Thermodynamic energy and exergy analysis of low-temperature combustion strategies [14]

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Thermodynamic energy and exergy analysis of low-temperature combustion strategies [14]

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ARTICLE INFO Article ID: 03-14-03-0021 © 2021 SAE International doi:10.4271/03-14-03-0021

Thermodynamic Energy and Exergy Analysis of Low-Temperature Combustion Strategies Saeid Shirvani,1 Sasan Shirvani,1 Rolf Reitz, 2 and Fatemeh Salehi3 KN Toosi University of Technology, Islamic Republic of Iran University of Wisconsin-Madison, USA 3 Macquarie University, Australia

1

2

Abstract Increasing thermal and fuel efficiency in Internal Combustion Engines (ICEs) requires thorough investigations on the combustion process and its thermodynamics. The first law of thermodynamics expresses the balance of the energy, while the second law specifies the maximum achievable work. In this article, Low-Temperature Combustion (LTC) strategies, including Homogeneous Charge Compression Ignition (HCCI), Reactivity Controlled Compression Ignition (RCCI), Partially Premixed Combustion (PPC), and Direct Dual-Fuel Stratification (DDFS) were analyzed by the first and second law approaches, and they were compared with ideal-diesel cycle and Conventional Diesel Combustion (CDC). HCCI and RCCI had the highest exergy efficiency of 50.8% and 49.2%, respectively, compared to other cases, and exergy destruction in these cases was the lowest (25.3% and 27.5%, respectively). Although all mentioned LTC strategies met the EURO6 regulation for NOX and soot, the thermodynamic analysis confirmed that DDFS was superior in terms of emissions and engine performance. The DDFS strategy uses two direct injectors in the combustion chamber, and the injection parameters have influential effects on the energy and exergy distributions as well as emissions. These injection parameters in a DDFS engine with direct diesel and gasoline injectors are energy fractions for each fuel (Ed and Eg) and their injection timing (SOI2 and SOI3), spray angles (θd and θg), the injection pressures (Pd and Pg). Furthermore, the Exhaust Gas Recirculation (EGR) was also studied as another effective parameter. The suitable ranges for desired exergy efficiency and exergy destruction levels for the mentioned parameters are as follows: Ed and Eg should be in the range of 4-6% and 25-30% of the total energy input, respectively. −100° to −80° After Top Dead Center (ATDC) is suitable for the diesel injection timing (SOI2), and −8° to −6° is a preferred range for the gasoline injection timing (SOI3). The suitable spray angle for both injectors was found to be about 65°, and the injection pressure about 1500-2000 bar showed better results for both injectors.

History Received: Revised: Accepted: e-Available:

27 Oct 2020 09 Dec 2020 04 Jan 2021 03 Feb 2021

Keywords Energy analysis, Exergy analysis, DDFS, LTC, CFD simulation, ICCI

Citation Shirvani, S., Shirvani, S., Reitz, R., and Salehi, F., “Thermodynamic Energy and Exergy Analysis of Low-Temperature Combustion Strategies,” SAE Int. J. Engines 14(3):345-367, 2021, doi:10.4271/03-14-03-0021. ISSN: 1946-3936 e-ISSN: 1946-3944

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Introduction

I

nternal Combustion Engines (ICEs), fueled by fossil and bio-derived fuels, play a dominant role in supplying the mobility and power generation in the world. Twenty-five percent of the world’s energy demand belongs to the ICEs, and they contribute 10% of the world’s greenhouse gases. Other alternatives to ICEs are Electric Vehicles (EVs), but the application of EVs is still challenging. The replacement of the ICEs with EVs may need massive infrastructure and tremendous expense, and it will take decades for the replacement. Another challenge of EVs is the high cost and weight of batteries. Scholars hope to charge batteries by renewable sources of energy such as solar and wind, but it must be mentioned that these energy sources provide a minuscule fraction of the world’s energy supply [1]. It may conclude that there is still no reliable alternative to ICEs over their wide range of applications; thus, it is essential to continue improving and developing new ICE technologies to reach highly efficient engines with extremely low levels of pollutants. ICEs are categorized into spark and Compression Ignition (CI) engines. Most of the typical CI engines are operated on conventional diesel fuels that use a direct injector to inject diesel fuels into the combustion chamber. Conventional Diesel Combustion (CDC) operates on a diffusion-limited or mixingcontrolled mode. The rate of heat release and timing is controlled by the injection timing that provides sufficient controllability over the rate of heat release. However, the nature of the diff usion-limited combustion results in high levels of nitrogen oxides (NOX) and Particulate Matter (PM) due to locally fuel-rich zones at the flame front. Emission standards such as the Environmental Protection Agency (EPA) of the United States and European regulations mandate automotive manufactures to use aftertreatment to deal with NOX, PM, and other hazardous pollutants such as carbon monoxide (CO) and Unburned HydroCarbon (UHC). With considerable advancements in engines, hazardous pollutants have been reduced substantially in the past decade, although they have not met the requirements for efficient, extremely low-emission engines. Beatrice et al. [2, 3] performed experimental investigations on the effects of the injection strategy and intake/exhaust conditions in a high-speed light-duty engine to exceed the 100 kW/l (power density) as a milestone of the operation of modern diesel engines. They reported the following results to achieve the target: the boosted intake pressure should be  more than 3.5 bar, and the difference between exhaust and intake pressure should be 0.7-1 bar. The injection pressure needed to be more than 2700 bar, and the fuel flow rate was considered to be about 960 cm3/min. The inlet turbine temperature was reported to be more than 870°C. Small use of the Exhaust Gas Recirculation (EGR) rate could improve the trade-off between NOX, CO2, and power density. For the engine speeds over 4000 revolutions per minute (rpm), the engine performance decreased owing to the considerable pumping and friction losses. Low-Temperature Combustion (LTC) is a practical pathway to reach ultralow emissions combined with higher

engine performance and efficiency. LTC is a potential solution to reduce the expense of aftertreatment and decrease NOX and PM, but at the expense of an increase in UHC and CO amounts when compared with CDC. The most well-known LTC strategies can be categorized as Homogeneous Charge Compression Ignition (HCCI) in the 1980s; Premixed Charged Compression Ignition (PCCI), 1996; Reactivity Controlled Compression Ignition (RCCI), 2006; and Direct Dual-Fuel Stratification (DDFS), 2015. HCCI was introduced by Onishi et al. [4] and operated in a two-stroke sparked engine. As the name implies, a fully homogeneous charge enters the combustion chamber and the combustion is kinetically controlled. Heat addition occurs in a short duration and at an almost constant volume that makes HCCI combustion efficient like the ideal Otto cycle. The short combustion duration of HCCI results in high engine noise and limited operating range. HCCI combustion phasing can be controlled with the EGR rate, intake temperature, and pressure, and equivalence ratio. As a result, HCCI is not a practical strategy for the mobile application. From the emissions standpoint, HCCI combustion emits ultralow NOX and soot; however, UHC and CO are rather high. Many studies are ongoing to provide more control over HCCI combustion and develop methods to tackle the HCCI limits. Shiraishi [5] improved the HCCI combustion ignitability by implementing a low-temperature plasma igniter with 15 kHz frequency voltage at the top center of the chamber of a gasoline HCCI engine with a 15:1 compression ratio. The igniter discharge timing was about −315° After Top Dead Center (ATDC) in the intake stroke and showed the optimum results in terms of promoting HCCI autoignition. The igniter produced active ozone species in the combustion chamber, and the active radical of O initiated the combustion process of the HCCI engine. The plasma-assisted case showed better results than conventional spark-assisted cases when it comes to the HCCI strategy. Gawale et al. [6] conducted an experimental investigation on the effects of propanol/diesel blends on combustion and emissions in an HCCI engine. The experiments were combined with an optimization study to indicate the suitable mass flow rate of alternative fuels to diesel. By applying propanol, NOX and soot decreased compared to the baseline diesel engines, while UHC and CO increased. At higher load operating conditions, P100+Diesel HCCI showed good results in brake thermal efficiency compared to the conventional diesel engine. At 100% load, NOX reduced up to 50% compared to the low load, and soot decreased up to eight times in magnitude. PCCI was termed by Aoyama et al. [7] in 1996, which is a strategy that supplies a premixed charge by port injection while a portion of the fuel is directly injected in the combustion chamber at an early timing. This fuel stratification in the combustion chamber extends the combustion duration and provides better control over the rate of heat release compared to HCCI. PCCI is still a challenging strategy due to its cyclic variation, sensitivity to the initial conditions, and limited operating range. Kalghati et al. [8] showed that early pilot injection with the main injection near the Top Dead Center

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(TDC) can improve combustion stability and cyclic variation. They also reported that PM was reduced by utilizing gasoline instead of diesel fuel. Th is strategy is also called Partially Premixed Combustion (PPC). At higher operating loads, the need for EGR and high-octane fuels is increased to maintain combustion phasing. Other drawbacks of PCCI are engine noise and high levels of UHC and CO emissions [9, 10, 11]. Xu et al. [12] carried out an experimental investigation on the effects of the HCCI to PPC transition on emissions and engine performance in a heavy-duty engine fueled by 81% iso-octane by volume and 19% n-heptane. For the HCCI mode, Start Of Injection (SOI) timing was varied between −100° and −48° ATDC, and for the PPC regime, the timing was varied from −44° to −20° ATDC. For the HCCI regime, the mean in-cylinder temperature was lower than 1700 K; while in the PPC, it was lower than 1850 K. For both regimes, the main source of UHC was in the crevice volume, and the source of CO emissions was the fuel-lean region in the cylinder. Thermodynamic efficiency of the PPC case was reported higher than HCCI for SOIs between −44° and −31° ATDC. Belgiorno et al. [13] conducted an experimental study to investigate the effects of EGR, pilot quantity, and combustion phasing in a light-duty gasoline engine operated with the PPC strategy. The study’s main purpose was to minimize emissions and fuel consumption and obtain the engine’s calibration parameters. By setting the EGR rate to about 45%, advancing the combustion phasing up to 3-4° ATDC, and lowering the pilot quantity ratio by about 31%-33%, the engine managed to achieve 2% higher thermal efficiency, and a considerable reduction by a factor of 2 in soot, 0.5 g/kW-h in NOX and 5% in CO2 were obtained when compared with the CDC strategy. In another attempt, they used two different combinations of diesel, gasoline, and ethanol in the ratio of 68:17:15 and 58:12:30 in the light-duty engine. They reported that thermal efficiency improved by about 4% compared to the CDC case at the load of 8 bar BMEP; however, NOX slightly increased at higher loads (13 bar BMEP) by about 1 g/kW-h when going from high diesel portion to high ethanol portion of the mixtures. Soot was oxidized by utilizing the oxygenated fuels up to five times [14]. In 2020, they studied the effects of piston bowl features by inspiring the General Motors Company (GM)’s most innovative achievements to yield better results for PM production. By adopting a new piston profi le, which was a combination of a highly reentrant sharp-stepped bowl with inner radial lips, PM reduced by 25%-50% in magnitude at different operating loads [15]. RCCI was introduced by Inagaki et al. [16] in 2006. They utilized two different fuels with different reactivities to extend the operating range. Iso-octane was used as the low-reactivity fuel injected by a port-fuel injector to provide a partially homogeneous charge in the combustion chamber, while n-heptane was injected into the combustion chamber as the high-reactivity fuel to provide reactivity stratification. Combustion starts from the high-reactivity regions (richer n-heptane zones) and continues toward low-reactivity zones. Th is concept uses both fuel and reactivity stratification to extend the combustion duration more than HCCI and PCCI.

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RCCI also leads to better control over the rate of heat release; however, it is influenced by initial temperature and pressure and, hence, cannot operate at higher loads. Similar to HCCI and PCCI, RCCI emits ultra-low NOX and while it provides high thermal efficiency [17, 18]. As it is reported by Splitter et al. [19], RCCI can reach up to 60% thermal efficiency (from the perspective of the first law) under optimized conditions. Jonathan et al. [20] carried out an experimental study on the intermediate modes between the RCCI and CDC strategies in a light-duty 1.9L GM engine operated with a blend of propane and dimethyl ether as the low-reactivity fuel and diesel as the high-reactivity fuel. The Premixed Dual-Fuel Combustion (PDFC) mode was discovered as a bridge between RCCI and CDC at mid-loads; and by using PDFC, some drawbacks of RCCI such as high levels of Peak Pressure Rise Rate (PPRR) and UHC and CO can be improved. They reported that RCCI and to a lesser degree of PDFC showed similar performance regarding their versatility over different operating rages. Pandian and Krishnasamy [21] performed an experimental investigation on the Homogeneous Charge Reactivity Controlled Compression Ignition (HCRCCI) strategy. They reported that this strategy could reduce UHC and CO emissions due to the incomplete combustion in the RCCI strategy. Under optimized conditions, the brake thermal efficiency of the HCRCCI increased, and UHC and CO reduced by about 37% and 87%, respectively. In addition, NOX and soot in the HCRCCI case were at almost zero levels. The main drawback of the LTC strategies is low control over the rate of heat release because the combustion is kinetically controlled, and it is sensitive to small changes in initial conditions (e.g., intake temperature or pressure). The DDFS is a promising LTC strategy that provides better control over the rate of heat release using the near-TDC injection. The DDFS strategy was introduced by Wissink and Reitz in 2015 [22], and they managed to improve combustion controllability by implementing two direct injectors in the combustion chamber for gasoline and diesel injections. The gasoline injector provides a partially premixed charge by an early injection (at about −340° ATDC). The diesel injector injects the high-reactivity fuel to provide a reactivity stratification that benefits from the RCCI concept. In the third injection, the gasoline injector directly controls the rate of heat release by the near-TDC injection. By adopting this method, NOX and soot emissions are at acceptable levels (meets the EPA2010 and EURO6 mandates), and UHC and CO are decreased due to the nature of the diff usion-limited combustion [23, 24, 25]. Compared to RCCI, the primary drawback of DDFS is the higher levels of PM due to the diff usion-limited nature of the near-TDC injection. One practical pathway to address this problem is using oxygenated fuels like ethanol blends with gasoline. Shirvani et al. [26] performed a numerical study showing that by using E10 (10% ethanol blended in gasoline by volume) as the low-reactivity fuel, it is possible to decrease PM up to 40%; however, NOX increased by 2%. Ethanol improved fuel oxidization, so Gross Thermal Efficiency (GTE) slightly increased in the E10 case compared to the gasoline/ diesel DDFS. In another attempt, they showed that by using

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conventional diesel-like piston profiles (omega type), NOX and CO reduced by 16% and 27%; respectively, compared to the modified-piston [27]. DDFS’s thermal efficiency and CO2 emissions are comparable to the RCCI strategy while it provides direct control over the rate of heat release [28]. Li et al. [29] performed experimental investigations on the DDFS mode, and they denoted it as Intelligent Charge Compression Ignition (ICCI). The ICCI engine fueled by gasoline and diesel showed thermal efficiency near 50%, and NOX emissions were at 0.12 g/kWh. By increasing the gasoline ratio, combustion phasing retarded and NOX decreased, but UHC and CO increased. At the load of 8 bar IMEP, the best results for thermal efficiency were achieved and the NOX level was measured below 0.1 g/kW-h. In another attempt [30], they demonstrated that the ICCI engine fueled by methanol and biodiesel could achieve up to 53% thermal efficiency and acceptable levels of UHC and CO less than 0.6 g/kWh. The optimal methanol injection timings were obtained at −340° to −300° ATDC, and biodiesel injection was at −48° to −42° ATDC. Energy and exergy analyses have been previously taken into consideration to not only quantify the energy portion from the perspective of the fi rst law but also to study the potential of waste heat recovery from the second law’s perspective. Mahabadipour et al. [31] developed a numerical model for exergy analysis in a methane/diesel dual-fuel engine. Exergy destruction was computed for several diesel injection timings at fi xed 6.0 bar IMEP. They reported that about 99% of the total exergy input is chemical fuel exergy of methane and diesel fuels; however, only 1% of exergy input consists of physical exergy of the input flow. It was also reported that the combustion process significantly contributed to the exergy destruction, which is inevitable. They also mentioned that SOI timing cannot mitigate exergy destruction by itself. Li et al. [32] conducted thermodynamic analyses from the first and second law perspectives for three combustion regimes: HCCI, RCCI, and CDC. The results showed that HCCI had the highest exergy and energy efficiency, while CDC yielded the worst. The CDC case had more potential for heat recovery due to the higher exhaust pressure and temperature, and it also demonstrated the highest levels of exergy destruction, while HCCI had the lowest level. The position of CA90 affected thermal efficiency and contributed to reducing exhaust loss. In the RCCI strategy, utilizing the heat transfer and exhaust loss was less affected by the CA50 compared to the CDC and HCCI. In another study [33], they demonstrated the chemical reactions and combustion process as the most important sources of the exergy destruction in engines using the CDC strategy. However, for RCCI and HCCI engines, the main source of the exergy destruction was reported in the transition between the low-temperature heat release and the hightemperature heat release. Making the combustion process more homogeneous could decrease the exergy destruction, and consequently, the HCCI and RCCI showed the lowest levels of exergy destruction compared to the CDC case. In this study, a three-dimensional Computational Fluid Dynamics (3D-CFD) model was developed and validated against experimental data for different modes, including

kinetically controlled combustion (RCCI and HCCI) as well as kinetically controlled combustion combined with diffusionlimited combustion (DDFS). The main objective of this study is to provide a comparative study on different LTC strategies (HCCI, RCCI, PPC, DDFS), CDC, and ideal-diesel cycle from the perspectives of the first and second laws of thermodynamics. Exergy destruction should be minimized from the thermodynamic point of view. To the best of our knowledge, there are few studies available that compare LTC strategies from the perspective of the first and second laws, and there is no available energy and exergy analysis on the DDFS strategy that investigates the effects of operating parameters (injection parameters and EGR) on the exergy efficiency and exergy destruction.

Material and Methods In this study, a 3D-CFD model was developed and validated against experimental data for different modes: RCCI, DDFS, and CDC. Experimental data were taken from [22, 23]. A mesh sensitivity study was performed to achieve a suitable mesh refi nement, considering both accuracy and computational time. The results of the mesh study are provided in the Appendix. Table 1 gives the engine and injector’s specification of a single-cylinder research engine, whereas Table 2 shows the operating parameters of the DDFS mode with a 4.71 kJ cylinder energy input.

TABLE 1   Geometrical and technical specification of the engine and gasoline and diesel injectors used for the DDFS strategy (experimental data taken from [23]).

Piston type

Modified piston (bathtub)

Displacement (L)

2.44

Bore (mm)

137.2

Stroke (mm)

165.1

Connecting rod length (mm)

261.6

Squish height (mm)

1.57

Number of valves per cylinder

4

IVO (° ATDC)

335

IVC (° ATDC)

−143

EVO (° ATDC)

130

EVC (° ATDC)

−355

Swirl ratio

0.7

Compression ratio

14.88:1

Diesel injector Nozzle angle (°)

148

Hole diameter (μm)

141

Number of holes

7

Gasoline injector Nozzle angle (°)

143

Hole diameter (μm)

117

Number of holes

10

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TABLE 2   Experimental operating conditions of the engine

using the DDFS strategy (experimental data taken from [23]). EGR (%)

39.5

TEGR (°C)

90.3

Tin (°C)

49.3

Pin (kPa) (supercharged)

186.2

Equivalence ratio (—)

0.57

Gross IMEP (bar)

9.41

Qfuel (kJ/cyc)

4.71

Fuel1

Diesel

Injector name

CRI1

Injection pressure (bar)

500

SOI2 (° BTDC)

60

Dur2 (ms)

0.6

Total energy ratio (%)

7.0

Fuel2

Gasoline

Injector name

CRI2

Injection pressure (bar)

1000

SOI1 (° BTDC)

340

Dur1 (ms)

1.4

SOI3 (° BTDC)

4

Dur3 (ms)

0.8

Computational Model The 3D-CFD model was used to simulate HCCI, RCCI, PPC, DDFS, and CDC combustion for a closed-cycle simulation. The simulation started from −180° to 180° ATDC for all cases. In the experiments, gasoline was injected at −340° ATDC. For the HCCI, RCCI, PPC, and DDFS simulations, gasoline was considered homogeneous at the start of the simulation.

The CFD model uses a hybrid method that solves the Eulerian Reynolds-averaged Navier-Stokes equations for the gaseous phase while adopts a Lagrangian approach to simulate the dynamics of the injected droplets until their vaporization. The Kelvin-Helmholtz and Rayleigh-Taylor (KH-RT) hybrid models were used to simulate primary and secondary breakups for spray modeling in all cases [34]. The drop drag submodel is combined with a drop drag coefficient that changes dynamically with the gaseous field updated in the CFD model [35]. The Standard Discrete Droplet Model (DDM) was employed, which uses a technique for simulating the behavior of atomized evaporating liquid parcels in the gaseous environment [36]. The No-Time-Counter (NTC) method, which is faster than the O’Rourke model, was employed to calculate droplet collisions in Lagrangian spray simulation [37]. To simulate turbulence transport in the combustion simulation, the Re-Normalization Group (RNG) k-ε model was utilized. This model showed promising results when it comes to ICEs, especially with combustion resulting from spray injection [38]. Detailed chemical kinetics using a multi-zone model was applied for combustion modeling [22, 39] that accelerates the simulation while it is suitable for multicomponent fuels like gasoline and diesel [40]. A reduced chemistry mechanism for iso-octane and n-heptane combustion with 108 species and 435 reactions was adopted. The Hiroyasu soot model and extended Zeldovich were employed for the soot and NOX predictions [41, 42]. Fixed wall boundary conditions with no roughness were assumed for the boundary conditions of the head cylinder, cylinder wall, and piston surface. Wall temperature values were chosen based on the experimental tests provided in Ref. [43] for the RCCI and CDC combustion. Figure 1 shows the locations of gasoline and diesel injectors in the combustion chamber. Mesh refi nement technique was used for critical regions near injector, piston surface, head, and wall cylinder

  FIGURE 1    Combustion chamber, piston profile geometry, and gasoline and diesel injector positions.

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Experimental operating parameters

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  FIGURE 2    Validation of the cylinder pressure and AHRR for the numerical predictions and experimental data.

zones. Senecal et al. [44] recommended that the optimum grid size for the spray simulation should not exceed 0.25 mm. Thus the mentioned mesh size was chosen as a criterion for the mesh refinements in the critical zones. The mesh sensitivity study confirmed that the mesh resolution of 1.4 mm in other regions is suitable, which is adopted in all cases.

Model Validation The 3D-CFD model was validated against experimental data for different modes. Figure 2 illustrates the cylinder pressure and Apparent Heat Release Rate (AHRR) traces, whereas Figure 3 shows the emission results for the DDFS combustion case. The case with a 4.71 kJ cylinder energy input was

considered as the base case, and the validation was performed for different modes. The predicted results for the DDFS case are provided here, and other results are available in the Appendix. It can be  seen that the predicted pressure and AHRR agree well with the measurements [23].

The First and Second Law of Thermodynamics Methodology in CI Engines In this section, the concepts and equations used for the first and second law of thermodynamics analyses are provided and discussed.

  FIGURE 3    Emission results for numerical predictions and experimental data.

0.6

0.47

0.5 0.4 0.3

Num. results

Exp. data 0.5

6

5.06

5

4.9

4

0.291 0.28

3

0.2

2

0.1

1

0

0

NOx (g/kW-h)

Num. results

Soot×100 (g/kW-h)

2.15

1.66

UHC (g/kW-h)

CO (g/kW-h)

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Exp. data

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  FIGURE 4    The first law approach to the CI engines.

QH

101

Pressure (MPa)

Combustion loss

100

Heat transfer loss

Input energy

W

Exhaust loss

QL

Gross work

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10–1 0

0.5

1

1.5 Volume (m3)

2

2.5

3 10–3

(b)

(a)

The First Law of Thermodynamics Approach for CI Engines Figure 4 schematically illustrates the concept of the first law in the CI engines. The energy input that equals mf × LHVf is supplied from the injected fuel, as shown in Figure 4(a). The energy portion of the cylinder energy input exits as the heat loss, which is denoted by QL Losses, categorized as heat transfer, exhaust, and combustion losses. Figure 4(b) is the energy Sankey diagram, which is a description of the energy conservation or the first law. Gross indicated work and gross indicated power per cycle are predicted based on Work gross =

180

∫ pdV

Eq. (1)

where Efuel is the fuel energy, m is the species mass, and LHV is the lower heating value. The thermal efficiency of the first law is finally defined as η1st = GTE =

Work gross E fuel

Eq. (5)

Exhaust loss (EXL) is the ratio of the difference between output and input enthalpies to Efuel, and it is expressed as Exhaust loss =



m i Hi − ∑ Intakem i Hi

Exhaust

E fuel

Eq. (6)

where mi and Hi are the mass and enthalpy of each species. Combustion efficiency (ηcomb) is the ratio of the lower heating value of burned species to the total energy input, and it is expressed as

−180

Pgross =

Work gross × n R Vd N

Eq. (2)

where p is pressure, V is volume, P is power, Vd is displacement volume, N is engine speed, and nR for the four-stroke engine is 2. Gross Indicated Mean Effective Pressure (IMEPgross) is calculated using IMEPgross =

Pgross n R Vd N

Eq. (3)

The total cylinder energy input is then computed as E fuel = m gasoline × LHVgasolnie + m diesel × LHVdiesel

Eq. (4)

ηcomb =

∑ m i × LHVi − m CO × LHVCO − m UHC × LHVUHC E fuel Eq. (7)

Combustion loss (CL) is due to the unburned species such as CO and UHC, and it is defined as Combustion loss ( CL ) = 1 − ηcomb

Eq. (8)

The heat transfer loss is the ratio of heat transfer, which passes through the cylinder walls and piston surface, to the cylinder energy input (Efuel). The first law of thermodynamics expresses the conservation of energy, and hence GTE + CL + HTL + EXL = 1

Eq. (9)

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The adiabatic AHRR is calculated as dQad dp γ dV 1 = p + V γ − 1 dθ γ − 1 dθ dθ

Eq. (10)

where θ is the crank angle degree and γ is the ratio of specific heats. PPRR is a noise metric, and it is defined based on  dp  PPRR =    dθ max

Eq. (11)

The Second Law of Thermodynamics Approach for CI Engines The first law of thermodynamics is useful to study efficiencies, but it does not manifest thermodynamic irreversibility, which means lost work. Therefore, the second law is used to quantify the thermodynamic irreversibility. Exergy is the maximum theoretical work that can be achieved by a system. The second law of the thermodynamics approach evaluates the quality of energy during the combustion process. Thus the second law provides a thorough insight into the thermodynamic process of combustion. The second law can identify energy sources and wasted energy in a system. The availability of a system means the maximum power that can be extracted from a system under reversible processes when the system reaches thermal, mechanical, and chemical equilibrium with its surroundings. In short, the second law specifies what fraction of the first law is achievable and what fraction will be destroyed inevitably. Figure 5(a) is an illustration of the second law and work availability for the idea-diesel cycle. Equilibrium is a state in which the system comes to thermal, mechanical, and chemical equilibrium to the environment. Thermal equilibrium means there is no difference

in temperature compared to the surroundings, so heat transfer does not occur. Mechanical equilibrium is a state in which there is no pressure gradient between the system and surroundings; thus no output work is attainable in this state. Chemical equilibrium is a state when no chemical reactions or mass diffusion will occur between the system and surroundings. In this study, the Ultimate Dead State (UDS) is defined, where P0 = 101 kPa, T0 = 298 K, and mole fractions of O2 , N2 , H 2O, and CO2 are 0.2035, 0.7567, 0.0303, and 0.0003, respectively. Exergy input can be divided into two components—the physical and chemical exergy: EX = EX physical + EX chemical

Eq. (12)

Physical exergy is the maximum work that can be attained from a system when it reaches thermal and mechanical equilibrium or, in other words, the temperature and pressure of the system reach its surroundings or the UDS. Chemical exergy means maximum work in which the system completes chemical equilibrium with the environment, or the Gibbs free energy of a system equals that of the environment or UDS. Figure 5(b) shows the Sankey diagram for the exergy flow in the cylinder. Fuel exergy is obtained as follows, and the thermomechanical exergy of the injection is about 0.2% of the LHV based on Ref. [45] and can be considered negligible.    EX fuel =  ∑ Hi − ∑ Hi − T0  ∑ Si − ∑ Si   Products Products    Reactants  Reactants +0.002LHV

Eq. (13)

where H and S are the enthalpy and entropy of each species. T0 is the temperature at UDS. Exhaust exergy comprises of

  FIGURE 5    The second law approach to the CI engines.

2200 2000

Energy destruction

1800

QH

1400 1200

Combustion loss Heat transfer loss Exhaust loss

Chemical energy

W

1000 800

QL

600

200 0.0405

Physical energy

Available energy

400

Gross work

Unavailable energy 0.041

0.0415

0.042 0.0425 Entropy (kJ/K)

(a)

0.043

0.0435

0.044

(b)

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Temperature (K)

1600

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thermomechanical and chemical terms and calculated as follows: at ) exhaust ( at ) exhauust EX exhaust = EX (thermomechanical + EX chemical

Eq. (14)

 k k ∑ i =1Hi ( Te ) − ∑ i =1Hi ( T0 )   k  P   k − T0  ∑ i =1Si ( Te ) − ∑ i =1Si ( T0 ) − R × ln  e     P0     at ) exhaust EX (thermomechanical = EX fuel

Eq. (15)

( at ) ehxuast

EX chemical

 Xe  k R × T0 × ∑ i =1 ln  0i   Xi  = EX fuel

EX incomplete

Eq. (16)

( at ) exhaust

EX heat

© SAE International

  FIGURE 6    Problem formulation flowchart.

∫ ( p − p0 ) dV EX fuel

Eq. (19)

where p is cylinder pressure and dV is the differential change in the cylinder volume. According to the second law of thermodynamics and exergy balance in the system, exergy destruction is calculated based on Eq. (20) Entropy can be obtained using the ideal-gas law: T P S = S0 + m × C P × ln   − m × R × ln   Eq. (21)  T0   P0  where S0, T0, and P0 are entropy, temperature, and pressure at the UDS. m is the mass of the cylinder mixture, T is the mean temperature, R is the gas constant, and P is cylinder pressure.

Eq. (17)

Results and Discussion

The exergy of heat transfer is also defined as  T  ∫ 1 − T0g dQ  =  EX fuel

EX work = η2 nd =

EX destruction = 1 − EX work − EX heat − ( EX exhaust − Ex intake ) − EX incomplete

where Pe and Te are the exhaust pressure and temperature, respectively. P0 is the pressure at UDS. R is the gas constant of the mixture. X ie is the mole fraction of the i-th species at the exhaust, and X 0i is the mole fraction of each species at UDS. Similarly, the intake exergy is obtained as Equations 14 to 16. The exergy of the combustion loss is relevant to the exergies of UHC and CO at the exhaust and calculated using ( H − T 0S )CO + ( H − T 0S )UHC  = EX fuel

where Tg is the gas temperature in the boundary of the system. Q expresses the amount of heat transfer through boundary conditions. The exergy of output work or the second law thermal efficiency is obtained using

Eq. (18)

Figure 6 presents the adopted procedure in this study. Th is section first presents our comparative analysis on different LTC strategies from the first and second laws of thermodynamics perspectives and considering emission, performance, and fuel consumption to identify the promising strategy.

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The selected strategy is further studied to understand the effects of influencing parameters, including energy fractions of the injections, the SOI timings, injection pressures, spray angles, and EGR rate on exergy destruction and exergy efficiency. These results are then used to identify the optimum range of parameters where exergy destruction is minimum, whereas exergy efficiency is maximum.

highest peak of cylinder pressure, owing to the fast combustion in a short duration. As a result, the HCCI case has a narrower and higher peak of AHRR. In contrast to the HCCI, the CDC case has the lowest peak pressure and lengthened AHRR, which is due to the nature of the diff usion-limited combustion and longer diesel injection duration. Table 4 shows the pressure metrics and performance parameters, including gross IMEP, GTE, PPRR, CA50, Indicated Specific Fuel Consumption (ISFC), and combustion duration of all cases. As it is evident, the HCCI strategy has the highest GTE (or the first law thermal efficiency) in both combustion phasings. The highest GTE of the HCCI can be attributed to the fast combustion and short combustion duration of the HCCI. This is similar to the Otto cycle, which means heat addition at constant volume since all regions undergo combustion at a very short duration. The temperature cut planes of the HCCI case are illustrated in Figure 8, justifying this claim. The RCCI has a longer combustion duration than HCCI due to the reactivity stratification of the n-heptane fuel in the combustion chamber, as shown in Figure 8 and Table 4. The RCCI case also yielded a high GTE (ranked two). By moving toward diff usion-limited combustion like PPC, DDFS, and CDC, the combustion duration increases, which has negative effects on GTE. Although the HCCI and RCCI cases showed high GTE, the PPRR (or engine noise) of these cases is at unacceptable levels (higher than 10 bar/deg). This

Simulation Parameters of the LTC Strategies The cylinder energy input, CA50, boundary conditions, intake pressure, and temperature were considered constant to conduct a reliable comparison between different strategies from the thermodynamic standpoint. To maintain the combustion phasing constant for the different strategies, the injection timings and EGR rate were accordingly modified. All simulation parameters are presented in Table 3. The simulations were performed for two combustion phasings CA50 = 3.7° and 6.5° ATDC. Figure 7 illustrates the pressure and AHRR profi les for all strategies for CA50 = 3.7° ATDC. AHRR is also magnified in Figure 7 to highlight the differences between cases. The results for the ideal-diesel cycle are also presented based on the thermodynamic equations to provide a better comparison. The HCCI strategy has the

TABLE 3   Simulation parameters for all cases.

TBDC (K)

DDFS

PPC

CA50 = CA50 = 3.7° 6.5°

CA50 = 3.7°

385

385

CDC CA50 = 6.5°

CA50 = 3.7°

CA50 = 6.5°

385

HCCI

RCCI

CA50 = CA50 = 3.7° 6.5°

CA50 = CA50 = 3.7° 6.5°

385

385

PBDC (bar)

1.98

1.98

1.98

1.98

1.98

Engine speed (rpm)

1300

1300

1300

1300

1300

Cylinder wall temperature (K)

425

425

425

425

425

Head cylinder temperature (K)

425

425

425

425

425

Piston wall temperature (K)

470

470

470

470

470

EGR (%)

40

47

40

40

0

0

50

55

56

60

SOI1 (Gasoline injection) SOI1 (° ATDC)

−340

−340

−340

−340

Energy fraction (%)

59

66

100

93

Duration (ms)

1.4

1.57

2.38

2.21

Pinjection (bar)

1000

1000

1000

1000

SOI2 (Diesel injection) SOI2 (° ATDC)

−60

Energy fraction (%)

7

−80

100

−11

−8

7

−80

Duration (ms)

0.18

2.5

0.18

Pinjection (bar)

500

1000

500

−80

SOI3 (° ATDC)

−4

Duration (ms)

0.80

−2.8

−9 0.80

Energy fraction (%)

34

34

Pinjection (bar)

1000

1000

−5

© SAE International

SOI3 (Gasoline injection)

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  FIGURE 7    Cylinder pressure and AHRR traces for all cases at CA50 = 3.7° ATDC.

can have devastating effects on the engine’s mechanical parts. The EGR rate of the HCCI and RCCI cases is over 50% at this engine load. It is necessary to emphasize that EGR rates above 60% are not recommended in experiments. Hence, the HCCI and RCCI strategies are not attainable at higher engine loads, which is one of the primary limitations of these strategies. In contrast, DDFS and CDC have acceptable PPRR and engine noise levels. From the perspective of fuel consumption, the kinetically controlled strategies (HCCI and RCCI) have the lowest fuel consumption, while the CDC, as a diff usionlimited strategy, yields the highest ISFC. Figure 8 depicts the differences in temperature contours of the combustion regimes. As previously mentioned, the

HCCI case has a short combustion duration and all regions undergo fast and simultaneous combustion. In the RCCI case, the n-heptane fuel provides a reactivity gradient (or stratification) that permits more extended combustion than HCCI. This reactivity stratification in the RCCI strategy grants RCCI better control over the combustion process compared to HCCI. RCCI still has a kinetically controlled combustion, which means it is not as good as CDC. The CDC strategy is a completely diff usion-limited combustion that causes locally high-temperature regions, which are the source of NOX and soot formation. The DDFS strategy benefits from the RCCI concept and also has direct control over combustion due to the near-TDC injection. The near-TDC injection in the DDFS

TABLE 4   Operating parameters for all cases.

Case 1 (CA50 = 3.7° ATDC) DDFS

PPC

CDC

HCCI

RCCI

Gross IMEP (bar)

10.22

10.59

9.31

11.00

10.80

GTE (%)

51.7

53.9

47.1

55.6

54.6

PPRR (bar/deg)

8.3

27.6

6.6

35.4

22.3

CA50 (° ATDC)

3.7

3.7

3.7

3.7

3.7

ISFC (g/kW-h)

157.3

151.9

172.8

146.1

149.0

Combustion duration (CAD)

28.1

9.6

38.0

4.7

7.9

© SAE International

Case 2 (CA50 = 6.5° ATDC) DDFS

PPC

CDC

HCCI

RCCI

Gross IMEP (bar)

9.91

10.45

9.26

10.93

10.69

GTE (%)

50.1

53.2

46.8

55.2

54.0

PPRR (bar/deg)

6.9

29.6

6.4

26.5

14.7

CA50 (° ATDC)

6.5

6.5

6.5

6.5

6.5

ISFC (g/kW-h)

162.1

153.9

173.6

147.1

150.4

Combustion duration (CAD)

28.3

10.7

39.1

5.4

9.1

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  FIGURE 8    Temperature cut planes for all strategies at similar CA50, but different CA90.

CA50 = 3.7° ATDC

CA90

DDFS

PPC

CDC

HCCI Temperature (K)

strategy has a shorter duration, causing less NOX and soot formation. The results confirm that the iso-octane near-TDC injection in the DDFS leads to five times lower soot formation compared to n-heptane injection [25]. In contrast to the DDFS, reactivity gradients are marginal in the PPC case as shown in Figure 8.

The First Law of Thermodynamics Approach The energy distribution of the first law is illustrated in Figure 9 for two different combustion phasing (CA50 = 3.7° and 6.5° ATDC). The ideal-diesel cycle has the highest GTE because it is isentropic in the expansion and compression strokes. Moreover, incomplete combustion and heat transfer losses irreversibility are zero. Although the HCCI and RCCI cases have higher GTE than other cases, they show a rather high combustion loss due to incomplete combustion and high levels of UHC and CO. The PPC and DDFS cases can be categorized

© SAE International

RCCI

with medium levels of GTE and combustion loss. The combustion loss of the CDC case can be considered negligible. Figure 10 shows the cylinder mean temperature and the heat transfer rates of all cases. Although the CDC case has the lowest mean temperature, it has the highest heat transfer rate due to its long injection duration and the nature of diff usionlimited (or mixing-controlled) combustion of the CDC. The negative effect of the heat transfer rate of the CDC on the GTE is also evident in Figure 9. The HCCI case presents the lowest heat transfer rate among all cases. The exhaust loss of the CDC case, shown in Figure 9, is higher than other LTC strategies, which could be due to the long injection duration of CDC. Exhaust loss is a potential heat source for recovery applications such as turbocharging and aftertreatment purposes.

The Second Law of Thermodynamics Approach The first law of thermodynamics is capable of quantifying the disposition of fuel energy input, and it is a statement of energy

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  FIGURE 9    Energy distribution from the perspective of the first law.

conservation. From the second law of thermodynamics perspective, irreversibility causes energy degradation during the whole process. Exergy analysis is used to evaluate the maximum achievable efficiency for engines. In ICEs, more than 90% of exergy destruction happens during combustion. Other exergy destructions due to pumping and fuel/air mixing process are smaller by two orders of magnitude than exergy destruction in the combustion process [46, 47]. Figure 11 depicts the Temperature-Entropy (T-S) diagram for all strategies and their deviation from the ideal-diesel cycle. The combustion process causes irreversibility in IC engines, and

it is known as the main source of entropy generation, while heat transfer reduces entropy. Among all cases, the HCCI strategy has the highest levels of entropy, owing to the lowest heat transfer rate and maximum cylinder mean temperature as shown in Figure 10. The CDC case has the lowest levels of entropy compared to other LTC strategies due to high levels of heat transfer and lower cylinder temperature and pressure. Figure 12 shows the exergy distribution of all cases as well as the ideal-diesel cycle. All LTC strategies have comparable exergy efficiency (thermal efficiency of the second law) to the ideal-diesel cycle; however, the CDC case has the lowest

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  FIGURE 10    Bulk temperature and heat transfer rate for all cases.

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  FIGURE 11    Temperature-entropy diagram for all cases.

exergy efficiency. Compared to the fi rst law’s thermal efficiency, the second law’s thermal efficiency provides a more realistic criterion to analyze different strategies because the first law shows absolute work without considering the ambient pressure. The variations of the exergy were obtained from the second law of thermodynamics, demonstrating exergy destruction has close relations to the homogeneity of the charge during the combustion process. Thus the CDC has the highest exergy destruction, while advanced LTC strategies

have lower exergy destruction. Among all strategies, HCCI has the highest thermal efficiency in both first law and second law analyses. In addition, HCCI has the lowest exergy destruction compared to other cases. The second law expresses how much of the exhaust, heat transfer, and combustion losses of the first law are attainable with recovery potentials. Take the DDFS strategy as an example; the exhaust energy balance of DDFS for the CA50 = 3.7° ATDC case is shown in Figure 9. The exhaust energy balance is about 37.4% of the fuel energy input. From the

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  FIGURE 12    Energy distribution from the perspective of the second law.

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TABLE 5   Emissions results for all cases.

CA50 = 3.7 DDFS

PPC

CDC

HCCI

RCCI 0.014

NOX (g/kW-h)

0.28

0.19

9.56

0.009

Soot (g/kW-h)

0.005

0.003

0.083

0.0002 0.001

UHC (g/kW-h)

1.66

2.12

0.27

4.31

3.14

CO (g/kW-h)

4.91

3.13

0.84

3.53

5.16

CO2

476.37

455.34

529.51

450.115

452.6

© SAE International

CA50 = 6.5 DDFS

PPC

CDC

HCCI

RCCI

NOX (g/kW-h)

0.15

0.18

6.83

0.005

0.004

Soot (g/kW-h)

0.0054

0.009

0.10

0.0004 0.002

UHC (g/kW-h)

2.34

3.56

0.52

5.76

3.92

CO (g/kW-h)

6.5

5.45

1.13

3.59

6.83

CO2

466.62

456.24

528.3

453.56

458.43

second law perspective, the exhaust exergy of this case is about 17.1% of the fuel exergy input, showing that the highest amount of available energy can be recovered from the exhaust gases in the DDFS case. The exergy destruction is the difference between 37.4% and 17.1%, which is the portion that cannot be recovered. The same conclusion can be drawn for other cases by comparing Figures 9 and 12.

Comparison of Emissions of LTC Strategies Th is section compares the LTC strategies and CDC cases from the emission point of view. As shown in Table 5, the HCCI and RCCI strategies have ultralow NOX and soot. However, these cases lead to higher UHC and CO. The DDFS and PPC cases have higher NOX and soot compared to the

HCCI and RCCI, but they still can meet the EURO6 emission mandate (0.4 g/kW-h and 0.01 g/kW-h for NOX and soot). Consequently, all of the LTC strategies met the EURO6 mandate for NOX and soot without using aftertreatment. On the other hand, the CDC strategy has significantly high levels of NOX and soot, which are 34 and 16.6 times higher than those in the DDFS case. Meeting EURO6 without using extensive aftertreatment seems impossible in the CDC strategy. UHC and CO emissions for all cases are listed in Table 5 for both CA50s. The EURO6 limits for UHC and CO are 0.13 g/kW-h and 1.5 g/kW-h, respectively. Only the CDC case meets the CO limit. Among the LTC strategies, the DDFS has a higher potential for catalytic converter activation (higher exhaust temperature) owing to the near-TDC injection. Table 6 presents a qualitative study of all strategies by taking engine performance, emissions, fuel consumption, and other technical specifications into account. The green color represents a positive attribute, the red color shows a negative quality, and the yellow ones can be  categorized in the medium level. Controllability is a parameter that indicates the ability of each strategy to control combustion phasing directly. HCCI and RCCI are kinetically controlled, and there is no direct control over combustion. In contrast to HCCI and RCCI, DDFS, PPC, and CDC can provide direct control over the combustion. Engine noise level can be indicated by the PPRR index. The acceptable level of the PPRR is less than 10 bar/ deg. DDFS and CDC are the only strategies that meet this criterion. HCCI and RCCI offer limited ranges of operation, and they are sensitive to the initial boundary conditions. In the high engine load operation, it is impossible to maintain the combustion phasing in the HCCI or RCCI strategies. However, DDFS and CDC are not confined to initial boundary conditions and work over wider operating ranges.

TABLE 6   A qualitative study of LTC strategies (red represents a negative quality, yellow represents a medium level, and green

represents a positive quality).

DDFS Controllability Engine noise Operating range Downstream recovery (turbocharger and aftertreatment) Fuel consumption Sensitivity to the changes in initial pressure and temperature NOX © SAE International

Soot UHC + CO Meeting the EURO6 emission mandate for NOX and soot (without aftertreatment) Cyclic variations Exergy destruction

PPC

CDC

HCCI

RCCI

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Downstream recovery is a parameter indicating the potential of a strategy to use the exhaust enthalpy for some applications such as turbocharging or aftertreatment purposes. The HCCI and RCCI cases have low exhaust temperature and enthalpy, especially at low-load operating conditions that have less capability to activate the aftertreatment for UHC and CO oxidization. Conversely, the DDFS, PPC, and CDC have higher exhaust gas temperature and enthalpy owing to the near-TDC injection that is beneficial for the aftertreatment and heat recovery purposes. Fuel consumption and emission were previously discussed. Cyclic variation is another important parameter in IC engines. RCCI and HCCI have higher cyclic variations, particularly at higher loads. Experimental tests performed by Wissink and Reitz [23] clearly depicted these variations in the cylinder pressure and AHRR for the RCCI and PPC strategies. Regarding these experiments, the DDFS has low cyclic variations comparable to the CDC strategy. Exergy destruction is another important thermodynamic parameter for the engine design that must be  kept at the lowest level. Kinetically controlled strategies have the lowest levels of exergy destruction, while the CDC strategy has the highest level of exergy destruction. Table 6 gives a qualitative comparison between all strategies based on the crucial parameters. It can be seen that the DDFS strategy is the best, especially when it comes to mobility applications. Thus it was selected for further investigations to minimize exergy destruction and maximize exergy efficiency.

Effects of Injection Parameters and EGR Rate on the Exergy Destruction and Second Law Thermal Efficiency in the DDFS Strategy In this section, the effects of diesel and gasoline energy fractions and their timings, spray angles, injection pressure for each injector, and EGR rate are studied for the DDFS strategy.

Effects of Diesel Energy Fraction (Ed) and Its Start of Injection (SOI2) The diesel energy fraction was swept from 4% to 10% of the total cylinder energy input, and diesel injection timing (SOI2) was swept from −100° to −30° ATDC. The minimum and maximum range were selected based on the previous numerical study [26] and experimental tests performed by Wissink and Reitz [24]. It is necessary to declare that for SOI2 before −100° ATDC, reactivity stratification is not achieved, and for the SOI2 after −30° ATDC, the combustion regime will change from DDFS to the CDC. Thus SOI2 and Ed must be in the mentioned ranges to meet the RCCI concept in the DDFS strategy. Figure 13 shows the effects of diesel energy fraction (Ed) and timing sweeps (SOI2) on exergy destruction and

  FIGURE 13    The effects of diesel energy fraction (Ed) and timing (SOI2) sweeps on exergy destruction and efficiency at a

© SAE International

constant Eg = 34%, SOI3 = −4° ATDC, θd = 74°, θg = 71.5°, Pd = 500 bar, Pg = 1000 bar, and EGR = 40%.

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efficiency. The other parameters were considered constant as the base case to explore the effects of Ed and SOI2 on exergy destruction and efficiency. The exergy destruction increased by increasing Ed and retarding SOI2. When the amount of diesel fuel increased, the combustion regime moved toward a diffusion-limited regime. It can be seen that as SOI2 retarded toward the TDC, the combustion regime also changed to the diff usion-limited combustion. Consequently, exergy destruction is increased, which has negative effects on exergy efficiency. Considering minimum exergy destruction and maximum exergy efficiency, the suitable ranges of Ed and SOI2 are 4%-6% and −100° to −80° ATDC, respectively, as shown in Figure 13.

Effects of Gasoline Energy Fraction (Eg) and Its Start of Injection (SOI3) This section investigates the effects of gasoline energy fraction (Eg) of the near-TDC injection and its timing (SOI3) on exergy destruction and efficiency. Retarded SOI3 resulted in higher amounts of exergy destruction and lower exergy efficiency. The same trends were observed by increasing the amount of Eg. As a result, considering minimum exergy destruction and maximum exergy efficiency, the suitable ranges for Eg and SOI3 is 25%-30% of the total energy input and −8° to −6° ATDC, respectively, as shown in Figure 14. It is necessary to declare that for the SOI3 before −8° ATDC, the combustion

phasing occurs before TDC that causes negative work and is detrimental to the engine.

Effects of Spray Angles of Gasoline and Diesel Injectors The diesel and gasoline (θd and θg) spray angles are studied by sweeping from 40° to 80° to explore their effects on exergy destruction and efficiency. As is shown in Figure 15, for θg less than 50°, the worst results for exergy destruction and efficiency were observed. Increasing θ g up to 80°, the exergy destruction decreased and hence exergy efficiency increased. Consequently, wide spray angles for gasoline injection (θg) are recommended. In contrast to θg, diesel spray angles more than 70° do not show satisfying results for both exergy destruction and efficiency. A medium spray angle for diesel injection can be  then recommended. As it is evident in Figure  15, the suitable ranges for selecting θd and θ g are between 55°-65° and 65°-75°, respectively.

Effects of Injection Pressure of Gasoline and Diesel Injectors Diesel and gasoline injection pressures were swept from 500 to 2000 bar. Exergy destruction increased when gasoline and diesel injection pressure were at the lowest values (500 bar).

  FIGURE 14    The effects of gasoline energy fraction (Eg) and timing (SOI3) sweeps on exergy destruction and efficiency at

constant Ed = 7%, SOI2 = −60° ATDC, θd = 74°, θg = 71.5°, Pd = 500 bar, Pg = 1000 bar, and EGR = 40%.

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  FIGURE 15    The effects of diesel and gasoline spray angle (θd and θg) sweeps on exergy destruction and efficiency at a constant

© SAE International

Ed = 7%, SOI2 = −60° ATDC, Eg = 34%, SOI3 = −4° ATDC, Pd = 500 bar, Pg = 1000 bar, and EGR = 40%.

At this state, injection duration was the longest among all simulations. The long injection duration of both injectors caused higher exergy destruction, and when injection duration decreased (or injection pressure increased), exergy efficiency improved. In addition, the lower injection pressure of the gasoline injector means a longer duration of the diff usionlimited combustion for DDFS, which has devastating effects on exergy. To avoid exergy destruction, it is necessary to choose the highest injection pressure that contributes to a shorter combustion duration. The suitable ranges for selecting gasoline and diesel injection pressures, as demonstrated in Figure 16, are 1500-2000 bar for both injectors. Gasoline injection pressure seems to have higher influential impacts on exergy destruction rather than the diesel injection since the gasoline injection is the main injection that occurs near TDC, leading to a significant role in controlling diff usion-limited combustion. It supplies 34% of the total energy input, whereas the diesel injection contribution is only 7%.

Effects of the EGR Rate The EGR rate varies from 16% to 65% to explore its effects on exergy destruction and efficiency. For EGR values lower than 16%, combustion phasing (CA50) occurs before the TDC, and it has negative effects on the engine, resulting in negative work. EGR has devastating effects on the combustion process. As shown in Figure 17, exergy destruction is at its lowest value (26.5%) for the EGR = 16%, and exergy efficiency (the second law thermal efficiency) is at its highest amount (48.5%). By increasing the EGR rate, combustion phasing retarded and the peak cylinder pressure decreased; thus, exergy efficiency

reduced up to 41.4%. Compared to the case with a 16% EGR rate, the exergy destruction increased by approximately 2.5%. It can be concluded that a suitable range for the EGR rate is about 16% to 25%.

Summary/Conclusions A 3D-CFD model was developed and validated against experimental data for different loads. The DDFS case was selected to be the base case with 4.71 kJ/cyc total cylinder energy input. To make a reliable comparison, 4.71 kJ/cyc energy input with constant boundary conditions was used for different strategies, including HCCI, RCCI, PPC, and CDC. Energy and exergy analyses were performed for two combustion phasing modes 3.7° and 6.5° ATDC. The main findings are as follows: 1. The HCCI strategy had the highest thermal efficiency obtained using the first and second laws. HCCI showed a 5% lower thermal efficiency with the first law than the ideal-diesel cycle, but it achieved a comparable exergy efficiency to the ideal-diesel cycle. Other strategies can be ranked from highest to lowest based on the thermal efficiency of both laws as RCCI, PPC, DDFS, and CDC. 2. Heat transfer analysis confirmed that the highest magnitude occurred for the CDC strategy with 14.8% of the total energy input; however, the HCCI and RCCI cases had the lowest heat transfer losses (5.6% and 5.8%, respectively). The high value of the CDC’s heat transfer loss can be attributed to the long nearTDC injection duration.

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  FIGURE 16    The effects of diesel and gasoline injection pressure (Pd and Pg) sweeps on exergy destruction and efficiency at a

© SAE International

constant Ed = 7%, SOI2 = −60° ATDC, Eg = 34%, SOI3 = −4° ATDC, θd = 74°, θg = 71.5°, and EGR = 40%.

3. All LTC strategies met the EURO6 limits for NOX and soot emissions; however, the CDC strategy had 24 and 8 times higher NOX and soot emissions than the EURO6 limits, respectively. None of the LTC strategies could meet the EURO6 limits for UHC and CO, but the DDFS strategy has the potential to use an aftertreatment at suitable conversion efficiency, owing to the higher exhaust temperature caused by the near-TDC injection. 4. A qualitative study was performed to compare all strategies based on emissions, performance, fuel consumption, and pressure metrics such as PPRR. It was found that among all strategies, the DDFS is the best for IC engines, particularly considering future mobility purposes. 5. The DDFS strategy was studied from the standpoint of thermodynamics to minimize exergy destruction

and maximize exergy efficiency. The effect of injection parameters including diesel energy fraction (Ed) and its timing (SOI2), gasoline energy fraction (Eg) and its timing (SOI3), spray angles for both injectors (θd and θg), the injection pressure of both injectors (Pd and Pg), and EGR rate were studied on exergy destruction and efficiency. 6. The suitable ranges for the mentioned parameters were 4% < Ed < 6%; −100 < SOI2 < −80° ATDC; 25% < Eg < 30%; −8 < SOI3 < −6° ATDC; 55° < θd < 65°; 65° < θg < 75°; 1500 bar < Pd, Pg < 2000 bar, and 16% < EGR < 25%. Among all considered parameters for the DDFS strategy, the EGR rate had the most significant effects on exergy destruction and efficiency. EVs have many advantages over combustion-based vehicles, such as zero emissions, charging from renewable

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  FIGURE 17    EGR effects on the exergy destruction and second law thermal efficiency.

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sources, smooth torque, high efficiency, and acceleration while moving off; however, they have considerable drawbacks, which make developing the ICEs more necessary for the near future. High costs of batteries, low range of driving, low power density, unsatisfactory battery lifespan, the need for massive infrastructure for charging stations are the most considerable bottleneck of the EVs. On the other hand, Hybrid Electric Vehicles (HEVs) can be  perfect alternatives to EVs. As mentioned previously, the DDFS strategy has comparable thermal efficiency to the RCCI and HCCI, ultralow emissions, high performance, and robust control that makes this strategy very suitable for HEVs. Thus it is necessary to make engines more efficient from the perspective of thermodynamics, so the author believes optimizing the injection parameters of the DDFS from the thermodynamic and emission approaches would be  an interesting future research subject for other scholars.

Contact Information [email protected] [email protected]

Definitions/Abbreviations AHRR - Apparent Heat Release Rate AMR - Adaptive Mesh Refinement ATDC - After Top Dead Center BTDC - Before Top Dead Center BMEP - Brake Mean Effective Pressure CAD - Crank Angle Degree CDC - Conventional Diesel Combustion CFD - Computational Fluid Dynamics CRI - Common Rail Injector DDFS - Direct Dual-Fuel Stratification DDM - Discrete Droplet Model Dur - Duration Ed - Diesel energy fraction Eg - Gasoline energy fraction for the near-TDC injection EGR - Exhaust Gas Recirculation EPA - Environmental Protection Agency EVC - Exhaust Valve Closing EVO - Exhaust Valve Opening EX Destruction - Exergy Destruction HCCI - Homogeneous Charge Compression Ignition ICE - Internal Combustion Engine 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 LTC - Low-Temperature Combustion NOX - Nitrogen Oxides NTC - No Time Counter PCCI - Premixed Charge Compression Ignition Pd - Diesel injection pressure Pg - Gasoline injection pressure PM - Particulate Matter PPC - Partially Premixed Combustion PPRR - Peak Pressure Rise Rate RCCI - Reactivity Controlled Compression Ignition RNG k-ε - Re-Normalization Group k-ε RON - Research Octane Number SOI1 - Start Of Injection for first gasoline injection SOI2 - Start Of Injection for diesel fuel SOI3 - Start Of Injection for near-TDC injection TDC - Top Dead Center T-S diagram - Temperature-Entropy diagram UHC - Unburned HydroCarbon 𝜂2nd law - Exergy efficiency of the second law thermal efficiency θd - Diesel spray angle θg - Gasoline spray angle

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Appendix Mesh Sensitivity Study A mesh sensitivity study was performed to determine a suitable size for grids regarding the accuracy and computational time. Th ree mesh sizes, including 2  mm (coarse case), 1.4  mm (medium case), and 1 mm (fine case), were chosen and applied for the numerical CFD code. Predictions of three mesh sizes are shown in Figure A.1 and Table A.1. The 1.4 mm mesh size (medium case) was chosen for all CFD results, owing to its satisfying results based on accuracy and computational time.

RCCI (Iso-octane and n-Heptane) and CDC Combustion for Other Loads To show the reliability of the developed 3D-CFD model, the model was used for other different modes of combustion, including gasoline/diesel RCCI (at 6.5 and 9.5 bar IMEP) and CDC (at 9.9 bar IMEP), shown in Figure A.2. TABLE A.1   Predicted emissions results for different mesh sizes (system configuration was an Intel Corei7 6700k CPU with 32 GB RAM).

42. Heywood, J.B., Internal Combustion Engine Fundamentals (Singapore: Mc Graw Hill International Editions, 1988). 43. Gingrich, E., Ghandhi, J., and Reitz, R., “Experimental Investigation of Piston Heat Transfer in a Light Duty Engine under Conventional Diesel, Homogeneous Charge Compression Ignition, and Reactivity Controlled Compression Ignition Combustion Regimes,” SAE Int. J. Engines 7(1):375386, 2014, https://doi.org/10.4271/2014-01-1182. 44. Senecal, P., Pomraning, E., Richards, K., and Som, S., GridConvergent Spray Models for internal Combustion Engine CFD Simulations. in ASME 2012 Internal Combustion Engine Division Fall Technical Conference, Vancouver, BC,

NOX (g/ kW-h)

Soot (g/ kW-h)

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

Time Max. cell (h) Min. cell

Coarse mesh (2 mm)

0.291

0.003

4.79

~13

Medium mesh (1.4 mm)

0.280

Case

1.55

~690,000 ~126,000

Fine mesh 0.280 (1 mm)

0.005

4.90

1.66

~20

~960,000 ~252,000

0.005

4.90

1.66

~41

~2,580,000 ~462,000

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© SAE International

  FIGURE A.1    Cylinder pressure and AHRR traces for coarse, medium, and fine mesh sizes.

© SAE International

  FIGURE A.2    Cylinder pressure and AHRR predictions vs. experimental data for RCCI and CDC modes, emissions are in g/kW-h (experimental data taken from [17, 23, 48]).

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