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 [275]

Table of contents :
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
Introduction
Computational approach
Computational model
Mesh study
Model validation
Results and discussion
Substitution of gasoline with E10 and E85
Diesel energy fraction and SOI2 sweeps for E10/diesel DDFS
Energy fraction and SOI sweeps for near-TDC injection
Effects of diesel and E10 spray angles
Effects of injection pressure of SOI3
Conclusions
CRediT authorship contribution statement
Declaration of Competing Interest
mk:H1_15
References

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Fuel 275 (2020) 117870

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A study of using E10 and E85 under direct dual fuel stratification (DDFS) strategy: Exploring the effects of the reactivity-stratification and diffusionlimited injection on emissions and performance in an E10/diesel DDFS engine

T

Saeid Shirvania, , Sasan Shirvania, Amir H. Shamekhia, Rolf D. Reitzb ⁎

a b

Mechanical Engineering Department, K.N. Toosi University of Technology, No. 7, Pardis St., Molla Sadra Ave., Vanak Sq., Tehran, P.O. Box 19395-1999, Iran Engine Research Center, University of Wisconsin, Madison, Madison, WI, USA

ARTICLE INFO

ABSTRACT

Keywords: DDFS E10/diesel DDFS E85/diesel DDFS, LTC CFD simulation

One promising pathway to directly control combustion is Direct Dual Fuel Stratification (DDFS). DDFS has comparable thermal efficiency to RCCI, acceptable levels of emissions, and lower cyclic variation. The primary drawback of DDFS is soot production owing to the diffusion-limited nature of the near-TDC injection. In this paper, E10 (10% ethanol in gasoline by volume) and E85 were studied as alternative fuels to gasoline to tackle soot formation. In the first step, E10 and E85 were compared, and the best alternative to gasoline was chosen based on emissions and performance. E10 reduced soot by 40%, and E85 eradicated soot completely. However, E85 had 25 times higher NOX than gasoline. Next, diesel energy fraction and its start of injection (SOI2) were swept to explore the domain of the reactivity controlled regime. It was found that for SOI2s before −80° ATDC, the regime was premixed, and for SOI2s after −40° ATDC, the regime was diffusion-limited. Diesel energy fractions, more than 11%, yielded unintended combustion. Then, the energy fraction and injection timing of the near-TDC injection (SOI3) was swept and studied. The best domain for both injectors was discovered based on the EURO6 emission mandate. Spray angles of both injectors were swept from 50 to 80°, and 55° for diesel injector and 65° for the E10 injector showed the best results regarding emissions, performance, and fuel consumption. Finally, the effects of the injection pressure of SOI3 on emissions and performance were studied. The suitable domain for injection pressure was chosen based on EURO6.

1. Introduction The demand for ICEs arose, and humankind approached two crises, first environmental pollution, and second, the growth in fuel cost and extensive energy consumption. The instability of crude oil price and regulations mandate automotive industries to develop engine technology in order to decrease fuel consumption and engine emissions. From the study performed by PF Flynn et al. [1], it was found that Conventional Diesel Combustion (CDC) supplies a very rich mixture and with a diffusion-limited flame. This diffusion flame causes the formation of the Particulate Matter (PM) in rich zones, and high local temperature regions contribute to nitrogen oxides (NOX) formation. The levels of PM and NOX in CDC is very high, but by utilizing aftertreatments, there is a possibility to reach the high standards of emission mandates. Aftertreatments need to reach a standard temperature to have desirable conversion efficiency. The cost of implementation of ⁎

these aftertreatments and being a function of exhaust temperature were motivations to reach more efficient and cleaner combustion. Low Temperature Combustion (LTC) is a pathway to reach more efficient and cleaner combustion than CDC. Many strategies, such as Homogenous Charge Compression Ignition (HCCI), Premixed Charge Compression Ignition (PCCI), and Reactivity Controlled Compression Ignition (RCCI) are utilized to achieve LTC. LTC strategies result in ultra-low NOX and soot, but a high amount of Unburned Hydrocarbon (UHC) and CO compared to CDC. LTC strategies are controlled by chemical kinetics; thus, there is less authority over the rate and timing of heat release. That being said, they are very sensitive to small changes in boundary conditions [2–4]. Onishi et al. [5] designed a new two-stroke engine using HCCI strategy. The challenging point was high load operation in HCCI combustions. As a matter of fact, controlling HCCI engines was difficult, and it was found that Exhaust Gas Recirculation (EGR) can be a promising

Corresponding author. E-mail address: [email protected] (S. Shirvani).

https://doi.org/10.1016/j.fuel.2020.117870 Received 20 December 2019; Received in revised form 9 April 2020; Accepted 13 April 2020 0016-2361/ © 2020 Elsevier Ltd. All rights reserved.

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Nomenclature AHRR AMR ATDC BTDC CAD CDC CFD CRI DDFS DDM Dur E10/D E85/D ED EE10 EGR EPA EVC EVO FE G/D

HCCI ICE IMEP ISFC IVC IVO KH-RT LHV LTC NOX NTC PCCI PM PPC PPRR RCCI RNG k-ε RON SOI1 SOI2 SOI3 TDC UHC

Apparent Heat Release Rate Adaptive Mesh Refinement After Top Dead Center Before Top Dead Center Crank Angle Degree Conventional Diesel Combustion Computational Fluid Dynamics Common Rail Injector Direct Dual Fuel Stratification Discrete Droplet Model Duration E10/Diesel DDFS E85/Diesel DDFS Diesel energy fraction E10 energy fraction for near-TDC injection Exhaust Gas Recirculation Environmental Protection Agency Exhaust Valve Closing Exhaust Valve Opening Fixed Embedding Gasoline/Diesel DDFS

way. Several years later, Najt and Foster [6] applied HCCI strategy in a four-stroke engine, and they found that by means of controlling intake temperature and EGR rate, it is possible to control HCCI combustion. Short combustion duration related to HCCI strategy is a noticeable problem at high load operations, and some noise metrics such as Peak Pressure Rise Rate (PPRR) and Ringing Intensity (RI) are significantly high at high load operations. HCCI also has its own drawbacks. In the HCCI combustion, when the charge is perfectly homogenous, all regions undergo heat release at the same time. From the Otto perspective, this kind of heat addition is ideal, but extreme pressure oscillation in HCCI cause engine damage and noise. There is also a trade-off between thermal efficiency and noise, which limits the operation of HCCI at different loads. HCCI is kinetically controlled, and its sensitivity to boundary conditions is a challenge as well [7,8]. Scholars have focused on some methods to control the heterogeneity of the combustion chamber to control ignition timing and combustion phasing of HCCI. This led to the emergence of some partially premixed strategies like PCCI. In 1996, the first PCCI engine was manufactured with a port fuel injector that could achieve better authority over combustion control, better efficiency, and a significant reduction in NOX and soot emissions compared to CDC. Many scholars performed exhaustive researches on PCCI, and they found that with one fuel and changing intake temperature, charge stratification, EGR and fuel blending, the PCCI strategy is more controllable than HCCI [2,3,9–11]. In 2006, Kalghatgi et al. [12] performed an investigation to overcome HCCI limitations. They found that the use of fuel with higher resistance to auto-ignition compared to diesel could extend combustion duration without the need for EGR. It was reported that too early injection timings provide a well-mixed charge for combustion, and this method, named Partially Premixed Combustion (PPC), led to lower levels of NOX and soot. Gasoline PPC was found as a new method to control the heat release rate better than diesel HCCI. However, this concept had challenges for a PPC engine fueled by gasoline with an octane number greater than 90 at low loads. In order to achieve better control, spark assistance was provided, and a double injection strategy was used to reach better performance, but NOX and soot were at unacceptable levels [13–17]. From the studies that carried out by Yao et al. [18], it was found that in an HCCI engine, by utilizing two fuels with different Research Octane Numbers (RON) such as iso-octane (RON 100) and n-heptane

Homogenous Charge Compression Ignition Internal Combustion Engine Indicated Mean Effective Pressure Indicated Specific Fuel Consumption Intake Valve Closing Intake Valve Opening Kelvin Helmholtz-Raleigh Taylor Lower Heating Value Low Temperature Combustion Nitrogen Oxides No Time Counter Premixed Charge Compression Ignition Particulate Matter Partially Premixed Combustion Peak Pressure Rise Rate Reactivity Controlled Compression Ignition Re-Normalization Group k-ε Research Octane Number Start Of Injection for first gasoline injection Start Of Injection for diesel fuel Start Of Injection for near-TDC injection Top Dead Center Unburned Hydro Carbon

(RON 0), it is possible to control combustion process, engine performance and emissions at different loads. In 2006, Inagaki et al. [19] introduced a new strategy for controlling PCCI by using two different fuels (different reactivity), and today, it is known as Reactivity Controlled Compression Ignition (RCCI). In fact, in the RCCI, by means of reactivity stratification, combustion starts from regions with high reactivity and it continues to other regions with lower reactivity. By this method, scholars managed to extend combustion duration, and operating range. In addition, more control over combustion phasing was achieved. The emissions of RCCI are significantly low, and at many loads, RCCI can meet EURO6 emission regulations. Moreover, it was found that RCCI can reach to 60% gross thermal efficiency under optimized conditions. RCCI is a more promising method compared to HCCI and PPC because it has a smoother heat release rate, lower PPRR, and noise levels. Using different rates of EGR, different fractions of gasoline, and injection timing led to more authority over combustion control. RCCI combustion offers better control over combustion duration and combustion phasing at moderate loads, but there are significant limitations at higher loads. In order to maintain the combustion phasing at higher loads, the mass fraction of the low reactivity fuel needs to be increased, and this trend proceeds until the high reactivity fuel will diminish. Hence, RCCI is limited by the resistance of the low reactivity fuel against auto-ignition. As a result, controlling the timing and rate of heat release for different LTC strategies such as PPC, PCCI, and RCCI is difficult because they are kinetically controlled and offer low authority on the combustion. In addition, these strategies have other drawbacks such as high UHC and CO emissions compared to conventional diesel combustion, and they have a narrow operating range and impractical boundary conditions [20–25]. Shim et al. [26] performed an experimental study on the comparison between advanced combustion technologies such as HCCI, PCCI, RCCI, and CDC. It was reported that all strategies could decrease NOX and PM simultaneously with the use of aftertreatment. HCCI and PCCI had some challenges to maintain the combustion phasing, so the EGR rate was increased up to 40% to serve the purpose. UHC and CO emissions increased by about 38 times in advanced strategies compared to CDC. In the dual-fuel PCCI, indicated thermal efficiency was peaked at 45.3%, and by using natural gas, CO2 emission reduced by 14.3%. To have more control over combustion and decrease UHC and CO, 2

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in 2013 and 2014, Lim and Reitz [27,28] implemented a direct gasoline injector in the combustion chamber of an RCCI engine. Gasoline was injected in the combustion chamber with a narrow spray angle, and then diesel was injected with a wide spray angle. As a result, UHC and CO emissions were decreased by about 7% and 27%. In addition, combustion efficiency improved by this method. They adopted this method and managed to extend the RCCI operating range to 21 bar gross IMEP. In 2015 Wissink implemented two direct injectors in the combustion chamber to have more authority over the combustion phasing and the rate of heat release to overcome RCCI limitations. He introduced a new strategy called Direct Dual Fuel Stratification (DDFS). In this strategy, gasoline is injected directly in the combustion chamber at 340 °BTDC to make a partially homogenous charge, and like RCCI strategy, diesel fuel is injected directly in the combustion chamber 40–60 °BTDC to prepare a reactivity gradient to achieve precise control over the Low Temperature Heat Release (LTHR). After LTHR, a near-TDC injection makes more control over the rate of heat release. DDFS has comparable efficiency to RCCI, and emissions were at acceptable levels. In addition, pressure rise rate, noise level, EGR, and cyclic variation were improved in DDFS compared to RCCI. Wissink also compared DDFS with PPC and found that DDFS has higher efficiency by about 15%. One big concern of the DDFS is high PM production due to the nature of diffusion-limited flame of the near-TDC injection. Wissink adopted many methods to reduce PM, especially for high load operations. PM reduction methods that he used included post injection, increasing injection pressure, decreasing injection duration or using gasoline instead of diesel for the near-TDC injection. Post injection decreased PM, but it increased NOX. The same behavior was observed for the increasing injection pressure [29–32]. In 2017, Loung et al. [33] performed a numerical investigation on the effects of injection timing on the ignition of iso-octane using DDFS strategy and compared with RCCI. They used a 2D direct numerical simulation (2D DNS) with reduced chemistry of primary reference fuel oxidization. They found that the RCCI combustion exhibited shorter combustion duration with highly peaked heat release rate, while the DDFS combustion had a lower peak of heat release and longer combustion duration, which was attributed to the sequential injection of isooctane. They reported that the DDFS strategy can allow precise control over the rate of heat release and combustion phasing when compared with RCCI, and by adjusting the near-TDC injection and proper energy fraction, this purpose can be served. In 2020, Shirvani et al. [34] performed an investigation on the effects of piston-profile on emissions and performance under DDFS strategy. They found that two geometrical parameters, including the ratios of squish volume to clearance volume, and piston area to chamber area have significant impacts on emission formation. They investigated three conventional piston profiles, and then combined the benefits of them and proposed a new piston profile, which has small squish volume and omega type spray-guide shape on the piston. This piston is suitable for the combustion chamber with two direct injectors using the DDFS strategy. They managed to reduce NOX, soot, and CO by about 17, 2, and 27%, respectively, with the proposed piston. Due to global energy demands, energy consumption has been increased, which is combined with greenhouse gas and global warming concerns. Biomass is renewable and can be considered as an alternative to fossil fuels, which is suitable for satisfying a part of energy demands. Biomass is utilized for direct combustion in power production, but it has some drawbacks. Instead of using biomass for direct combustion, some gasification techniques have shown significant heed by alleviating the severe changes in climate and energy crisis. Biomass gasification can reduce hazardous gases significantly and can satisfy energy demands. Syngas contains hydrogen (H2), methane (CH4), carbon dioxide (CO2), carbon monoxide (CO), and nitrogen (N2) and water vapor. Some eminent scholars performed experimental and numerical investigations on biomass and syngas production to finalize syngas

production for various applications. H2 can be used as a clean source of energy, and CH4 from biomass gasification can be used easily in the transportation industry under LTC strategies like RCCI [35–37]. Another renewable source of energy that is very attractive for the transportation industry is alcohol fuels like methanol and ethanol. Alcohol fuels can reduce the tradeoff between NOX and soot in conventional diesel combustion. Ethanol is used widely in many countries on a large scale as a renewable fuel and a suitable alternative to fossil fuels. There are two common ways of ethanol production: ethylene hydration and fermentation. In the first way, ethanol is widely produced for industrial applications. Ethylene is reacted with steam and by using some chemicals such as phosphoric and sulfuric acids, and as a result, ethanol is produced. In the fermentation method, bioethanol is produced from food crops such as corn grain and sugarcane, but it may be limited because it may lead to a tradeoff between bioethanol production and food provisions. Alcohol fuels have some advantages: higher octane number, renewability, evaporative cooling effects, which leads to higher volumetric efficiency and lower combustion temperature, low sulfur content compared to gasoline, and containing oxygen facilitates fuel oxidization [38,39]. Ethanol can be used with pure gasoline in Spark Ignition (SI) engines for different compression ratios. It was reported that the application of alcohol in SI engines under optimized conditions can enhance engine brake thermal efficiency and emissions when compared with baseline [40]. Diesel engines generally have higher efficiency compared to gasoline engines, but this efficiency is a trade-off between soot and NOX. In fact, a simultaneous reduction in NOX and soot is difficult in conventional diesel combustion. Many scholars have studied the effects of using ethanol in diesel engines not only as a substitution for diesel but also for reducing NOX and soot simultaneously [41–44]. From the studies performed by scholars on the effects of ethanol/diesel blend properties, it was found that 30% ethanol by volume in diesel not only has little effects on fuel atomization but also increases ignition delay and improves combustion efficiency. In addition, NOX and soot levels can be improved under certain conditions [45–48]. Other researches were performed on the effects of different alcohol fuels such as methanol and butanol in diesel fuel. It was found that the application of alcohol fuels increases fuel oxidization, NOX, and thermal efficiency while it significantly reduces soot [49–51]. Li et al. [52] conducted a multi-objective optimization of methanol/ diesel in a dual fuel engine to find the baselines of optimized RCCI and DDFS at low load operating conditions. The model was validated against experimental data of a light-duty RCCI engine, and the strategy was changed numerically from RCCI to DDFS. It was found that DDFS has higher thermal efficiency, lower UHC and CO, and lower sensitivity to initial conditions at low loads. Moreover, in RCCI, a large spray included angle is preferred, while in DDFS, a small spray-included angle is more desirable for fuel oxidization. Long et al. [53] performed an experimental investigation on the effects of dual-direct injection in a light-duty engine. It was reported that diesel fuel jet-injection could control the combustion phasing and ignition timing robustly for four operating conditions. Advancing the near-TDC injection timing resulted in high gross thermal efficiency and high NOX emissions. For prolonged near-TDC injection, there is a trade-off between NOX and soot like CDC. As previously mentioned, the DDFS strategy is a practical pathway in internal combustion engines to achieve ultra-low emissions and high thermal efficiency. Unlike RCCI and HCCI, which are kinetically controlled, DDFS has great authority over the rate of heat release and combustion. The advantages of the DDFS outweighed the disadvantages except high PM production. One practical way to address this problem is using fuels with low C/H ratio, or oxygenated fuels such as ethanol. In this paper, E85 and E10 were chosen as alternatives to gasoline in the gasoline/diesel DDFS combustion. A numerical 3D-CFD model was developed and validated against experimental data. To the best of our knowledge, there is little information about using ethanol under the DDFS strategy to tackle PM. A comparative study was performed 3

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between E10 and E85 as alternatives to gasoline regarding emissions and performance criteria. The paper is a comprehensive study on the effects of diesel energy fraction and its timing on emissions and performance. In addition, the energy fraction of the near-TDC injection and its timing was swept, and the best region based on the EURO6 emissions mandate was chosen. The DDFS combustion can meet EURO6 without aftertreatment. The combustion chamber has two direct injectors, and their spray angles were swept, and the best regions for the spray angles were also chosen. Finally, the effects of injection pressure of the nearTDC injection were studied, and the suitable range of injection pressure for DDFS combustion was chosen based on the EURO6 emissions regulation.

cylinder, and crevice volume, which is a source of UHC. As Senecal et al. recommended in [65], the optimum grid size for spray simulation was about 0.25 mm; in this study, appropriate mesh resolution was performed for spray to reach this size. In this computational model, the minimum number of the grid for the compression stroke at TDC was about 252,000, and for the expansion stroke, it was about 960,000 cells. To ensure that the CFD model is reliable for the gasoline/diesel and E85/diesel combustion, the model was used for other loads in Appendix. The numerical results showed a reliable model and enough accuracy. The combustion chamber has two direct injectors, so the whole combustion chamber was modeled, and Fig. 1 shows the profile of the combustion chamber for the DDFS engine with crevice volume at TDC as well as the injectors’ positions.

2. Computational approach

2.2. Mesh study

In this study, a 3D-CFD model was developed and used to simulate DDFS combustion of the Caterpillar single cylinder engine. Mesh study was performed for three mesh sizes: coarse mesh (2 mm), medium mesh (1.4 mm), and fine mesh (1 mm). The best mesh size was chosen based on mesh accuracy and computational time. Experimental tests were performed by Wissink and Reitz at the University of Wisconsin Madison Lab [29,30]. The engine specification and operating conditions of the experiments are outlined in Tables 1 and 2, respectively.

The CFD computational model of DDFS combustion was validated against experimental data given in [29]. In order to examine its accuracy, mesh independency of the numerical model was studied. Fig. 2 illustrates a comparison between the numerical results and experimental data for three mesh sizes. Table 3 shows the results of mesh independency study for emissions. According to Fig. 2 and Table 3, the appropriate mesh size based on the accuracy and computational time was considered 1.4 mm.

2.1. Computational model

2.3. Model validation

In this paper, the CFD model was employed to simulate the DDFS combustion at 9.41 bar gross IMEP and 1300 rpm. This simulation considers closed-cycle for combustion simulation from IVC (143 °BTDC) to EVO (130 °ATDC). The gasoline charge was considered completely homogeneous at IVC for the simulation. A Lagrangian approach was employed to simulate spray dynamics from injection until vaporization. A drop drag sub-model, which the drop drag coefficient can be changed dynamically with flow conditions, was used in the CFD model [54]. Standard Discrete Droplet Model (DDM) was applied, which uses a technique for simulating the behavior of atomized evaporating liquid parcels in the gaseous environment [55]. The Kelvin-Helmholtz and Rayleigh-Taylor (KH-RT) hybrid model were used to simulate spray atomization for the primary and secondary break-ups. This model is capable of being used for both diesel and gasoline spray, especially for the DDFS strategy [56]. Both KH and RT instabilities were used to predict the secondary breakup of each droplet. O’Rourke submodel was applied for spray wall interaction and wall film [57]. The No Time Counter (NTC) method, which is faster than the O’Rourke model, was employed to calculate droplet collisions in Lagrangian spray simulation [58]. The Re-Normalization Group (RNG) k-ε model was used to simulate turbulence transport. This model is used for internal combustion engines, especially for diesel engines, which have compressing-expanding flows and spray combustion [59]. Detailed chemical kinetics using multi-zone modeling was used in this paper. This model accelerates combustion process, and it is appropriate for gasoline and diesel fuels and also for multi-component fuels [60]. The multi-zone chemistry solver was used to speed up combustion calculations for grouped cells with similar properties. To simulate engine combustion, a reduced chemistry mechanism for gasoline/diesel, which simulate diesel as nheptane and gasoline as iso-octane with 108 species and 435 reactions [30,61]. In order to predict the soot concentration, Hiroyasu soot model was used, and the extended Zeldovich procedure was applied to predict NOX emissions [62,63]. Fixed wall boundary conditions were used for the head cylinder and cylinder wall with a smooth wall assumption. The Law of the Wall boundary for temperature and velocity was applied for all wall boundaries [64]. The base grid size for global mesh was chosen 1.4 mm based on grid independency study. Some local refinements were done for some critical regions such as nozzles’ regions, cylinder wall, head

The numerical 3D-CFD model was validated against experimental data and Fig. 3 shows a comparison of the cylinder pressure and AHRR between numerical results and experimental data. Fig. 4 depicts the emissions results of the numerical model and experimental tests. 3. Results and discussion As previously mentioned, the DDFS strategy has supremacy over RCCI except its PM production owing to the near-TDC injection. In order to decrease the soot formation, and reduce the trade-off between NOX and soot in diffusion-limited combustion, it is proposed to use E10 Table 1 Engine geometry and direct injector specifications for the experimental setup (experimental data taken from [29]). Engine Specifications Engine Type Piston type Displacement (L) Bore (mm) Stroke (mm) Connecting rod length (mm) Squish height (mm) Number of valves per cylinder IVO (°ATDC) IVC (°ATDC) EVO (°ATDC) EVC (°ATDC) Swirl ratio Compression ratio Common rail injectors name (CRI1) Body style Nozzle angle (°) Hole diameter (μm) Number of holes Common rail injectors name (CRI2) Body style Nozzle angle (°) Hole diameter (μm) Number of holes

4

Caterpillar 3401E Single Cylinder Oil Test Engine (SCOTE) Modified-piston (wide shallow) 2.44 137.2 165.1 261.6 1.57 4 335 −143 130 −355 0.7 14.88:1 Bosch CRI2 series 148 141 7 Bosch CRI2 series 143 117 10

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Table 2 Experimental operating conditions used for the DDFS experiments (experimental data taken from [29]). Operating Conditions EGR (%) TEGR (°C) Tin (°C) Pin (kPa) (supercharged) Equivalence ratio Gross IMEP (bar) Qfuel (kJ/cyc) Fuel1 Injector name Injection pressure (bar) SOI2 (°BTDC) Dur2 (ms) Total energy ratio (%) Fuel2 Injector name Injection pressure (bar) SOI1 (°BTDC) Dur1 (ms) SOI3 (°BTDC) Dur3 (ms)

39.5 90.3 49.3 186.2 0.57 9.41 4.71 Diesel CRI1 500 60 0.6 7.0 Gasoline CRI2 1000 340 1.4 4 0.8

Fig. 2. Comparison of cylinder pressure and AHRR for different mesh sizes using DDFS strategy.

oxygen, and oxygen molecules lead to faster fuel oxidization and more rates of heat release in a shorter duration. Thus, the peak of AHRR increased. In the E85/D case, AHRR increased dramatically, and it resulted in a dramatic increase in the cylinder pressure. The PPRR of the E85/D case is 21.6 bar/deg and much higher than normal levels, given in Table 5. Emissions, gross IMEP, and Gross Thermal Efficiency (GTE) are presented in Table 5. In the E10/D case, NOX slightly increased by about 1.7% compared to the G/D case. However, soot significantly reduced by about 40%. This reduction can be attributed to the fact that oxygen molecules in E10 accelerate fuel oxidization and prevent the formation of soot and carbon clusters. UHC and CO in the E10/D case reduced by about 23.5%, 38.6%, respectively. Numerous studies were performed on the effects of E85 under the RCCI strategy; E85 is absolutely helpful for the RCCI strategy owing to its higher octane number than gasoline and making a higher reactivity gradient. However, in the direct injection near-TDC, this fuel results in a sudden and very quick increase in the rate of heat release and NOX emissions. In the E85/D DDFS case, NOX dramatically increased by approximately 25 orders of magnitude compared to the G/D case. As a result, E10 is more suitable than E85 for DDFS combustion. Soot significantly reduced in the E85/D case, and CO reduced by about 83.9% owing to this fact that E85 has fewer carbons in its structure compared to E10 and gasoline. UHC emission in the E85/D case was slightly higher than E10/D case because of the cooling effects of E85 on the combustion chamber, shown in Fig. 7. For constant in-cylinder energy in each case, the gross IMEP increased for the alcohol cases and resulted in higher GTE. Fig. 8 depicts temperature cut-planes at similar combustion phasing (CA50 = 3.9° ATDC) for all cases. As it is evident, the E85/D DDFS case has the highest local temperature zones which associate with thermal NOX formation. The E85 has much more oxygen molecules that resulted in higher and faster AHRR. Consequently, these high temperature zones justify the high amount of NOX and better oxidization for soot in the E85/D case compared to the E10/D and G/D cases. Fig. 9 shows isothermal surfaces (2000 and 2200 K) for all cases. The effects of E85 and

(10% ethanol in gasoline by volume) and E85 (85% ethanol in gasoline by volume) as two conventional alcohol fuels. In the first step, gasoline was substituted by E10 and E85, and a comparative study was performed to investigate their effects on emissions and performance. In the second step, one of them was chosen, and a numerical study on the injection timing and sweeps of energy fraction for SOI2 and SOI3 will be studied in the following sections. Fig. 5 shows the problem formulation in this study. 3.1. Substitution of gasoline with E10 and E85 In this section, the effects of using E10 and E85 as the low reactivity fuels are investigated. In-cylinder energy for Gasoline/Diesel (G/D), E10/Diesel (E10/D), and E85/Diesel (E85/D) using the DDFS strategy was considered constant. CA50 for the G/D case was about 3.9° ATDC. To make a fair comparison, energy fractions for the E10/D and E85/D were considered constant similar to the G/D case, and by changing the near-TDC Start Of Injection (SOI3), CA50 was kept constant for all cases. Simulation parameters for the G/D, E10/D, and E85/D cases are provided in Table 4. The EGR rate for the E85/D case was considered zero owing to the fact that E85 has a higher octane number and it has cooling effects on the combustion chamber. In general, for 9.41 bar IMEP, using EGR in E85/D is unnecessary according to the experimental tests provided in [4,66,67]. Diesel injection strategy, including SOI2, diesel energy fraction, and injection pressure and duration were considered constant in this section. The cooling effects of ethanol vaporization in the first injection (SOI1 = -340° ATDC) on the cylinder temperature were calculated in Appendix, and TIVC for closed-cycle simulation for each case is provided in Table 4. Fig. 6 illustrates the comparison of cylinder pressure and AHRR for all cases. Despite the same EGR rate for the E10/D case compared to the G/D case, the peak of AHRR of this case slightly increased. E10 contains

Fig. 1. Piston geometry profile with the crevice volume for the combustion chamber at TDC.

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Table 3 Numerical results of emissions for different mesh sizes using DDFS strategy (system configuration is Intel Corei7 6700 k with 32 GB RAM). Case Coarse mesh Medium mesh Fine mesh

base grid size (dx,dy,dz)

NOX(gr/kW-hr)

Soot (gr/kW-hr)

UHC (gr/kW-hr)

CO (gr/kW-hr)

Computational time

Max. cell Min. cell

2 mm 1.4 mm 1 mm

0.291 0.280 0.280

0.003 0.005 0.005

1.55 1.66 1.66

4.79 4.90 4.90

~13 h ~20 h ~41 h

~690,000~126,000 ~960,000~252,000 ~2,580,000~462,000

Generally, LTC strategies have a lower exhaust gas temperature compared to conventional diesel combustion, so the application of aftertreatment and turbocharger is challenging. In short, DDFS has greater potential for downstream recovery compared to RCCI owing to the near-TDC injection. Using E10 has negligible effects on the heat recovery parameter, but E85 significantly decreases the exhaust mean temperature, which is not suitable for the heat recovery parameter. 3.2. Diesel energy fraction and SOI2 sweeps for E10/diesel DDFS In this section, the diesel energy fraction and its start of injection (SOI2) were swept. As it is visualized in Fig. 10, diesel fuel supplies 7% of the total cylinder energy at −60° ATDC. Diesel energy fraction was swept from 0 to 11% of total energy by subtracting the energy from the first injection at −340° ATDC. Then, the diesel start of injection was swept from −100° ATDC to −30° ATDC to investigate its effects on the charge stratification. Fig. 11 shows the cylinder pressure and AHRR for the cases of 0%, 4%, 7%, and 11% diesel energy fractions. AHRR and peak of cylinder pressure advanced as the fraction of diesel increased. As it is evident, for the 11% diesel energy fraction, the peak of AHRR occurred before TDC, and it resulted in negative work. Moreover, the peak of low temperature heat release, in this case, was rather high, which means there was diffusion-limited combustion at 20 CAD before TDC. The diesel energy fraction of 11% and magnitudes above 11% is not suitable for E10/D DDFS combustion at this engine load. As the amount of diesel fuel decreases, the peak of cylinder pressure and AHRR retard, and in the 0% case, the strategy changes from DDFS to PPC. Fig. 12 illustrates the cylinder pressure and AHRR traces for E10/D DDFS with 7% diesel energy fraction and with a sweep of diesel start of injection (SOI2). By advancing SOI2, the cylinder pressures and AHRRs retarded. In addition, the peak of AHRR increased as the SOI2 advanced. SOI2 had significant effects on the in-cylinder stratification and reactivity gradient. Three regimes occurred before the low temperature heat release by changing SOI2. These regimes are premixed (PPC-like combustion), reactivity controlled, and diffusion-limited. As it is shown, the case with SOI2 = -40° ATDC, had a sudden increase in AHRR and it can be concluded for SOI2 = -40° ATDC, and after that, the regime was diffusion-limited. For the cases with advanced SOI2, homogeneity increased, and the regime changed into a premixed regime (PPC-like regime). The characteristic of this regime is a high peak of heat release rate and

Fig. 3. Comparison of cylinder pressure and AHRR between numerical results and experimental data.

E10 on the flame shape are obviously shown in this figure. As previously mentioned, improved heat release rate, less cyclic variation, comparable gross thermal efficiency to RCCI, lower engine noise levels in high load operations, and higher great potential heat recovery from downstream is achievable in the DDFS strategy. The primary drawback of the DDFS is high soot production compared to RCCI strategy. Table 6 presents a qualitative comparison of emissions and other performance parameters between E10/D and E85/D compared to the G/D as the base case under DDFS strategy. In this table, zero is negligible variation, plus conveys an improvement, and minus shows an aggravation for each case compared to G/D DDFS case. The table shows that E10 is a better alternative to gasoline for DDFS strategy because not only soot, UHC + CO, and GTE improved, but also NOX, ISFC, and PPRR had negligible changes. As a result, by using E10, soot as the primary drawback of the DDFS can be reduced significantly up to 40% at this load. The E85 also reduced soot production, but at the expense of 25 times higher NOX. PPRR and fuel consumption are much higher than E10. The heat recovery parameter in this table presents an important concept in internal combustion engines. In order to reach higher loads, it is necessary to have high boost pressure for the intake manifold.

Exp. data 0.6 0.5 0.4 0.3 0.2 0.1 0

Num. results

Exp. data 6 5 4 3 2 1 0

0.47 0.5 0.291 0.28

NOx (gr/kW-hr)

soot×100 (gr/kW-hr)

Num. results 5.06

2.15

4.9

1.66

UHC (gr/kW-hr)

CO (gr/kW-hr)

Fig. 4. Comparison of the emissions between numerical results and experimental data. 6

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Fig. 5. Problem formulation in this study. Table 4 Simulation parameters for the G/D, E10/D, and E85/D cases using DDFS strategy.

EGR (%) TIVC (°C) ISFC (gr/kW-hr) Qfuel (kJ/cyc) Fuel1 Injector name Injection pressure (bar) SOI2 (°BTDC) Dur2 (ms) Total energy ratio (%) Fuel2 Injector name Injection pressure (bar) SOI3 (°BTDC) Dur3 (ms)

G/D DDFS

E10/D DDFS

E85/D DDFS

39.5 380 158.4 4.71 Diesel CRI1 500 60 0.6 7.0 Gasoline CRI2 1000 4.0 0.8

39.5 379.2 162.4 4.71 Diesel CRI1 500 60 0.6 7.0 E10 CRI2 1000 4.5 0.85

0 370.2 225.9 4.71 Diesel CRI1 500 60 0.6 7.0 E85 CRI2 1000 7.0 1.1

Fig. 6. Comparison of cylinder pressure and AHRR for G/D, E10/D, and E85/D using DDFS strategy.

combustion becomes retarded and resembles PPC combustion. As can be seen from the shape of AHHR (start, end, and peak) in Fig. 12, reactivity stratification, which is adjusted by SOI2, affects the rate of heat release, its start, end, and also emission formation. Adjusting reactivity stratification, which is a benefit of RCCI, makes it possible to manage and control the combustion. Having this benefit from RCCI is a great characteristic of DDFS. The second phase of combustion control is the near-TDC injection. Fig. 13 shows emission contours for the E10/D case with sweeps of SOI2 and diesel energy fraction (Ed). The upper limit of the Ed was

considered 11% of the total cylinder energy because, as it is shown in Fig. 12, for the magnitudes of Ed higher than 11%, combustion became advanced, and AHRR peaked before TDC, which was not suitable and resulted in negative work. As previously mentioned, three regimes occurred by sweeps of SOI2 and Ed. Diffusion-limited and PPC-like regimes are demonstrated in Fig. 13. The characteristic of the diffusionlimited regime is high amounts of NOX and soot emissions, and the PPClike regime has high amounts of UHC and CO. For SOI2s between −80 7

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Table 5 Emissions and performance for the G/D, E10/D, and E85/D using DDFS strategy.

G/D E10/D E85/D

IMEP (bar)

GTE (%)

NOX (gr/kW-h)

Soot (gr/kW-h)

UHC (gr/kW-h)

CO (gr/kW-h)

CA50 (°ATDC)

PPRR (bar/deg)

10.15 10.25 10.52

51.3 52.4 55.9

0.280 0.285 7.037

0.005 0.003 0.00001

1.66 1.27 1.44

4.90 3.01 0.79

3.9 3.9 3.9

8.0 9.1 21.6

between −80 to −40° ATDC) and the diffusion-limited regime (-40 to −30° ATDC) had the lowest amounts of PPRR. The reactivity-controlled regime had the highest. In this region, ISFC was at its the lowest level. To summarize this section, reactivity-controlled regime (SOI2 between −80 to −40° ATDC), shows the best results of emissions, GTE, PPRR and fuel consumption for the E10/D DDFS combustion. This region benefits from the RCCI concepts and its characteristics for better results of emissions and performance and more authority over combustion. 3.3. Energy fraction and SOI sweeps for near-TDC injection In this section, the energy fraction and SOI3 of E10 for the near-TDC injection were swept. Energy fraction sweeps were 20 to 40% of the total cylinder energy by subtracting from the first E10 injection. Fig. 15 shows cylinder pressure and AHRR traces for energy faction sweeps at fixed SOI3 = -4.5° ATDC. AHRR was extended by increasing the nearTDC energy fraction. The peak of AHRR and cylinder pressure drops, and it resulted in retarded combustion phasing. The opposite trend is evident by reducing the energy fraction. Fig. 16 illustrates the sweeps of SOI3 for a fixed energy fraction of 34%. SOI3 was swept from −8° ATDC to −2° ATDC. As SOI3 advanced, the peak of cylinder pressure and AHRR increased, this was combined with advanced CA50. The effects of E10 energy fraction and SOI3 sweeps on emissions and performance are depicted in Figs. 17 and 18. As previously mentioned, advancing SOI3 resulted in higher cylinder pressure and AHRR. Thus, thermal NOX increased, shown in Fig. 17. On the other hand, soot, UHC, and CO decreased by advancing SOI3 owing to higher cylinder pressure and temperature and better oxidization. By retarding SOI3, soot, UHC and CO increased owing to incomplete combustion. From the emissions’ perspective and EURO6 emission mandate (NOX = 0.4 g/ kW-h and soot = 0.01 g/kW-h), suitable SOI3 and E10 energy fraction are −4 to −2° ATDC and 20 to 30% energy fraction, respectively, to meet this standard without aftertreatment in the E10/D DDFS combustion. Fig. 18 shows the effects of energy fraction and SOI3 of the nearTDC injection’s sweeps on operating emissions and performance. As previously mentioned, advanced SOI3 yielded higher cylinder pressure, and higher GTE, owing to higher output work. GTE increased by advancing SOI3, but at the expense of higher engine noise and PPRR.

Fig. 7. Comparison of mean temperature diagrams between G/D, E10/D, and E85/D using DDFS strategy.

to −40° ATDC, the regime is reactivity-controlled, which is the best regime for DDFS that can be obtained by adjusting SOI2 and diesel energy fraction. As the EURO6 emission standard is 0.4 g/kW-h and 0.01 g/kW-h for NOX and soot,respectively, this engine can meet EURO6 emission mandate without using aftertreatments by adopting E10 as the low-reactivity fuel using DDFS strategy. However, for the Ed above 8% and SOI2s after −40° ATDC (diffusion-limited regime), the engine cannot meet EURO6 for NOX. As it is illustrated in Fig. 13, UHC and CO amounts are rather high in PPC-like regime owing to higher homogeneity of the charge that can result in incomplete combustion compared to the reactivity-controlled regime. From all emissions perspective, the suitable regions for choosing Ed and SOI2 are 4 to 8% of and −80 to −40° ATDC (reactivity-controlled regime). Fig. 14 depicts the effects of SOI2, and Ed sweeps on GTE, PPRR, and indicated specific fuel consumption. The PPC-like regime had higher PPRR compared to other regimes, which indicates higher noise levels. When the amount of diesel fuel decreased with early diesel injection (SOI2 = -100° ATDC), owing to higher homoginity of the charge, PPRR increased. On the other hand, the reactivity-controlled regime (SOI2

G/D DDFS

E10/D DDFS

E85/D DDFS

Temperature (K)

Fig. 8. Temperature cut-planes for G/D, E10/D, and E85/D at similar CA50 (3.9° ATDC) using DDFS strategy. 8

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G/D DDFS

E10/D DDFS

E85/D DDFS

Fig. 9. Iso-thermal surfaces (2000 and 2200 K) for G/D, E10/D, and E85/D at 8° ATDC using DDFS strategy. Table 6 Qualitative comparison of emissions, gross thermal efficiency, ISFC, PPRR, and the heat recovery for E85/D and E10/D under DDFS strategy. Case E10/D DDFS E85/D DDFS

NOX

Soot

UHC + CO

GTE

ISFC

PPRR

Heat recovery

0 –

+ +

+ +

+ +

0 –

0 –

0 –

E10

80 60

Diesel

59

34

40 20

0

7 -340

-200

-100 -60

Diesel energy fraction sweeps (%): 0 to 11

-4.5 0 Fig. 12. Comparison of cylinder pressure and AHRR for E10/D DDFS with a sweep of SOI2 at fixed ED = 7%.

SOI2 sweeps (°ATDC): -100 to -30

Fig. 10. Visualization of energy fraction and start of injection for E10/D DDFS (base case), and sweeps range.

Fig. 11. Comparison of cylinder pressure and AHRR for E10/D DDFS with a sweep of diesel energy fraction at fixed SOI2 = −60° ATDC. Fig. 13. Emission contours for E10/D DDFS with sweeps of diesel injection (SOI2) and diesel energy fraction (Ed) with fixed SOI3 = −4.5° ATDC and EE10 = 34%.

Retarded SOI3 and high EE10 resulted in high fuel consumption. From the emission, GTE and fuel consumption perspectives, the suitable domain for choosing SOI3 and EE10, is −4 to −2° ATDC, and 20 to 30% E10 energy fraction for the E10/D DDFS combustion.

angle (θg) on emissions and performance are investigated. In the first step, θd was swept between 50 and 80° with fixed diesel energy fraction and variable SOI2. Other fixed assumptions are SOI3 = -4.5° ATDC and EE10 = 34% for the near-TDC injection. Fig. 19 shows emission contours for the sweeps of θd and SOI2. As previously mentioned, the reactivity-controlled regime occurred between SOI2 = -80 to −40° ATDC. Thus, in this section, the sweeps of

3.4. Effects of diesel and E10 spray angles As it is shown in Fig. 1, the combustion chamber has two direct injectors with two spray angles that can affect emission formation. In this section, the effects of the diesel spray angle (θd) and E10 spray 9

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Fig. 16. Comparison of cylinder pressure and AHRR for E10/D DDFS with a sweep of the near-TDC injection (SOI3), at fixed energy fraction of 34%.

Fig. 14. Performance contours for E10/D DDFS with sweeps of diesel injection (SOI2) and diesel energy fraction (Ed) with fixed SOI3 = -4.5° ATDC and EE10 = 34%.

Fig. 17. Emission contours for E10/D DDFS with sweeps of near-TDC injection (SOI3) and E10 energy fraction (EE10) with fixed SOI2 = −60° ATDC and Ed = 7%.

Fig. 15. Comparison of cylinder pressure and AHRR for E10/D DDFS with a sweep energy fraction of the injection near TDC (SOI3 = −4.5° ATDC).

Figs. 21 and 22 illustrate the emission and performance contours with sweeps of θg and SOI3 for the E10/D DDFS combustion with fixed Ed = 7%, SOI2 = -60° ATDC, and EE10 = 34%. By advancing SOI3 to −8° ATDC, NOX significantly increased, and when θg is between 60 and 70°, NOX had the highest amount by about two times than the EURO6 emission mandate. From all emissions’ perspectives, SOI3 between −4 to −2° ATDC and θg between 60 and 70° have the best results for emissions. In addition, this region has the best result for GTE and ISFC shown in Fig. 22. PPRR for this region is less than 10 bar/deg, which means acceptable engine noise level. In brief, diesel and E10 spray angles affect emission formations and engine performance. The study showed that for medium spray angles for both spray angles is more preferable regarding emission and performance criteria. In short, for diesel spray angle (θd) between 55 and 60° and diesel start of injection (SOI2) between −65 to −55° ATDC indicated the best results of emissions and performance. For the nearTDC injection, E10 spray angle between 60 and 70° and near-TDC

SOI2 was considered for −80 to −40° ATDC. The highest amount of NOX occurred for θd = 70°, while θd = 55° had the best results for NOX. θd = 55° and SOI2 between −65 to −55° ATDC showed the best results for UHC, CO and soot emissions. From the emission’s standpoint, the θd between 55 and 60° and SOI2 between −65 to −55° ATDC, are suitable intervals to achive lower emissions in the E10/D DDFS combustion. Fig. 20 shows the performance contours for the sweeps of θd and SOI2. GTE had the best results for the regions where UHC and CO had the lowest amounts, indicating that UHC and CO oxidization resulted in higher output work and higher GTE. In addition, fuel consumption had the lowest magnitude in this region. As previously mentioned, the suitable interval for choosing θd and SOI2 are between 55 and 60° and −65 to −55° ATDC. This region satisfies both emissions and engine performance. PPRR has acceptable levels for this region (less than 10 bar/deg) and satisfies the engine noise. In the second step, the effects of the spray angle of the near-TDC injection of E10 (θg) on emissions and performance are investigated. 10

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Fig. 20. Performance contours for E10/D DDFS with sweeps of diesel spray angle (θd) and SOI2 with fixed Ed = 7%, SOI3 = −4° ATDC, and EE10 = 34%.

Fig. 18. Performance contours for E10/D DDFS with sweeps of near-TDC injection (SOI3) and E10 energy fraction (EE10) with fixed SOI2 = −60° ATDC and Ed = 7%.

Fig. 21. Emission contours for E10/D DDFS with sweeps of E10 spray angle of the near-TDC injection (θg) and SOI3 with fixed Ed = 7%, SOI2 = −60° ATDC, and EE10 = 34%.

Fig. 19. Emission contours for E10/D DDFS with sweeps of diesel spray angle (θd) and SOI2 with fixed Ed = 7%, SOI3 = −4° ATDC, and EE10 = 34%.

injection timing between −4 to −2° ATDC showed the best results of emissions and performance.

500, 1000, and 1500 bar injection pressure. The peak of AHRR and cylinder pressure increased as the injection pressure increased. Fig. 24 shows emissions and performance for third injection pressure sweeps. As the injection pressure increased, NOX increased dramatically by about 50% in the 1500 bar case compared to the 1000 bar (base case). On the other hand, no significant soot reduction was observed. NOX does not meet EURO6 HD for injection pressures above 1250 bar. As the injection pressure increased, UHC and CO decreased dramatically, owing to the better fuel atomization and oxidization. According to this figure, gross IMEP increased by boosting injection

3.5. Effects of injection pressure of SOI3 One of the drawbacks of DDFS strategy is soot production owing to the diffusion-limited nature of the near-TDC injection. In this section, the effects of the injection pressure of the third injection (E10 injection near TDC) are investigated. The injection pressure was varied from 500 bar to 1500 bar. The increase in injection pressure atomizes fuel properly and leads to a faster rate of heat release and facilitates soot oxidization. Fig. 23 illustrates cylinder pressure and AHRR traces for 11

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owing to the diffusion-limited nature of the near-TDC injection. One practical approach towards this drawback is the use of alcohol fuels (E10 or E85) as alternatives to gasoline. In this paper, a 3D-CFD model was developed and validated against experimental data in order to simulate the combustion process of a heavy-duty engine using DDFS strategy at 9.41 bar gross IMEP and 1300 rpm. The main results are as follows:

• At the first step, E10 and E85, as two conventional candidates as

• Fig. 22. Performance contours for E10/D DDFS with sweeps of E10 spray angle of the near-TDC injection (θg) and SOI3 with fixed Ed = 7%, SOI2 = −60° ATDC, and EE10 = 34%.

•

•

Fig. 23. Comparison of cylinder pressure and AHRR with a sweep of injection pressure for the near-TDC injection.

•

pressure, but a negligible change in GTE (about 0.3%) was observed. Higher injection pressure advanced CA50 and led to higher PPRR by about 2% in the 1500 bar case compared to the base case (1000 bar). Fig. 25 depicts the iso-thermal surfaces for 500, 1000, and 1500 bar injection pressures. Higher injection pressure resulted in better flame penetration and more numbers of high-temperature zones. This figure also shows the differences in the flame shapes as a result of increasing injection pressure.

•

4. Conclusions The Direct Dual Fuel Stratification (DDFS) strategy is a new pathway that benefits both RCCI and PPC combustion simultaneously. DDFS has improved control and more authority over the rate of heat release, comparable thermal efficiency to RCCI, less cyclic variation, better potential for downstream recovery when compared to RCCI and etc. One of the primary drawbacks of DDFS is high soot production 12

alternatives to gasoline, were studied. In the E85/D DDFS case, NOX increased dramatically by a factor of 25 times. PPRR also increased dramatically by about 2.7 times compared to the G/D DDFS case. Owing to these main drawbacks, E85 was not a good candidate as an alternative to gasoline in DDFS combustion. On the other hand, in the E10/D DDFS case, soot (the primary drawback of DDFS) decreased by about 40%, while NOX increased only less than 2%. In general, E10 was a suitable alternative to gasoline for the DDFS strategy because it not only improved soot, UHC, CO, and GTE, but also a negligible change in fuel consumption, NOX, and PPRR was observed. In order to investigate the role of diesel energy fraction and its start of injection on fuel stratification, diesel energy fraction was swept from 0% to 11%, and also SOI2 was swept from −100° ATDC to −30° ATDC. It was found that 11% diesel energy fraction and above this magnitude was not suitable at this engine load because not only the peak of low temperature heat release was rather high which was a sign of diffusion-limited regime, but also the peak of AHRR occurred before TDC, which resulted in negative work. The 0% diesel energy fraction was not suitable as well because the strategy is no longer DDFS and changes into PPC. By sweeping SOI2 and diesel energy fraction (Ed), three regimes were discovered: PPC-like (partially premixed), reactivity-controlled, and diffusion-limited regimes. The PPC-like regime was not suitable owing to high UHC, CO, and soot emissions. The diffusionlimited regime was not acceptable as well due to the high levels of NOX. The reactivity-controlled regime (SOI2s between −80 to −40° ATDC) was the best regime regarding the best results of emissions, thermal efficiency, and fuel consumption. In this region, if Ed was between 4 and 6%, the E10/D DDFS combustion can meet the EURO6 emission mandate without aftertreatments. The effects of SOI and energy fractions of the near-TDC injection of the E10/D DDFS combustion were investigated. SOI3 was swept from −8 to 4° ATDC, and EE10 was swept from 20 to 40% of the total cylinder energy. It was found that the best region from emission, performance, and fuel consumption perspectives was SOI3 between −4 to −2° ATDC and EE10 between 20 and 30%. In this region, the combustion can meet EURO6 without aftertreatments. As the spray angles have significant effects on emission formation, the spray angles of both injectors were swept from 50 to 80°. It was found that a medium spray angle about θd = 55° for the diesel injector had the best results for emissions. This magnitude also yielded the best results for gross thermal efficiency and fuel consumption. For the E10 injector (near-TDC injection), θg between 60 and 70° showed the best results regarding emission and performance. Another promising way to decrease soot is an increase in near-TDC injection pressure. The peak of AHRR and cylinder pressure increased by boosting the injection pressure. In addition, CA50 and the peak of AHRR advanced. An increase in the injection pressure resulted in high NOX for the E10/D DDFS case (with ED = 7%). Above 1250 bar injection pressures, NOX does not meet the EURO6 standard. However, soot and CO emissions decreased slightly for the injection pressures above 1000 bar. From the emissions perspective, from 750 bar to 1250, bar injection pressure has acceptable results. Gross thermal efficiency and gross IMEP reduced significantly, for low injection pressures (near 500 bar).

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Fig. 24. Emissions and performance of E10/D DDFS with a sweep of injection pressure for the near-TDC injection.

500 bar

1000 bar (base case)

1500 bar

Fig. 25. Iso-thermal surfaces (2000 and 2200 K) for E10/D DDFS with a sweep of injection pressure for the near-TDC injection.

For future works, it is proposed that E10/D DDFS can be optimized for this load and other loads. In addition, the effects of other alcohol fuels can be investigated as well. DDFS is a novel strategy with lots of supremacies over RCCI, and due to its near-TDC injection, new piston profile like those conventional ones can be proposed for DDFS and investigated. DDFS has a robust control over the rate of heat release than other LTC strategies, and also it is less sensitive to boundary conditions; thus, it is a suitable strategy to be used in automotive industries, especially for hybrid cars, which demand high thermal efficiency and low fuel consumption.

Investigation, Validation, Visualization, Writing - original draft. Sasan Shirvani: Conceptualization, Methodology, Software, Investigation, Validation, Visualization, Writing - original draft. Amir H. Shamekhi: Supervision, Resources. Rolf D. Reitz: Writing - review & editing. Declaration of Competing Interest The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.

CRediT authorship contribution statement Saeid

Shirvani:

Conceptualization,

Methodology,

Software,

Appendix In order to show that the developed 3D-CFD model is reliable with 1.4 mm mesh size, the CFD model was used for other different loads using gasoline/diesel 6.5 and 9.25 bar (50% EGR) under the RCCI strategy, shown in Fig. A.1. As Fig. A.2 shows, E85/diesel numerical model was used for different loads, including 9.6 and 11.6. According the results in this figure, the model is reliable for E85/D combustion. 13

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Exp. NOx

Num. NOx

Exp. soot

Num. soot

Exp. UHC+CO

Num. UHC+CO

0.05

20

0.04

15

0.03

10

0.02

5

0.01 0

G/D RCCI 6.5 bar and 9.25 bar gross IMEP (50% EGR) at 1300 rpm (Exp. data was taken from [29, 68]).

6.5 bar

0

9.0 bar

Comparison of emissions in gr/kW-hr for experimental data and numerical results (Exp. data was taken from and [29, 68])

Fig. A1. Comparison of cylinder pressure, AHRR, and emissions between the numerical results and experimental data for gasoline/diesel RCCI combustion.

Exp. NOx

Num. soot ×10

Exp. UHC+CO

Num. UHC+CO

0.15

15

0.1

10

0.05

5

0

E85/D RCCI 9.6 bar and 11.6 bar, 1300 rpm and no EGR (Exp. data was taken from [4]).

Num. NOx

Exp. soot×10

9.6 bar

11.6 bar

0

Comparison of emissions in gr/kW-hr for experimental data and numerical results (Exp. data was taken from [4]).

Fig. A2. Comparison of cylinder pressure, AHRR, and emissions between the numerical results and experimental data for E85/diesel RCCI combustion.

All operating parameters to calculate engine performance are as follows. Gross indicated work and gross indicated power per cycle are calculated based on Eqs. (A.1) and (A.2), respectively.

Work gross = Pgross =

180 180

pdV

(A.1)

Work gross × nR (A.2)

Vd N

where; p is pressure and V is volume, P is power, Vd is displacement volume and nR for the four-stroke engine equals 2, and N is engine speed. Gross indicated mean effective Pressure (IMEPgross ) is obtained from Eq. (A.3) [63]:

IMEPgross =

Pgross nR (A.3)

Vd N

Fuel energy is obtained from Eq. (A.4), and gross thermal efficiency (GTE) is as follows:

Efuel = mgasoline × LHVgasoline + methanol × LHVeth + mdiesel LHVdiesel gross

= GTE =

(A.4)

Workgross (A.5)

Efuel

where; Efuel is the fuel energy, m is the mass for each species, and LHV is the lower heating value. The adiabatic apparent heat release rate is calculated based on Eq. (A.6) [30]:

dQad = d

dV p + 1 d

1 1

V

dp d

(A.6) 14

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where;

is the crank angle degree, γ is the ratio of specific heats. PPRR is a noise metric, and it can be defined as follows:

PPRR =

dp d

(A.7)

max

In this paper, ethanol is considered fully evaporated and homogenous at IVC. Ethanol has a considerable enthalpy of vaporization and it cools the combustion chamber. The CFD simulation is performed based on closed-cycle and it is assumed that ethanol is fully evaporated at IVC. The first injection in this model starts at 340° BTDC and occurs before the start of simulation 143° BTDC (IVC). Since injection pressure is high (about 1000 bar), ethanol has enough time to become fully vaporized, and the combustion chamber is hot enough in the real engine to make ethanol fully vaporized. Moreover, the injector is a wide spray angle with ten nozzles, which facilitates the vaporization process. Thus, the amount of fuel at first injection is considered fully homogenous for the start of the simulation. The cooling effects of ethanol can be calculated based on Eqs. (A.8) and (A.9), and Table A.1 gives the physical properties of ethanol and iso-octane. Table A2 shows the cooling effects of ethanol (E10 and E85) on the combustion chamber.

Q1 = mEth × cp, Eth × (Tboiling

Tinitial ) + mEth × hfg , Eth + mGas × cp, Gas × (Tboiling

(A.8)

Tinitial ) + mGas × hfg, Gas

(A.9)

Q1 = Q2 = mair × cp, air × T where, m is mass for each species, cp is the specific heat capacity. Table A1 Physical properties of iso-octane and ethanol (data taken from [67]).

Lower Heating Value (MJ/kg) Research Octane Number (RON) Motored Octane Number (MON) Liquid Density @ 25° C (kg/m3) Enthalpy of Vaporization (kJ/kg) Initial Boiling Point (°C) Final Boiling Point

Iso-octane

Ethanol

44.3 100 100 692 272 99 99

26.9 107 89 785 840 78 78

Table A2 Temperate variation of the ethanol vaporization effect. Ethanol fraction by volume (%)

T (°K)

E0 (pure gasoline) E10 E85

6.5 7.3 16.3

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