering (1988). 39. Richards, K.S.P. and Pomraning, E., CONVERGE (v2. 2.0) (Madison, WI: Convergent Science. Inc, 2014). 40. 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. 41. Marinov, N.M., “A Detailed Chemical Kinetic Model for High Temperature Ethanol Oxidation,” International Journal of Chemical Kinetics 31(3):183-220, 1999. 42. Converge Science Inc, CONVERGE 2.2.0 Theory Manual. 2013. 43. Wang, H., DelVescovo, D., Yao, M., and Reitz, R.D., “Numerical Study of RCCI and HCCI Combustion Processes Using Gasoline, Diesel, Iso-Butanol and DTBP Cetane Improver,” SAE Int. J. Engines 8(2):831-845, 2015, https://doi. org/10.4271/2015-01-0850. 4 4. Jo, S., Park, S., Kim, H.J., and Lee, J.-T., “Combustion Improvement and Emission Reduction through Control of Ethanol Ratio and Intake Air Temperature in Reactivity Controlled Compression Ignition Combustion Engine,” Applied Energy 250:1418-1431, 2019. 45. Prikhodko, V.Y., Curran, S.J., Barone, T.L., Lewis, S.A. et al., “Emission Characteristics of a Diesel Engine Operating with In-Cylinder Gasoline and Diesel Fuel Blending,” SAE Int. J. Fuels Lubr. 3(2):946-955, 2010, https://doi.org/10.4271/201001-2266. 46. Splitter, D., Hanson, R., Kokjohn, S., and Reitz, R.D., “Reactivity Controlled Compression Ignition (RCCI) HeavyDuty Engine Operation at Mid-and High-Loads with Conventional and Alternative Fuels,” SAE Technical Paper 2011-01-0363, 2011, https://doi.org/10.4271/2011-01-0363. 47. 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).

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

AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

Appendix

Calculating Operating Parameters

Mesh Study

Some operating parameters for emissions and engine performance were calculated as follows. Gross indicated work and gross power per cycle are defined as follows [38]:

Figure A.1 shows the comparison of pressure and AHRR traces between different mesh sizes, including 2, 1.4, and 1 mm. As it is shown, the differences between 1.4 (medium mesh) and 1 (fine mesh) mm cases are negligible. Table A.1 presents the results of emissions and operating parameters for different mesh sizes. Regarding computation time and accuracy, the appropriate mesh size was considered 1.4 mm for all simulations in this study. Numerical simulation was performed by a system with 16.0 GB RAM, and the processor is an Intel® Core™ i7-6700K CPU at 4.00 GHz.

180

Wc ,i =



ò pdV

Eq. (A.1)

-180

Pi =



WC ,i N nR

Eq. (A.2)

where p is pressure, V is volume, N is engine speed, and nR for the four-stroke engine equals 2. Gross indicated mean effective pressure equals

Reaction Mechanism Validation for Different Fuels

IMEP =



Pn i R Vd N

Eq. (A.3)

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

The accuracy of the model was examined for isooctane, n-heptane, and ethanol combustion under different loads. Figure A.2 illustrates the results of cylinder pressure and AHRR traces as well as emissions for gasoline/diesel and E85/ diesel combustion at various loads. Table A.2 shows the properties of the fuels used in the numerical simulations.

Ein = x LR fuel .LHVLR fuel + x HR fuel .LHVHR fuel Eq. (A.4)



GTE =



Gross work Ein

Eq. (A.5)

where xLRfuel is the mass of low-reactivity fuels in kg and xHRfuel is the mass of high-reactivity fuels such as diesel. LHV is the lower heating value in kJ/kg for each species. Combustion efficiency, combustion, exhaust, and heat transfer losses are defined as follows [47]:

 FIGURE A.1   Cylinder pressure and AHRR traces for

different mesh sizes.



hcomb =

å m i ´ LHVi - m CO ´ LHVCO - m UHC ´ LHVUHC å m i ´ LHVi Eq. (A.6)



© SAE International.



EL =

CL = 1 - hcomb

Eq. (A.7)

m exh h exh - m int h int E fuel

Eq. (A.8)

HTL = 1 - ( GTE + EL + CL )

Eq. (A.9)

where mi is the mass of species, mCO is the mass of carbon monoxides at EVO, and mUHC is the mass of UHC at EVO. hexh and hint are the enthalpies for exhaust (at EVO) and intake

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

13

TABLE A.1  Investigation on emissions for different mesh sizes.

Case

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

Soot (g/kW-h) UHC (g/kW-h)

CO (g/kW-h)

Computation Maximum and time minimum cells

Coarse mesh

2

0.015

0.0011

5.731

6.382

3.5 h

Medium mesh

1.4

0.011

0.0013

6.101

7.187

7h

Fine mesh

1

0.011

0.0013

6.101

7.187

14 h

~115,000

© SAE International.

~20,000 ~219,000 ~49,000 ~433,000 ~81,000

where θ is the crank angle degree and 𝛾 is the ratio of specific heats. For assessing noise level, peak pressure rise rate (PPRR) is defined as follows:

(at IVC) species, respectively. EL is the fraction of fuel energy attributed to the exhaust losses. Combustion loss (CL) is attributed to unburned species (CO and UHC) at EVO and is equal to the ratio of their LHV to Efuel, and HTL is the fraction of fuel energy attributed to the heat transfer loss. The adiabatic apparent heat release rate is calculated based on Equation A.10 [47]:

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

æ dp ö PPRR = ç ÷ è dq ømax



Eq. (A.10)

Eq. (A.11)

 FIGURE A.2   Comparison of cylinder pressure, AHRR, and emissions between the numerical results and experimental data for gasoline/diesel and E85/diesel combustion (experimental data taken from [16, 46]).

Exp. NOx

Num. NOx

Exp. soot×10

Num. soot×10

Exp. UHC+CO

Num. UHC+CO 8 6 4 2 0

0.4 0.3 0.2 0.1 0 DDFS

© SAE International.

G/D DDFS 9.41 bar (40% EGR) and G/D RCCI 9.25 bar gross IMEP (50% EGR).

RCCI

Emissions in gr/kW-hr for experimental data and numerical results (Exp. data taken from [16]). Exp. NOx

Num. NOx

Exp. soot×10

Num. soot ×10

Exp. UHC+CO

Num. UHC+CO

0.15

15

0.1

10

0.05

5

0

0 9.6 bar

E85/D RCCI 9.6 bar and 11.6 bar gross IMEP and no EGR.

11.6 bar

Emissions in gr/kW-hr for experimental data and numerical results (Exp. data taken from [46]).

Downloaded from SAE International by Saeid Shirvani, Thursday, September 02, 2021 14

AN INVESTIGATION OF THE REPLACEMENT OF E10, E85, AND METHANE

TABLE A.2  Fuel properties used in the simulations.

Chemical formula

Isooctane

n-Heptane

Ethanol

Methane

C8H18

C7H16

C2H5OH

CH4

44.3

44.6

26.9

50.0 120

Lower heating value (MJ/kg) Research octane number

100

0

107

Enthalpy of vaporization (KJ/kg)

272

316

840

Liquid density at 25°C (g/cc)

0.692

0.684

0.785

Initial boiling point (°C)

99

98

78

Final boiling point (°C)

99

98

78

© 2020 SAE International. All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or transmitted, in any form or by any means, electronic, mechanical, photocopying, recording, or otherwise, without the prior written permission of SAE International. Positions and opinions advanced in this work are those of the author(s) and not necessarily those of SAE International. Responsibility for the content of the work lies solely with the author(s). ISSN 0148-7191

© SAE International.

Molecular structure