Effects of Different Injection Strategies on Combustion and Emission Characteristics of Diesel Engine Fueled with Dual Fuel

processes

Article Effects of Different Injection Strategies on Combustion and Emission Characteristics of Diesel Engine Fueled with Dual Fuel

Jianbin Luo 1,2 , Zhonghang Liu 1, Jie Wang 1, Heyang Chen 1, Zhiqing Zhang 1,2 , Boying Qin 3 and Shuwan Cui 1,*

1 School of Mechanical and Transportation, Guangxi University of Science and Technology, Liuzhou 545006, China; [email protected] (J.L.); [email protected] (Z.L.); [email protected] (J.W.); [email protected] (H.C.); [email protected] (Z.Z.) 2 Institute of the New Energy and Energy-Saving&Emission-Reduction, Guangxi University of Science and Technology, Liuzhou 545006, China 3 School of Science, Guangxi University of Science and Technology, Liuzhou 545006, China; [email protected] * Correspondence: [email protected]

Abstract: In this work, an effective numerical simulation method was developed and used to analyze the effects of natural gas mixing ratio and pilot-main injection, main-post injection, and pilot-main- post injection strategies on the combustion and emission characteristics of diesel engine fueled with dual fuel. Firstly, the one-dimensional calculation model and three-dimensional CFD model of the engine were established by AVL-BOOST and AVL-Fire, respectively. In addition, the simplified Citation: Luo, J.; Liu, Z.; Wang, J.; chemical kinetics mechanism was adopted, which could accurately calculate the combustion and Chen, H.; Zhang, Z.; Qin, B.; Cui, S. emission characteristics of the engine. The results show that the cylinder pressure and heat release Effects of Different Injection rate decrease with the increase of the natural gas mixing ratio and the NO emission is reduced. When Strategies on Combustion and x Emission Characteristics of Diesel the NG mixing ratio is 50%, the NOx and CO emission are reduced by 47% and 45%, respectively. ◦ Engine Fueled with Dual Fuel. When the SODI3 is 24 CA ATDC, the NOx emission is reduced by 29.6%. In addition, with suitable Processes 2021, 9, 1300. https:// pilot-main injection and pilot-main-post injection strategies, the combustion in the cylinder can be doi.org/10.3390/pr9081300 improved and the trade-off relationship between NOx and soot can be relaxed. Thus, the proper main-post injection strategy can improve the combustion and emission characteristics, especially the Academic Editor: Alessandro reduction in the NOx and CO emissions. D’Adamo Keywords: diesel-natural gas; dual-fuel engine; diesel injection strategy; combustion; emission Received: 16 June 2021 Accepted: 23 July 2021 Published: 27 July 2021 1. Introduction Publisher’s Note: MDPI stays neutral Since the first reciprocating internal combustion engine came out, there have been with regard to jurisdictional claims in published maps and institutional affil- more than 100 years of its history of development up to now. Nowadays, the internal iations. combustion engine is everywhere. Because of its high thermal efficiency, high power per unit weight and volume, and strong mobility, it is widely used in automobiles, engineering equipment, agricultural machinery, ship transportation, and other fields. The international automobile and internal combustion engine industry generally believes that in the fore- seeable future, the internal combustion engine will still be the main motive among these Copyright: © 2021 by the authors. power sources [1]. However, the internal combustion engine and automobile industry not Licensee MDPI, Basel, Switzerland. only promotes the rapid development of the global economy and the continuous progress This article is an open access article of human society, but also worsens the global problem of energy crisis and environmental distributed under the terms and conditions of the Creative Commons pollution [2,3]. With the trend of global warming and the serious contradiction between Attribution (CC BY) license (https:// energy supply and demand, solving the environmental pollution and energy consumption creativecommons.org/licenses/by/ caused by traditional internal combustion engines has become a global urgent mission [4]. 4.0/).

Processes 2021, 9, 1300. https://doi.org/10.3390/pr9081300 https://www.mdpi.com/journal/processes Processes 2021, 9, 1300 2 of 28

The harmful emission from diesel engine mainly includes carbon monoxide (CO), nitrogen oxides (NOx), carbon hydrides (HC), sulfides (SO2, SO3), and particles [5]. These pollution air will cause serious air pollution [6], and endanger human health [7]. For example, CO2 emission from automobiles is the main cause of global warming, sulfide dissolved in rain will form acid rain; nitrogen oxides, and carbon hydride under certain conditions will generate secondary pollutants, namely photochemical smog [8]. In view of these harmful emissions, the countries all over the world have been enacting more and more strict regulations [9], and researchers have also done a lot of researches on the reduction in emissions [10]. The main methods can be divided into three categories: internal purification technology, the external purification technology, and a quest for alternative fuels [11]. Among them, engine internal purification is to start from the source of the engine, fully consider the generation mechanism and influencing factors of emissions [12]. External purification technology mainly refers to the use of particulate filter, oxidation catalytic converter, and selective catalytic reduction technology to make the emissions be treated outside the cylinder [13]. Questing for alternative fuels refers to the use of alcohol fuels, biodiesel, natural gas, ether fuels, and other internationally recognized clean alternative fuels. Considering the short-term non-renewable nature of petroleum, the world faced with the challenge of depletion of oil resources, and the search for alternative fuels should be one of the main development directions [14]. Among many alternative clean fuels, nature gas (NG) has eye-popping reserves. By the end of 2019, the proven recoverable reserves of NG in the world were 198.8 trillion cubic meters. Therefore, NG has been recognized as the most ideal alternative fuel for internal combustion engines due to its advantages of clean combustion, large storage capacity, low price, and wide source [15]. NG engine technology is receiving more and more attention and has become the key research object of the automobile and internal combustion engine industry [16]. With more and more in-depth research, researchers have found that there are many problems to solve when using NG as engine fuel directly, among which the most prominent ones are: Firstly, NG is a gaseous fuel, and the intake efficiency will be reduced if using out of cylinder premixed method [17]. Secondly, the high ignition temperature of natural gas requires high ignition energy [18]. Thirdly, when the mixture ratio of NG and air is not uniform, it will cause a series of engine problems [18]. For example, a thicker mixture will affect the detonation phenomenon, and if the mixture ratio is too thin, it will lead to an increase of CO and THC emission [19]. For some of the above problems, the researchers also put forward some solutions, the diesel NG dual-fuel technology of reaction-controlled compression ignition combustion (RCCI) is one of the most popular technologies [20]. As a low activity fuel, NG is injected into the cylinder through the outside intake manifold during the intake stroke, and mixed with the air in the cylinder in advance to form a combustible mixture [21]. Diesel, as a high activity fuel, enters the cylinder by direct injection at the end of the compression stroke, and, finally, the diesel burns to ignite the mixture [22]. In this mode, NG and air are mixed in advance to form a uniform mixture, diesel atomization forms many ignition sources in the cylinder, and the compression ratio in the cylinder is high. The clean combustion of the mixture is improved due to the advantages. Moreover, this technology does not need to make major changes to the engine [23]. Based on the original diesel engine and function, it can be equipped with gas supply, injection, and control devices [24]. The improved engine can freely switch the fuel to pure diesel or diesel NG dual fuel, and can independently control the mixing ratio of NG. Many institutions and scholars have done a lot of researches on its combustion and emission characteristics, combustion system development, and fuel injection device. Kim et al. [25] thought that the ignition performance of the main fuel plays an important role in the engine performance and emission during the combustion of dual fuel. The experiment on a heavy single-cylinder engine showed that the fuel with high cetane number and aromatics has higher combustion efficiency, lower methane leakage, and THC and CO emission, but it will be at the cost of higher NOx emission. Processes 2021, 9, 1300 3 of 28

Lee et al. [26] studied the emission characteristics of non-methane hydrocarbons (NMHC) and methane under low load conditions on a 6 L heavy engine. The results showed that methane accounted for the largest proportion of the emissions (52–87%) under the general dual-fuel combustion condition due to the low combustion efficiency. The problem was solved by advanced injection time. Millo et al. [27] studied the ignition process of dual-fuel engines through detailed experimental analysis. Based on the detailed chemical reaction mechanism of 0-D, a high precision ignition delay model was developed. Muhssen et al. [28] used FLUENT software to analyze and optimize the mixer of CNG-diesel dual- fuel engine, and the mixture fuel showed good uniformity. Shen et al. [29] had carried out numerical researches on four different piston clearances and piston geometry, pointing out that the main source of carbon hydride emission of NG-diesel dual-fuel engine was similar to spark-ignition engine mainly caused by piston gap. Yang et al. [30] studied the mass, quantity, and particle size distribution characteristics of diesel NG dual-fuel engines under different NG air mixing conditions. The results showed that the mixing state of the mixture has a significant effect on the combustion process. The dual-fuel can effectively reduce particulate emissions, but 60% of them are ultrafine particles. Based on the previous literature and recent research, it can be seen that there are still some bottlenecks to be overcome for this spark ignition diesel NG engine. Compared with pure diesel engines, the dual-fuel can improve the thermal efficiency at low load and reduce the HC and CO emission. In addition, the NG and air enter the cylinder through the intake pipe, so it will occupy part of the gas volume, resulting in a loss of volumetric efficiency. For spark ignition diesel NG dual-fuel engine, the combustion process can be divided into three stages: Firstly, the premixed combustion of premixed diesel NG air mixture [31]. Secondly, diffusion combustion of pilot injected diesel and ignition and flame propagation of natural gas. Finally, the diffusion combustion of residual diesel and the afterburning of natural gas. It is not difficult to see from the combustion process that the first injection time and times of diesel will have an important impact on the combustion results, and it is feasible to reduce these problems by optimizing the injection strategy (such as the number of fuel injections, time, and pressure) At present, some researchers have carried out researches on these issues. For example, Yousefi et al. [32–34] have carried out important research from the aspect of fuel injection strategy to explore the effects of post-injection strategy on greenhouse gas emission of dual-fuel engine under high load. They found the split injection and injection timing are important. Huang et al. [35] and Wang et al. [36] also studied the effects of multiple injections on the combustion and emission of diesel NG dual-fuel engine. They had found the same conclusion. However, most of the previous studies focused on the effects of pilot-main injection on engine combustion and emission characteristics. There is less comprehensive comparative analysis on the four injection strategies of main, pilot-main, main-post, and pilot-main-post. In this work, firstly, the 1D and 3D computing models were established by using AVL- BOOST and AVL-Fire software. Then, some boundary conditions of 3D CFD calculation were obtained based on the 1D model in the previous work. Finally, the corresponding cases of different NG mixing ratios and different injection strategies were simulated by a 3D model. This work helps to solve some of the difficulties currently faced by diesel NG dual-fuel engine.

2. Numerical Methods 2.1. 1-D Numerical Calculation Due to the limitation of the experiment, it is not easy to get some boundary-condition. Therefore, the boundary condition can be obtained by the 1D simulation model. The 1D simulation model of the six-cylinder diesel engine is established by using the AVL-BOOST software. The diagram of the 1D model is shown in Figure1, the control interface of the software can accurately control the proportion of NG involved in the reaction, which makes the combustion control of some multi-component gas or liquid fuels simple. Processes 2021, 9, x FOR PEER REVIEW 4 of 29

Processes 2021, 9, 1300interface of the software can accurately control the proportion of NG involved in the re‐ 4 of 28 action, which makes the combustion control of some multi‐component gas or liquid fuels simple.

Figure 1. SimulationFigure 1. modelSimulation of an model entire dieselof an engine:entire diesel SB—system engine: boundary; SB—system MP—measuring boundary; MP—measuring point; CL—air cleaner; TC—turbocharger;point; CO—cooler;CL—air cleaner; PL—plenum; TC—turbocharger; I—injector; C—cylinder;CO—cooler; R—restriction;PL—plenum; E—engine;I—injector; CAT—catalyticC—cylinder; oxidizer; R—restriction; E—engine; CAT—catalytic oxidizer; PF—diesel particulate filter. PF—diesel particulate ﬁlter.

2.1.1. Combustion2.1.1. Model Combustion Model In the quasi‐dimensionalIn the quasi-dimensional combustion model combustion “mixed model control “mixed combustion” control combustion”(MCC) (MCC) provided by theprovided software by is the selected software as the is selected combustion as the model. combustion In this model.model, Inthe this effective model, the effective flow area and flowturbulent area and kinetic turbulent energy kinetic of the energy injection of the holes injection can be holes calculated can be by calculated input‐ by inputting ting the numberthe of number injection of injectionholes, the holes, diameter the diameter of the orifice, of the the orifice, flow the coefficient flow coefficient of the of the orifice, orifice, and theand pressure the pressure of the of injection the injection track. track. The total The totalcombustion combustion heat heatrelease release rate rateis is [37]: [37]: dQtotal dQPMC dQMCC dddQQQtotal PMC= MCC + (1)  (1) ddddϕ dϕ dϕ

where Qtotal is wherethe heatQ totalreleaseis the rate heat of release combustion, rate of J/deg; combustion, QPMC is J/deg; the totalQPMC heatis release the total rate heat release rate of premixed combustion,of premixed J/deg; combustion, QMCC is J/deg;the totalQ MCCheat isrelease the total rate heat of diffusion release rate combustion, of diffusion combustion, ◦ J/deg; φ is the J/deg;crankshaftϕ is theangle, crankshaft °CA. angle, CA. The premixed Thecombustion premixed was combustion described was by vibe described function by [37]: vibe function [37]: dQPMC dQPMC QPMC a m −a·y(m+1) QPMC a = may·ꞏ(mm1+ 1)y e (2) (2) dϕꞏ1mye∆ϕc d c ϕ − ϕid  id y = , m = 2, a = 6.908 (3) yma,2,6.908∆ϕc (3)  c ◦ ◦ where ϕid is the ignition delay, ; ∆ϕc is the combustion duration, . where φid is the ignitionFor the delay, diffusion °; Δφ combustionc is the combustion part, the duration, heat release °. rate is as follows [38]: For the diffusion combustion part, the heat release rate is as follows [38]: dQ dQMCC MCC = Ccomb· f1(mF, QMCC) f2(k, V) (4)  CfmQfkVcombꞏ,12 F MCC , (4) d dϕ QMCC CEGR Q CEGR f1(mF, Q) =MCCmF − woxygen,available (5) fmQ1,F,  m F w oxygenLCV available (5) LCV √ k ( ) = √ f2 k, kV CRate 3 (6) fkVC2 ,  Rate V (6) 3 V where mF is the mass of fuel injected, kg; Ccomb is the combustion constant; LCV is the where mF is thelow mass caloriﬁc of fuel value injected, of combustion, kg; Ccomb is the kJ/kg; combustionCRate is constant; the mixing LCV rate is the constant; low CEGR is the 2 2 calorific valueinﬂuence of combustion, factor ofkJ/kg; EGR; CkRateis is the the local mixing turbulent rate constant; kinetic energy CEGR is density, the influence m /x ; woxygen,available is the mass fraction of oxygen in the mixture at the time of injection, %; V is the cylinder volume, m3.

Processes 2021, 9, 1300 5 of 28

The dissipation rate is proportional to the turbulent kinetic energy:

dE kin = 0.5 · C · m · v2 − C · E1.5 (7) dt turb F F Diss kin E k = kin (8) vF(1 + λDi f f mstoich)

mF vF = (9) ρF · µA

where the Cturb is the turbulence generation constant; CDiss is the dissipation rate constant; n is the engine speed, r/min; mstoich is the stoichiometric fresh charge mass, kg; λDiff is the excess air coefficient of diffusion combustion; vF is the fuel injection rate, mg/cycle; Ekin 3 represents the kinetic jet energy, J; ρF is the fuel density, g/cm ; A is the sectional area of injection hole, mm2; µ is the flow coefficient of injection hole. According to the Andree and Pacherneg model [37], the ignition delay time was calculated:

dI TUB − Tre f id = (10) dϕ ϕid · Qre f

◦ where Iid is the ignition delay integral, CA; Tref is the reference temperature = 505 K; TUB is the temperature of potential area, K; Qref is the reference reaction energy, kJ. When the integral value of Iid is equal to 1.0 (=ϕid), The value of ignition delay is equal to ◦ ◦ τid = ϕid − ϕSOI, among, ϕSOI is the injection start time, CA; ϕid is the ignition delay, CA. The fuel evaporation rate and droplet diameter can be calculated by Equations (11) and (12):

Td ve = 0.70353 · 4159 T pc·e d (11) q dd = dd,02 − ve · t (12)

where Td is the equilibrium temperature of isothermal droplet evaporation, K; pc is the cylinder pressure, Pa; dd,0 is the initial droplet diameter, mm; dd is the actual droplet 2 diameter, mm; ve is the fuel evaporation rate, g/(cm ·s); t is the fuel evaporation time, s.

2.1.2. Heat Transfer Model The heat transfer from the combustion chamber of the cylinder to the wall surface (cylinder head, piston, and cylinder liner) was calculated according to the following equation [38]: Qwi = f · Ai · αw · (Tc − Twi) (13)

where Qwi is the wall heat transfer coefficient (cylinder head, piston and cylinder liner), 2 2 W/(m ·K); Ai is the wall heat transfer area (cylinder head, piston, and cylinder liner), m ; 2 αw is the convective heat transfer coefficient, W/(m ·K); Tc is the gas temperature in the cylinder, K; Twi is the wall temperature (cylinder head, piston, and cylinder liner), K; f is the heat transfer adjustment factor, which is 1 by default. The Woschni 1978 is employed to calculate the convective heat transfer coefficient in the high-pressure circulation cylinder of the engine at full load [36]. These calculation equations are as follows:

0.8 −0.2 0.8 −0.53 VD · Tc,1 αw = 130 · D · pc · Tc C1 · cm + C2 · · (pc − pc,0) (14) pc,1 · Vc,1

cu C1 = 2.28 + 0.308 · (15) cm 2 · L · n c = (16) m 60 Processes 2021, 9, x FOR PEER REVIEW 6 of 29

0.8  0.2 0.8 0.53 VTDc ,1  wccmcc130DpTCcC 12   ( pp ,0 ) (14) pVcc,1 ,1

cu C1 2.28 0.308  (15) cm Processes 2021, 9, 1300 6 of 28 2Ln c  (16) m 60

where C2 = 0.00324; D is the cylinder diameter, m; cm is the average speed of piston, m/s; cu where C2 = 0.00324; D is the cylinder diameter, m; cm is the average speed of piston, m/s; is the rotation speed of circumferential air flow, m/s; VD is the displacement of single cu is the rotation3 speed of circumferential air ﬂow, m/s; VD is the displacement of single cylinder, m3 ; pc is the gas pressure in the cylinder, Pa; Pc,0 is the initial cylinder pressure, cylinder, m ; pc is the gas pressure in the cylinder, Pa; Pc,0 is the initial cylinder pressure, Pa; Pc,1 is the pressure at the inlet valve closing (IVC), Pa; Vc,1 is the volume of the cylinder Pa; P is the pressure at the inlet valve closing (IVC), Pa; V is the volume of the cylinder c,1 3 c,1 at IVC, m3 ; Tc,1 is the cylinder temperature at IVC, K; L is the stroke of the engine, m. at IVC, m ; Tc,1 is the cylinder temperature at IVC, K; L is the stroke of the engine, m.

2.2.2.2. 3D-CFD 3D‐CFD Calculations Calculations TheThe 3D 3D computing computing model model (as (as shown shown in in Figure Figure2) 2) is is generated generated by by the the ESE-Diesel ESE‐Diesel modulemodule in in AVL-Fire. AVL‐Fire. In In order order to to shorten shorten the the calculation calculation time, time, the the one one eighth eighth of of the the overall overall modelmodel was was used used for for calculation calculation due due to to the the principle principle of of axial axial symmetry symmetry of of eight eight nozzles. nozzles. TheThe grid grid model model at at the the bottom bottom dead dead center center is is shown shown as as follows: follows:

FigureFigure 2. 2.Grid Grid model model for for CFD CFD simulation. simulation.

2.2.1.2.2.1. Fluid Fluid Flow Flow and and Combustion Combustion Model Model TheThe fluid fluid flow flow simulation simulation was was carried carried out out in in the the AVL-Fire AVL‐Fire environment, environment, and and the the improvedimproved turbulence turbulence and and spray/wall spray/wall interaction interaction models models was was used used in thein the process process of simula- of sim‐ tion.ulation. The three-dimensionalThe three‐dimensional CFD CFD model model considers considers the effect the effect of turbulent of turbulent fluid flow fluid on flow fluid. on Thefluid. AVL-Fire The AVL solves‐Fire thesolves momentum, the momentum, energy, energy, three-dimensional three‐dimensional transient transient conservation conser‐ equationsvation equations of mass, turbulent of mass, fluid turbulent flow, and speciesfluid usingflow, the and temporal-differencing species using the scheme tem‐ andporal a finite‐differencing volume. scheme In addition, and the a finite Arbitrary volume. Lagrangian–Eulerian In addition, the method Arbitrary is employed Lagrangi‐ toan–Eulerian calculate the method diffusion is andemployed convection to calculate terms from the chemicaldiffusion sourceand convection terms. Nevertheless, terms from the default codes are not suitable for the simulation due to the two reasons. The default chemical source terms. Nevertheless, the default codes are not suitable for the simulation codes in AVL-Fire environment do not contain the accurate thermos-physical properties of due to the two reasons. The default codes in AVL‐Fire environment do not contain the different fuels. Secondly, the combustion in cylinder should require breaking C-H bonds accurate thermos‐physical properties of different fuels. Secondly, the combustion in cyl‐ and lots of intermediate chemical species are dissociation/formation. The phenomenon of inder should require breaking C‐H bonds and lots of intermediate chemical species are dual fuel is more obvious due to the variety of hydrocarbons with different thermochemical dissociation/formation. The phenomenon of dual fuel is more obvious due to the variety properties and molecular structure. of hydrocarbons with different thermochemical properties and molecular structure. The combustion of the diesel engine includes premixed combustion and diffusion combustion, and diffusion combustion was the main combustion. Therefore, the diesel ignited gas engine model was selected as the ignition model, which was combined with the ECFM-3Z model. In the ECFM-3Z model [39], the transport equations are solved for the averaged quantities of chemical species O2,N2, CO2, CO, H2,H2O, O, H, N, OH, and NO. Here, the averaged means these quantities are the global quantities for the three mixing zones (that is in the whole cell). This equation is classically modeled as: ∂ρyex ∂ρueiyex ∂ κ κt ∂yex + − + = ωx (17) ∂t ∂xi ∂xi Sc Sct ∂xi Processes 2021, 9, 1300 7 of 28

3 where ωx is the combustion source term, kg/(m ·s); yex is the averaged mass fraction of species, %; ρ is the average density, g/cm3; Sc is the Schmidt number; κ is the dynamic 2 viscosity, N·s/m ; uei is the unburned mass fraction of species, %. u The fuel is divided in two parts: the fuel present in the fresh gases, yeFu and the fuel b present in the burnt gases, yeFu. mu mu /V ρu yu = Fu = Fu = Fu (18) eFu m m/V ρ

mb mb /V ρb yb = Fu = Fu = Fu (19) eFu m m/V ρ u b u With yeFu = yeFu + yeFu as the mean fuel mass fraction in the computational cell,%; ρFu b 3 (respectively, ρFu) is the fuel density in the fresh gases(respectively, burnt gases), g/cm ; u b mFu (respectively, mFu) is the mass of the fuel contained in fresh gases (respectively, burnt u gases), mg. A transport equation is used to compute yeFu: u u u ∂ρyeFu ∂ρueiyeFu ∂ κ κt ∂yeFu u u + − + = ρSeFu + ωFu (20) ∂t ∂xi ∂xi Sc Sct ∂xi

u 3 where SeFu is the source term quantifying the fuel evaporation in fresh gases, kg/(m ·s); u ωFu is a source term taking auto-ignition, premixed ﬂame and mixing between mixed unburned and mixed burnt areas into account, kg/(m3·s).

2.2.2. Turbulence Model and Turbulence Diffusion Model In the process of 3D simulation, the scale of turbulence in the cylinder of the internal combustion engine changes greatly, but the CFD software can not directly solve the minimum vortices through the grid, so the turbulence model is introduced in the simulation process. Turbulence models can be divided into Reynolds averaged viscosity model, Reynolds stress model, large eddy model, and separated eddy model. The k-zeta-f four equation turbulence model is used in AVL-Fire due to its good accuracy and stability [40]. There are random and irregular turbulent vortices in the ﬂow process of fuel droplets injected into the cylinder of the diesel engine. The interaction of turbulent vortices interferes with the trajectory of fuel droplets. Therefore, the selection of an appropriate droplet turbulent diffusion model is the premise to ensure the accuracy of the simulation. In this paper, the Fire software’s own enable model was as follows [41]:

1 2 2 Y0 = ( k) · sign(2Rn − 1) · er f −1(2Rn − 1) (21) i 3 i i

" 2 # k k 3 1 tturb = min Cτ , Cl 0 (22) ε ε ug + u − ud

0 3 2 −2 −1 where Yi is the gas phase pulse velocity, m /h; k is turbulent kinetic energy, m s ; erf is the inverse of Gaussian function; Rni is the random number of each vector component, which is determined by the minimum time of turbulence breakup tturb; Cl, Cτ is the model constant, Cl = 0.1644, Cτ = 1.1.

2.2.3. Spray, Collision, Evaporation, and Emission Models Kelvin–Helmholtz/Rayleigh–Taylor (KH-RT) was used to simulate the atomization process of spray [42]. The atomization process of the model can be divided into primary atomization and secondary atomization. In the KH model, due to the unstable growth of the disturbance wave along the ﬂow direction, a breakup will be formed. In the RT model, when the droplet velocity decreases rapidly in the direction of windward, the disturbance wave will be formed at its stagnation point. The fragmentation formed by the unstable Processes 2021, 9, 1300 8 of 28

growth can be applied to the case of relatively high relative velocity and high air resistance at the initial time of spray. The continuous competition between these two phenomena will lead to the breakup of droplets. Wave model equation can be used to simulate KH splitting:

Ra = M1Λ (23)

3.7M2Ra τa = (24) ΛΩ

Λ = f (Wec, Ohd) (25)

Ω = f (Wec, Ohd) (26)

where Wec is the Weber number of gas; Ohd is the number of Onseg; Ra is the radius of new oil drop, mm; M1 is the model constant; M2 is the braking rime cnstant; Ω is the wave growth rate, mm/s; Λ is the length of wave, mm. τa is the breaking time, s. The RT disturbance is described by the fastest growing frequency Ω and the corresponding wave number K:

s 1.5 2 gt | (ρd − ρc) | Ωt = √ · (27) 3 3σ ρd + ρc

1 τt = M5 (28) Ωt r g | (ρ − ρ ) | K = t d c (29) t 3σ π Λt = M4 (30) Kt 2 where gt is the deceleration in the direction of travel, -m/s . If the wavelength Λ is small enough to be growing on the droplet’s surface and the characteristic RT break-up time τt has passed, the droplets atomize and their new sizes are assumed to be proportional to the RT wavelength; σ is the surface tension coefficient; Kt is the number of waves; ρd is the 3 3 density of liquid phase, g/cm ; ρc is the density of gas phase, g/cm ; M4 and M5 are the model adjustable parameters. Furthermore, droplet wall interaction has greatly influenced the direct injection engine, especially for some diesel engines with small cylinder diameter, which will cause most of the fuel atomization evaporation and wall collision at the same time, making the combustion and emission in the simulation results and the test error larger, for instance, incomplete combustion of fuel will worsen the emission of carbon hydride and soot particles. The interaction between the droplet and the wall depends on the velocity, diameter, properties, surface roughness, and temperature of the droplet. In this paper, the Walljet1 model was used to simulate the interaction between droplet and wall [42]. In the simulation process, the mass exchange between the droplet and the oil film on the cylinder wall is not considered, and the model in this paper is satisfied. The model is based on Weber number and Reynolds number spray and wall impingement model. The springback or reflection of droplets is dual by Weber number. Dukowicz model is used for droplet heating and evaporation [43], which can fully consider the evaporation process of the droplet in a non-condensing state. For the simulation of emission, the extended Zeldovich model with mature theory is used for NOx emission [44], and Hiroyasu Nagle Strickland-Constable model is used for soot model [45].

2.2.4. Boundary Condition To obtain the boundary conditions of 3D calculation, the 1D calculation model is built by AVL-BOOST software. After verifying the results of 1D model calculation with the experiment results and 3D model simulation results, some boundary conditions which Processes 2021, 9, 1300 9 of 28

are inconvenient to measure are obtained by 1D calculation. For example, the volume, temperature, and pressure of IVC are 249.30 mL, 420 K, and 197.37 kPa, respectively. In addition, some boundary conditions were calculated by the following equations [4]: (1) Initial turbulent kinetic energy (TKE) of cylinder:

3 3 C 2 E = × U2 = × m (31) TKE 2 0 2 2 n C = 2 · L · (32) m 60

where U0 is the turbulent ﬂuctuation velocity, m/s. (2) Initial cylinder turbulence scale length:

L L = V (33) TSL 2

where Lv is the max. lift of the valve. (3) Fuel injection mass per cycle:

be peτn 3 Vb = × 10 (34) 120n f ρF

where Vb is the fuel injection mass per cycle, g/cycle; be is the brake speciﬁc fuel consumption, g/(kw·h); pe is the power, kW; τn is the number of strokes; nf is the diesel engine speed, r/min; i is the number of cylinders.

3. Engine Setup and Experimental Cases 3.1. Engine Setup The experiment was carried out on a six-cylinder turbocharged intercooled four-stroke diesel engine reﬁtted into a diesel engine. The speciﬁc parameters of the engine are shown in Table1, and the layout of the engine is shown in Figure3. Among them, the intake system is installed on the intake manifold, and the natural gas enters the cylinder together with the air to form a homogeneous mixture, which can accurately control the injection quantity of NG. The common rail injection system was powered by a high-pressure fuel pump. The improved diesel engine was directly connected with the hydraulic dynamometer. NOx, CO, and HC emissions were measured by Horiba exhaust gas analyzer, the AVL DEWE-2010CA diesel engine combustion monitoring system was adopted, and the electronic unit pump control system was used to accurately control the fuel injection quantity of the engine. In the experiment, the injection pressure of the injector remains constant at 150 MPa. In addition, the diesel and natural gas properties parameters were shown in Table2.

Table 1. Engine speciﬁcation.

Parameter Unit Value Number of cylinders - 6 Compression ratio - 17 Bore × Stroke mm 126 × 155 Displacement L 11.5 Connecting rod mm 235 Cooling system - water-cooling Fuel injection - Common rail system Inlet valve closing deg 564 Exhaust valve opening deg 848 Number of holes - 8 Hole diameter µm 0.121 Included spray angle deg 145 Injection pressure MPa 150 Processes 2021, 9, x FOR PEER REVIEW 10 of 29

gine are shown in Table 1, and the layout of the engine is shown in Figure 3. Among them, the intake system is installed on the intake manifold, and the natural gas enters the cylinder together with the air to form a homogeneous mixture, which can accurately control the injection quantity of NG. The common rail injection system was powered by a high‐pressure fuel pump. The improved diesel engine was directly connected with the hydraulic dynamometer. NOx, CO, and HC emissions were measured by Horiba exhaust gas analyzer, the AVL DEWE‐2010CA diesel engine combustion monitoring system was adopted, and the electronic unit pump control system was used to accurately control the fuel injection quantity of the engine. In the experiment, the injection pressure of the in‐ Processes 2021, 9, 1300 10 of 28 jector remains constant at 150 MPa. In addition, the diesel and natural gas properties parameters were shown in Table 2.

Natural gas compressor ECU Flowmeter 2 Flowmeter 1 Fuel tank Pressure gauge Sensors

DOC DPF Flowmeter 3 Valve 1 Electric Diesel engine Gas analyzer dynamometer

Gas analyzer Valve 2 Valve 3 Sampled-data PC display systems

2 Writ . P 6 ro tec e ted

T 6 ape .0 in Us e

C U 100 leaning se

Tape

2 00 T T o o Re H Op Wa P ress B Un H Close this Ins H Op Wa L mo a e ioa a er t a e oa n hn t it d n n n it ve dl dl T dl th L d T e is Lig ut e ape e is ap h ton igh C e t t Ha Op e ompr e nd ra l ss e te

Ov errideD e n s tyi d i g i t a l U n lo a d

S De ns i el ec T t ty Z 8 7

FigureFigure 3.3. Schematics ofof experimentalexperimental device.device.

Table 2.1. DieselEngine and specification. natural gas properties parameters. Parameter Unit Value Related Parameters Natural Gas Diesel Number of cylinders ‐ 6 ◦ Ignition temperatureCompression ( ratioC) ‐ 65017 260 Octane number RON 130 20–30 Bore × Stroke mm 126 × 155 Theoretical air fuel ratio 17.25 14.6 Displacement L 11.5 Low calorific value (MJ·kg−1) 50.05 44.4 Connecting rod mm 235 High calorific value (MJ·kg−1) 55.54 47.1 Calorific valueCooling of mixture system ‐ water‐cooling 3230 3790 (kJ/mFuel3 )injection ‐ Common rail system BoilingInlet point valve (◦C) closing −161deg 564 280 DensityExhaust (kg/m valve3) opening 0.716deg 848 814 CHNumber4 (%) of holes ‐ 97.68 - Hole diameter μm 0.121 3.2. ExperimentalIncluded Cases spray angle deg 145 Injection pressure MPa 150 Numerical simulation technology can be used to analyze the combustion process in the cylinder and provide information of intermediate products and related free radicals. In the paper, the process of numerical simulation, which starts from IVC (564 ◦CA) to EVO (848 ◦CA) is studied. In the initial stage of the simulation, it was assumed that NG and air were uniformly mixed, and the pressure and temperature of the mixture were also homogeneous. During the experimental process, the engine speed was at the constant speed of 1800 rpm. Then, the cycle simulation under different injection strategies was shown in Figure4. Among them, Case 1 calculated the results of four injection strategies of pure diesel, while Case 2 calculated the results of single main injection when burning dual fuel with different NG mixing ratios, which were used as the basis of comparative analysis. According to Equation (33), the energy fraction of natural gas is calculated by the low calorific value and mass flow rate of fuel. Similarly, according to Equation (34), the proportion of the mass flow rate of pre injection, main injection, or post injection to the total mass of diesel is calculated. M LHV %NG = NG NG (35) MNG LHVNG + MD LHVD Processes 2021, 9, 1300 11 of 28

MD-pilot,main,post %FR = (36) MD-total

where %NG is the energy fraction of natural gas, %; MNG (respectively, MD) is the mass of NG (respectively, diesel), mg; LHVNG (respectively, LHVD) is the low heat value of NG (respectively, diesel), MJ·kg−1;%FR is the percentage of injection volume consumed by

Processes 2021, 9, x FOR PEER REVIEWeach injection strategy to total cycle injection volume, %; MD-pilot,main,post is the mass12 of of29 diesel injected for each injection strategy, mg; MD-total is the total mass of diesel injected in cycle, mg.

Figure 4. DifferentDifferent injection strategy cases.

3.3. Model Model Validation Validation InIn AVL AVL-BOOST‐BOOST and AVL AVL-Fire‐Fire environment, the default codes do not have accurate thermophysicalthermophysical properties, which cannotcannot accuratelyaccurately describedescribe thethe formation/dissociation formation/dissociation of intermediate intermediate chemical chemical species, species, especially especially for for multi multi component component fuels. fuels. Therefore, Therefore, a sim a‐ plifiedsimpliﬁed chemical chemical kinetic kinetic mechanism mechanism of diesel of diesel NG dual NG dualfuel is fuel used is usedto calculate to calculate the com the‐ combustion process of cylinder, and the HRR of the 3D CFD simulation is obtained from bustion process of cylinder, and the HRR of the 3D CFD simulation is obtained from the the reaction heat source term, which is proved by Reference [36]. In addition, the HRR for reaction heat source term, which is proved by Reference [36]. In addition, the HRR for 1D 1D simulations and experiments can be obtained by the following equation: simulations and experiments can be obtained by the following equation: dp dp dQdQT γ ddcyppcy11 cy cy T=ppVcy + Vcy (37) dθd1d1d γ−1 cydθ γ − 1 cy dθ ◦ where ddQQTT/dθθ isis thethe HHR, HHR, J/deg; J/deg;θ θis is the the crankshaft crankshaft angle, angle,; p cy°; ispcy cylinder is cylinder pressure, pressure, MPa; 3 MPa;V is cylinder V is cylinder volume, volume, m ; γ mis3 the; γ is speciﬁc the specific heat capacity,heat capacity, J/(kg ·J/(kg.K).K). The experiment was carried out on a six six-cylinder‐cylinder diesel engine under different load conditions.conditions. The The cylinder cylinder pressure pressure and heat release rate (HRR) from the simulation results were compared with the experimental results and shown in Figure 5. The results show that the cylinder pressure, HRR, NOx, and CO emissions obtained by numerical simula‐ tion are in agreement with the experimental results. The maximum error is less than 5%. Therefore, the accuracy of the model is reliable (see Figure 5a–c).

Processes 2021, 9, 1300 12 of 28

were compared with the experimental results and shown in Figure5. The results show that Processes 2021, 9, x FOR PEER REVIEWthe cylinder pressure, HRR, NOx, and CO emissions obtained by numerical simulation13 of are 29 in agreement with the experimental results. The maximum error is less than 5%. Therefore, the accuracy of the model is reliable (see Figure5a–c).

10 16 Experiment Experiment 9 1-D model 14 1-D model 3-D model 3-D model 8 1800rpm 30%load 24 12 1800rpm-75%load 7 20 10 80 6 16 8 5 60 12 6

4 (J/deg) 40 CylinderPressure(MPa) CylinderPressure(MPa) HRR(J/deg)

8 HRR 4 3 20 2 4 2

1 0 0 0 -40 -20 0 20 40 -40 -20 0 20 40 Crank angle(°CA) Crank angle(°CA) (a) In‐cylinder pressure (b) HRR 0.006

NOx mass fraction 0.005 CO mass fraction

0.004

0.003

0.002 Experiment value (‰)

0.001

0.000 0.000 0.001 0.002 0.003 0.004 0.005 0.006 Simulation value (‰)

Figure 5. Comparisons of experiment and simulation results.

4. Results and Discussion In the paper, D100 is the pure diesel, D75 is the 25% NG addition blends with 75% diesel by mass;mass; D60 D60 is is the the 40% 40% NG NG addition addition blends blends with with 60% 60% diesel diesel by mass;by mass; D50 D50 is the is 50% the 50%NG additionNG addition blends blends with 50% with diesel 50% bydiesel mass; by D40 mass; is the D40 60% is the NG 60% addition NG blendsaddition with blends 40% withdiesel 40% by mass;diesel D25 by mass; is the D25 75% is NG the addition 75% NG blends addition with blends 25% diesel with by25% mass. diesel The by results mass. Theof Case results 1 and of Case Case 1 2 and are usedCase 2 as are the used basis as for the comparative basis for comparative analysis with analysis Cases with 3–5. Cases The 3–5.effects The of differenteffects of NG different mixing ratio,NG mixing injection ratio, strategies injection (main strategies injection, (main pilot-main injection, injection, pi‐ lotmain-post‐main injection, injection, main and‐post pilot-main-post injection, and injection), pilot‐main pilot‐post injection injection), start pilot time injection (SODI1), start and timepost injection(SODI1), startand timepost (SODI3)injection onstart the time combustion (SODI3) and on emissionthe combustion characteristics and emission of the characteristicsdiesel NG dual-fuel of the engine diesel NG are analyzed.dual‐fuel engine In the pure are analyzed. diesel case, In the the single-hole pure diesel cycle case, fuel the ◦ singleinjection‐hole is 19.60cycle mg,fuel theinjection start ofis diesel19.60 mg, injection the start (SODI2) of diesel at − injection8 CA in (SODI2) the main at injection −8 °CA inphase, the main and theinjection injection phase, duration and the in injection the main duration injection in phase the main is constant. injection In phase the cases is con of‐ stant.the subsequent In the cases addition of the subsequent of NG, the fueladdition mass of of NG, the the pre-injection, fuel mass mainof the injection, pre‐injection, and mainpost-injection injection, is and controlled post‐injection by adjusting is controlled the injection by adjusting duration. the injection duration. In the process of simulation calculation, the deﬂagrationdeflagration will easily occur when thethe main injection injection time time is is set set too too early early in in the the dual dual fuel fuel mixed mixed combustion combustion case case [34]. [34 It]. is It due is due to theto the fact fact that that the the chemical chemical reaction reaction sensitivity sensitivity of of the the mixture mixture in in the the cylinder increasesincreases when NG is added into the homogeneous mixture. If the injection advance angle is ear‐ lier, the combustion in the cylinder will release heat prematurely. Due to the effects of heat conduction and heat radiation, the combustible mixture in some areas of the cylin‐ der will reach the ignition point before the flame spreads. It leads to a sudden increase in

Processes 2021, 9, 1300 13 of 28

Processes 2021, 9, x FOR PEER REVIEW 14 of 29 when NG is added into the homogeneous mixture. If the injection advance angle is earlier, the combustion in the cylinder will release heat prematurely. Due to the effects of heat conduction and heat radiation, the combustible mixture in some areas of the cylinder will regionalreach the pressure ignition pointand temperature. before the flame Thus, spreads. based It on leads the toprevious a sudden research increase experiences in regional [36],pressure the main and temperature. injection time Thus, was basedset at − on8 °CA the previous ATDC. research experiences [36], the main injection time was set at −8 ◦CA ATDC. 4.1. Combustion Characteristics 4.1. Combustion Characteristics Figure 6a–c shows the cylinder pressure and HRR of the pure diesel case and the Figure6a–c shows the cylinder pressure and HRR of the pure diesel case and the dual‐fuel case with different NG mixing ratios. It can be seen that pure diesel (D100) has dual-fuel case with different NG mixing ratios. It can be seen that pure diesel (D100) has the highest peak cylinder pressure, followed by D75, D60, D50, D40, and D25. Based on the highest peak cylinder pressure, followed by D75, D60, D50, D40, and D25. Based on the the analysis of the HRR, the diesel injection time in the pure diesel case is earlier than that analysis of the HRR, the diesel injection time in the pure diesel case is earlier than that in in the dual‐fuel case. In Figure 6a, the cylinder pressure will be higher in the condition of the dual-fuel case. In Figure6a, the cylinder pressure will be higher in the condition of the the ahead‐of‐time heat release and the combustion phase, and more sufficient combus‐ ahead-of-time heat release and the combustion phase, and more sufficient combustion in tion in the compression stroke. In addition, as the proportion of NG mixing increases, the the compression stroke. In addition, as the proportion of NG mixing increases, the cylinder cylinder peak pressure generally shows a decreasing trend. It is due to the fact that NG peak pressure generally shows a decreasing trend. It is due to the fact that NG and air are and air are introduced into the cylinder in the form of a homogeneous mixture. As the introduced into the cylinder in the form of a homogeneous mixture. As the proportion of proportion of NG increases, the excess air coefficient and the intake air volume decrease. NG increases, the excess air coefficient and the intake air volume decrease. Moreover, as Moreover, as the mixing ratio of NG increases, the amount of diesel injected into the the mixing ratio of NG increases, the amount of diesel injected into the cylinder will also cylinderdecrease, will and also the spreaddecrease, speed and of the the spread flame willspeed be of slowed the flame down, will resulting be slowed in a lowdown, initial re‐ sultingHRR. The in a pressure low initial and HRR. temperature The pressure decrease and accordingly temperature in thedecrease initial accordingly combustion in stage. the initialSimilarly, combustion other studies stage. had Similarly, the same other results studies [46,47 had]. the same results [46,47].

18 100 D100 D100 16 D75 D75 D60 D60 80 14 D50 D50 D40 12 D40 D25 D25 60 10 SODI2=-8°CA ATDC

8 40 D100 D75 HRR(J/deg) 6 D60

Cylinder pressure(MPa) Cylinder D50 4 20 D25 2 D40

0 0 -40-30-20-100 10203040 -20-100 10203040 Crank angle(°CA) Crank angle(°CA) (a) In‐cylinder pressure (b) HRR

Figure 6. InIn-cylinder‐cylinder pressure and HRR under different natural gas mixing ratios.ratios.

In Figure 66b,b, itit cancan be be seen seen that that as as the the NG NG mixing mixing rate rate increases, increases, the the spontaneous spontaneous combustion of diesel is suppressed, although the amount of NG involved in the premixed combustion stage increases. This leads to a significant signiﬁcant trend of decreasing the peak HRR in the diffusion combustion phase, and as the NG mixing rate continues to increase,increase, thethe combustion delay also gradually increases, D100 has the smallest ignition delay angle, followed by D75, D60, D50, D40, and D25. It is due to the fact that the increase of the NG mixing ratio, resulting in a decrease in the overall fuel activity, the combustion start angle is delay andand combustioncombustion end end angle angle is is advance, advance, more more fuel fuel burning burning after after the the TDC. TDC. When When the theNG NG mixture mixture ratio ratio is 75%, is 75%, a signiﬁcant a significant increase increase in ignition in ignition delay delay angle angle is observed. is observed. Figures7 7 and and8 show8 show the the in-cylinder in‐cylinder temperature temperature distribution distribution and and diesel diesel distribution distribu‐ tionat different at different NG mixing NG mixing ratios. ratios. It can It be can seen be thatseen the that diesel the diesel injection injection time in time the in pure the diesel pure diesel case is earlier than that of the dual fuel, and a better‐atomized fuel is formed at −6 °CA ATDC. As the NG fraction in dual‐fuel increases, the lag time of diesel injection in‐ creases. Thus, the fuel atomization time was also delayed correspondingly. It can be seen that the premixed combustion in the cylinder is over, and the diffusion combustion phase

Processes 2021, 9, 1300 14 of 28

case is earlier than that of the dual fuel, and a better-atomized fuel is formed at −6 ◦CA Processes 2021, 9, x FOR PEER REVIEW 15 of 29 ATDC. As the NG fraction in dual-fuel increases, the lag time of diesel injection increases. Thus, the fuel atomization time was also delayed correspondingly. It can be seen that the premixed combustion in the cylinder is over, and the diffusion combustion phase has been hasentered been at entered the time at of the TDC time when of TDC the when diesel the fraction diesel is fraction high. Similarly, is high. Similarly, the start of the premixed start of premixedcombustion combustion is relatively is relatively late in dual late fuel in dual with fuel a high with NG a high fraction. NG fraction. In addition, In addition, it can be it canfound be thatfound it isthat basically it is basically at the stageat the where stage where the average the average temperature temperature of combustion of combustion in the incylinder the cylinder is the highest is the highest at the 15.7 at ◦theCA 15.7 ATDC. °CA With ATDC. the increaseWith the of increase NG mixing of NG proportion mixing proportionin dual fuel, in the dual high-temperature fuel, the high area‐temperature in the cylinder area distributes in the cylinder more widely, distributes resulting more in widely,a higher resulting average temperaturein a higher average in the cylinder. temperature in the cylinder.

−6 °CA ATDC 0 °CA TDC +15.7 °CA ATDC

D100

D75

D60

D50

D40

D25

Figure 7. In‐cylinder temperature distribution field with different NG mixing ratios. Figure 7. In-cylinder temperature distribution ﬁeld with different NG mixing ratios.

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Processes 2021, 9, 1300 15 of 28

−6 °CA ATDC 0 °CA TDC +4 °CA ATDC

D100

D75

D60

D50

D40

D25

Figure 8. In‐cylinder evaporated fuel mass distribution field with different NG mixing ratios. Figure 8. In-cylinder evaporated fuel mass distribution ﬁeld with different NG mixing ratios.

FigureFigure9 9shows shows the the cylinder cylinder pressure pressure and and HRR HRR in in the the cylinder cylinder with with a a single single main main injectioninjection and and pilot-main pilot‐main injections injections at differentat different SODI1. SODI1. When When the SODI1 the SODI1 at −30 ◦atCA −30 ATDC, °CA theATDC, peak the pressure peak pressure increases increases from 16.8 from MPa 16.8 to MPa 18.6 to MPa 18.6 compared MPa compared with the with single the single main injection.main injection. There There is an increaseis an increase of about of 10.7%about (See10.7% Figure (See9 a).Figure However, 9a). However, when SODI1 when furtherSODI1 advancesfurther advances from − 30from◦CA −30 ATDC °CA toATDC−38 to◦CA −38 ATDC °CA ATDC and − 40and◦CA −40 ATDC, °CA ATDC, the peak the pressurepeak pressure in cylinder in cylinder decreases decreases and the ignition and the delay ignition angle delay is further angle reduced. is further Based reduced. on the analysisBased on of the the analysis HRR, it is of obvious the HRR, that it compared is obvious with that a compared single main with injection, a single the main pilot-main injec‐ injectiontion, the strategypilot‐main improves injection the strategy combustion improves and increases the combustion the peak pressureand increases in the the cylinder peak (Seepressure Figure in9 theb). Withcylinder the (See advancement Figure 9b). of With the SODI1,the advancement the cylinder of pressurethe SODI1, and the the cylinder peak HRRpressure both and show the a peak trend HRR of ﬁrst both increase show a and trend then of decrease. first increase It is dueand tothen the decrease. fact that It the is in-cylinderdue to the fact mixture that the is mixed in‐cylinder more fullymixture with is themixed advance more offully the with SODI1, the increasingadvance of thethe chemicalSODI1, increasing sensitivity the of thechemical fuel. Therefore, sensitivity the of heatthe fuel. release Therefore, advance the and heat accompany release advance by low temperatureand accompany heat by release low temperature (LTHR), the combustionheat release delay (LTHR), angle the of combustion the diffusion delay combustion angle of phasethe diffusion is reduced, combustion as the SODI1 phase continuously is reduced, advance, as the theSODI1 combustion continuously delay angleadvance, further the decreases,combustion keeps delay the angle main further combustion decreases, close keeps to the the TDC. main combustion close to the TDC.

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20 150 20 Main injection 150 Main injection Main injection Main injection 18 SODI1=-24°CA ATDC SODI1=-24°CA ATDC 18 SODI1=-24°CA SODI1=-30°CA ATDC ATDC SODI1=-24°CA ATDC SODI1=-30°CA ATDC SODI1=-30°CA ATDC 16 SODI1=-38°CA ATDC 120 SODI1=-30°CA ATDC 16 SODI1=-38°CA SODI1=-40°CA ATDC ATDC 120 SODI1=-38°CA ATDC SODI1=-40°CA ATDC SODI1=-38°CA ATDC 14 SODI1=-40°CA ATDC 14 SODI1=-40°CA ATDC SODI2=-8°CA ATDC 12 90 SODI2=-8°CA ATDC 12 90 10 10 -38°CA -38°CA-30°CA 8 -30°CA 60 8 HRR(J/deg) 60 HRR(J/deg) 6 -40°CA Cylinder Pressure(MPa) 6 -40°CA Cylinder Pressure(MPa) 4 30 4 -24°CA 30 LTHR -24°CA LTHR 2 2 0 0 0 -40-30-20-100 10203040 0 -40-30-20-100 10203040 -40 -30 -20 -10 0 10 20 30 40 Crank angle(°CA) -40 -30 -20 -10Crank an 0gle(°CA 10) 20 30 40 Crank angle(°CA) Crank angle(°CA) (a) In‐cylinder pressure (b) HRR (a) In‐cylinder pressure (b) HRR Figure 9. In‐cylinder pressure and HRR under pilot‐main injection strategy. Figure 9. InIn-cylinder‐cylinder pressure and HRR under pilot pilot-main‐main injection strategy. Figure 10 shows the cylinder pressure and HRR with single main injection and Figure 1010 showsshows the the cylinder cylinder pressure pressure and and HRR HRR with with single single main main injection injection and main- and main‐post injection strategy at different SODI3. It can be seen that compared with single mainpost injection‐post injection strategy strategy at different at different SODI3. SODI3. It can It be can seen be that seen compared that compared with single with single main main injection, the peak pressure in the cylinder and the peak HRR are reduced. With the maininjection, injection, the peak the pressure peak pressure in the cylinder in the cylinder and the and peak the HRR peak are HRR reduced. are reduced. With the With delay the of delay of the SODI3, there is no effect on the peak cylinder pressure in the diffusion delaythe SODI3, of the there SODI3, is no effectthere onis theno peakeffect cylinder on the pressurepeak cylinder in the diffusionpressure combustion in the diffusion stage, combustion stage, but the peak HRR decreases, the post‐injection diesel slightly increases combustionbut the peak stage, HRR decreases,but the peak the HRR post-injection decreases, diesel the post slightly‐injection increases diesel the slightly heat releaseincreases in the heat release in the later combustion stage. With the◦ SODI3 from +10 °CA ATDC◦ delay the heat later release combustion in the stage. later combustion With the SODI3 stage. from With +10 theCA SODI3 ATDC from delay +10 to °CA +30 ATDCCA ATDC, delay to +30 °CA ATDC, the peak pressure in cylinder changes little, because SODI3 has no tothe +30 peak °CA pressure ATDC, in the cylinder peak changes pressure little, in cylinder because SODI3changes has little, no effect because on the SODI3 combustion has no effect on the combustion start stage (i.e., premixed combustion process), but the peak effectstart stage on the (i.e., combustion premixed combustionstart stage (i.e., process), premixed but the combustion peak HRR isprocess), signiﬁcantly but the lower. peak It HRR is significantly lower. It is due to the fact that only 75% of the original fuel injection HRRis due is to significantly the fact that lower. only 75%It is due of the to originalthe fact that fuel only injection 75% of in the original main injection fuel injection phase in the main injection phase and 25% of diesel in post injection is injected to cylinder, not inand the 25% main of dieselinjection in postphase injection and 25% is of injected diesel toin cylinder,post injection not conducive is injected to to the cylinder, premixed not conducive to the premixed combustion and diffusion combustion. conducivecombustion to and the diffusionpremixed combustion. combustion and diffusion combustion.

75 18 75 18 Main injection Main injection Main injection Main SODI3=+10°CA injection ATDC SODI3=+10°CA ATDC 16 SODI3=+10°CA ATDC 16 SODI3=+10°CA SODI3=+16°CA ATDC ATDC SODI3=+16°CA ATDC SODI3=+16°CA ATDC 60 SODI3=+16°CA ATDC SODI3=+24°CA ATDC 60 SODI3=+24°CA ATDC 14 SODI3=+24°CA ATDC 14 SODI3=+24°CA SODI3=+30°CA ATDC ATDC SODI3=+30°CA ATDC SODI3=+30°CA ATDC SODI2=-8°CA ATDC SODI2=-8°CA SODI3=+30°CA ATDC ATDC 12 SODI2=-8°CA ATDC SODI2=-8°CA ATDC 12 45 45 10 10 8 8 +16°CA 30 +16°CA 30HRR(J/deg) 6 +10°CA HRR(J/deg) 6 +10°CA Main Main Cylinder Pressure(MPa) Cylinder

Cylinder Pressure(MPa) Cylinder 4 +24°CA 4 15 +24°CA +30°CA 15 +30°CA 2 2 0 0 0 -40-30-20-100 10203040 0 -20-100 102030405060 -40-30-20-100Crank angle(°CA 10203040) -20-100 102030405060 Crank angle(°CA) Crank angle(°CA) Crank angle(°CA) (a) In‐cylinder pressure (b) HRR (a) In‐cylinder pressure (b) HRR Figure 10. In‐cylinder pressure and HRR under main‐post injection strategy Figure 10. In-cylinderIn‐cylinder pressure andand HRRHRR underunder main-postmain‐post injectioninjection strategy.strategy Figure 11 shows the cylinder pressure and HRR with single main injection and pi‐ Figure 11 shows the cylinder pressure and HRR with single main injection and pi‐ lot‐main‐post injection strategy with different SODI1 and SODI3. It can be seen that the lot‐main‐post injection strategy with different SODI1 and SODI3. It can be seen that the

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Processes 2021, 9, x FOR PEER REVIEW 18 of 29

Figure 11 shows the cylinder pressure and HRR with single main injection and pilot- main-post injection strategy with different SODI1 and SODI3. It can be seen that the peak peakpressure pressure in the in cylinder the cylinder and and the the peak peak HRR HRR is slightlyis slightly higher higher than than that that of of single single main injection. With With the the advance of the SODI1 and the delay of the SODI3, the peak cylinder pressure and the peak HRR both show a trend of ﬁrstfirst increase and then decrease,decrease, the postpost-injection‐injection diesel slightly increases the heat release in the later combustioncombustion stage. It is due to the fact that the pilot injection makes the mixture mix better, promotes the heat release in advance, reduces the combustion delay angle, and facilitates the premixed and diffusion combustion, mainmain combustion combustion close close to to the the TDC. TDC. At At the the same same time, time, the the increase increase in inpeak peak pressure pressure and and peak peak HRR HRR can becan seen be seen that pilotthat injectionpilot injection has a greaterhas a greater effect on effect engine on enginecombustion combustion compared compared to post to injection. post injection.

18 90 Main injection Main injection SODI1,3=-30/+13°CA ATDC SODI1,3=-30/+13°CA ATDC 16 SODI1,3=-32/+17°CA ATDC SODI1,3=-32/+17°CA ATDC SODI1,3=-36/+21°CA ATDC 75 SODI1,3=-40/+25°CA ATDC SODI1,3=-36/+21°CA ATDC 14 SODI2=-8°CA ATDC SODI1,3=-40/+25°CA ATDC SODI2=-8°CA ATDC 12 60

10 -32/+17°CA 45 8 -30/+13°CA HRR(J/deg) -36/+21°CA Main 6 30 Cylinder Pressur(MPa) 4 LTHR 15 2 -40/+25°CA

0 0 -40-30-20-100 10203040 -40 -30 -20 -10 0 10 20 30 40 50 60 Crank angle(°CA) Crank angle(°CA) (a) In‐cylinder pressure (b) HRR

Figure 11. InIn-cylinder‐cylinder pressure and HRR under pilot-main-postpilot‐main‐post injection strategy.

Figure 12a12a shows the cylinder pressure and HRR of pure diesel and dual fuel (75% diesel25% NG) with four different injection strategies, respectively. The results show that the effects of different injection strategies on the combustion characteristics of pure diesel and dual fuel are the same. Figure 12b12b further confirms confirms the previous simulation results. InIn-cylinder‐cylinder temperaturetemperature distribution distribution field field with with different different fuel injectionfuel injection strategies strategie is showns is ◦ shownas Figure as 13 Figure. Adding 13. aAdding pilot injection a pilot atinjection−30 CA at − ATDC30 °CA before ATDC the before single the main single injection main injectioncan significantly can significantly increase theincrease cylinder the peakcylinder pressure peak pressure and peak and HRR, peak advance HRR, theadvance heat ◦ therelease heat and release reduce and the reduce ignition the ignition delay angle. delay Adding angle. Adding a post-injection a post‐injection at +30 atCA +30 ATDC °CA ATDCafter the after single the main single injection main injection will result will in aresult decrease in a indecrease the peak in pressure the peak and pressure peak HRR. and peakIn addition, HRR. theIn addition, pilot-main-post the pilot injection‐main‐post strategy injection can slightly strategy increase can slightly the peak increase pressure the in peakthe cylinder pressure and in peakthe cylinder HRR, which and peak also advance HRR, which the heat also release advance and the reduce heat therelease ignition and reducedelay angle. the ignition delay angle.

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100 20 Main injection Main injection (SODI2=-8°CA ATDC) (SODI2=-8°CA ATDC) 18 Pilot-main injection Pilot-main injection (SODI1,2=-30/-8°CA ATDC) Main-post injection (SODI1,2=-30/-8°CA ATDC) 16 (SODI2,3=-8/+12°CA ATDC) 80 Main-post injection Pilot-main-post injection (SODI2,3=-8/+12°CA ATDC) 14 (SODI1,2,3=-40/-8/+25°CA ATDC) Pilot-main-post injection (SODI1,2,3=-40/-8/+25°CA ATDC) 12 60

10 Pilot-main-post 8 Main 40 HRR(J/deg) 6 Main-post Cylinder Pressure(MPa) 4 Pilot-main 20

0 0 -40-30-20-100 10203040 -40-30-20-100 10203040 Crank angle(°CA) Crank angle(°CA) (a) Single diesel 80 20 Main injection (SODI2=-8°CA ATDC) Main injection 18 Pilot-main injection (SODI2=-8°CA ATDC) (SODI1,2=-30/-8°CA ATDC) 70 Pilot-main injection Main-post injection (SODI1,2=-30/-8°CA ATDC) 16 (SODI2,3=-8/+16°CA ATDC) Main-post injection Pilot-main-post injection 60 (SODI2,3=-8/+16°CA ATDC) (SODI1,2,3=-40/-8/+25°CA ATDC) 14 Pilot-main-post injection (SODI1,2,3=-40/-8/+25°CA ATDC) 50 12

10 40 Main 8 Pilot-main HRR(J/deg) 30 6 Cylinder Pressure(MPa) Cylinder 20 4 Main-post

2 10 Pilot-main-post 0 0 -40 -30 -20 -10 0 10 20 30 40 -40-30-20-100 10203040 Crank angle(°CA) Crank angle(°CA) (b) Dual fuel

Figure 12. InIn-cylinder‐cylinder pressure and HRR of pure diesel and dual fuel with different injectioninjection strategies.strategies.

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Main and Pilot‐main injection −28 °CA ATDC −6 °CA ATDC 0 °CA TDC 15.7 °CA ATDC

SODI2 = −8 °CA ATDC

SODI1 = −30 °CA ATDC

SODI1 = −40 °CA ATDC

Main‐post injection −6 °CA ATDC 0 °CA TDC 18 °CA TDC +35 °CA ATDC

SODI3 = +16 °CA ATDC

SODI3 = +30 °CA ATDC

Pilot‐main‐post injection −28 °CA ATDC −6 °CA ATDC 0 °CA TDC +35 °CA ATDC

SODI1,3 = −30/+13 °CA ATDC

SODI1,3 = −40/+25 °CA ATDC

FigureFigure 13. 13.In-cylinder In‐cylinder temperature temperature distribution distribution ﬁeldfield withwith different fuel injection strategies. strategies.

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4.2. Emission Characteristics 4.2. Emission Characteristics 4.2.1. NOx Emission 4.2.1. NOx Emission x FigureFigure 14 shows the NOx emissionsemissions with with different NG mixing ratios and different x dieseldiesel injection strategies.strategies. It It can can be be seen seen that that the the NO NOx formation formation increases increases rapidly rapidly between be‐ tween0 and 300 and◦CA 30 ATDC. °CA ATDC. Compared Compared with the with pure the diesel pure case, diesel the case, combustion the combustion of dual fuelof dual can fuelsigniﬁcantly can significantly reduce NOreducex emission. NOx emission. Moreover, Moreover, NOx emission NOx emission ﬁrstly firstly decreases decreases and then and thenincreases increases with with the increase the increase of the of mixing the mixing ratio ratio of NG. of ForNG. example, For example, compared compared with purewith purediesel diesel case, case, NOx emissionNOx emission of D50 of isD50 reduced is reduced by 47%. by In47%. addition, In addition, when when the proportion the propor of‐ tionNG inof dual-fuelNG in dual increases‐fuel increases from 50% from to 60%, 50% NO to x60%,emission NOx emission increase by increase 24%. It by is due24%. to It the is duehigh to temperature the high temperature caused by the caused improved by the combustion. improved Withcombustion. the increase With of the the NGincrease mixing of theratio, NG the mixing mixture ratio, fuel the in the mixture cylinder fuel is in fully the mixedcylinder and is thefully combustion mixed and process the combustion a is more processcomplete a is and more rapid. complete and rapid.

0.0045 0.0030 Main injection D100 SODI1=-24°CA ATDC D75 SODI1=-30°CA ATDC 0.0025 D60 0.0036 SODI1=-38°CA ATDC D50 SODI1=-40°CA ATDC SODI2=-8°CA ATDC 0.0020 D40 D25 0.0027

0.0015 0.0018 mass fraction(‰) mass x massfraction(‰) x 0.0010 NO NO 0.0009 0.0005

0.0000 0.0000 -40-200 20406080 -40-30-20-100 1020304050607080 Crank angle(°CA) Crank angle(°CA) (a) Mixing ratio. (b) Pilot‐main injection. 0.0030 0.0035 Main injection Main injection SODI3=+10°CA ATDC SODI1,3=-30/+13°CA ATDC SODI3=+16°CA ATDC 0.0030 SODI1,3=-32/+17°CA ATDC 0.0025 SODI1,3=-36/+21°CA ATDC SODI3=+24°CA ATDC SODI1,3=-40/+25°CA ATDC SODI3=+30°CA ATDC SODI2=-8°CA ATDC SODI2=-8°CA ATDC 0.0025 0.0020

0.0020 0.0015 0.0015 massfraction(‰) 0.0010 x Mass Fraction(‰)

x 0.0010 NO NO 0.0005 0.0005

0.0000 0.0000 -40-30-20-100 1020304050607080 -40-30-20-100 1020304050607080 Crank angle(°CA) Crank angle(°CA) (c) Main‐post injection. (d) Pilot‐main‐post injection.

Figure 14. NOx emission for different NG mixing ratios and injection strategy. Figure 14. NOx emission for different NG mixing ratios and injection strategy.

FigureFigure 14b14b shows the NO x emissionemission with with the the pilot pilot-main‐main strategy. strategy. It can be seen that thethe pilot pilot-main‐main injection strategy can greatlygreatly increaseincrease NONOxx emission.emission. In addition,addition, NONOxx emissionsemissions show an increasing trend with the increase of the SODI1. Compared with main injection,injection, the the NOx emissionemission of of pilot pilot injection injection increase increase by 65.3% when the advance angle is −−4040 °CA◦CA ATDC. ATDC. It It is is due due to to the the high high temperature temperature caused caused by by the the increased increased advance advance angle. angle. The high temperature is beneficial beneﬁcial to the formation of NOx..

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Figure 14c shows the NOx emission with the main-post injection strategy. It can be seen that the main-post injection strategy can reduce NOx emission compared with main injection strategy. With the increase of SODI3, the NOx emission firstly decreases and then increases. Compared with main injection, the NOx emission of main-post injection strategy ◦ increases by 29.6%when the advance angle is +24 CA ATDC. In addition, NOx emission decreases by 7.6% when the SODI3 delay is from +24 ◦CA ATDC to +30 ◦CA ATDC. It is due to the fact that only 75% of the original fuel injection in the main injection phase and 25% of diesel in post-injection phase is injected to cylinder. Thus, it significantly reduces the temperature in the cylinder during the main fuel injection phase, thereby inhibiting the generation of NOx. Figure 14d shows the NOx emission with pilot-main-post injection strategy. It can be seen that the NOx emission firstly decreases and then increases with increase of advance angle. It is due to the fact that the increase of pilot injection and post injection. Only 50% original diesel is injected to cylinder in the main injection phase. However, when the advance angle of the SODI1 and the delay time of the SODI3 further increase, the sufficient pre-mixing allows the mixture to be widely distributed in the cylinder. After being ignited by diesel, the cylinder was filled with high-temperature gas, which greatly result in the increase of NOx emission.

4.2.2. Soot Emission The multiple injection technology not only improves the combustion process, but also reduces the pollutant emissions, and optimizes the trade-off relationship between soot and NOx [32]. Figure 15 shows the soot emission with different NG mixing ratios and different fuel injection strategies. It can be found that the soot mass fraction in the cylinder firstly increases and then decreases with the increase of NG mixing ratio. The soot emission reaches the highest and lowest when the ratio of NG is 60% and 75%, respectively. The soot emission is mainly distributed in the high temperature and oxygen-poor area, and then gradually reduced by oxidation. With the addition of pilot injection and post injection, the emission of soot changes from a single peak of the main injection to a double peak. When there is no pilot injection, soot emission is mainly distributed in the front of the combustion chamber. However, in the pilot injection and post injection stage, soot is mainly generated at the bottom and rear of the combustion chamber. The use of pilot-main injection will slightly increase the emission of soot. It is due to the fact that the amount of fuel injected in the pilot injection stage is less, and the local high temperature generated by combustion and the oxygen-deficient environment in the middle of the fuel beam create conditions for soot generation. When the pilot-main-post injection strategy was adopted, the emission curve of soot shows three peaks, and the three peaks show a trend of first decreases and then increases with the advance of the SODI1 and the delay of the SODI3. From the changing trend, it can be seen that the SODI3 has a greater impact on this strategy. Processes 2021, 9, 1300 22 of 28 Processes 2021, 9, x FOR PEER REVIEW 23 of 29

0.05 0.05 D100 Main injection D75 SODI1=-24°CA ATDC D60 SODI1=-30°CA ATDC 0.04 0.04 D50 SODI1=-38°CA ATDC SODI1=-40°CA ATDC D40 SODI2=-8°CA ATDC D25 0.03 0.03

0.02 0.02 Soot MassSoot Fraction(‰) Soot Mass Fraction(‰) Mass Soot 0.01 0.01

0.00 0.00 -90 -60 -30 0 30 60 90 -90 -60 -30 0 30 60 90 Crank angle(°CA) Crank angle(°CA) (a) Mixing ratio. (b) Pilot‐main injection. 0.05 Main injection Main injection SODI1，3=-30/+13°CA ATDC SODI3=+10°CA ATDC 0.05 SODI1，3=-32/+17°CA ATDC SODI3=+16°CA ATDC SODI1，3=-36/+21°CA ATDC 0.04 SODI3=+24°CA ATDC SODI1，3=-40/+25°CA ATDC SODI3=+30°CA ATDC 0.04 SODI2=-8°CA ATDC SODI2=-8°CA ATDC

0.03 0.03

0.02 0.02 Soot MassSoot Fraction(‰) Soot MassFraction(‰) Soot 0.01 0.01

0.00 0.00 -90 -60 -30 0 30 60 90 -90 -60 -30 0 30 60 90 Crank angle(°CA) Crank angle(°CA) (c) Main‐post injection. (d) Pilot‐main‐post injection. Figure 15. Soot emission for different NG mixing ratios and injection strategy. Figure 15. Soot emission for different NG mixing ratios and injection strategy. 4.2.3. HC Emission 4.2.3. HC Emission Figure 16 shows the HC emission and methane distribution with different NG Figure 16 shows the HC emission and methane distribution with different NG mixing mixing ratios and injection strategies. Figure 16a shows that the HC emission firstly in‐ ratios and injection strategies. Figure 16a shows that the HC emission ﬁrstly increases creases in the initial stage and then decreases with the increase of NG mixing ratio. in the initial stage and then decreases with the increase of NG mixing ratio. Compared Compared with pure diesel, the NG mixing generated more HC in the early stage. with pure diesel, the NG mixing generated more HC in the early stage. However, a large However, a large amount of HC was oxidized at −8 °CA ATDC. As seen in Figure 16b, amount of HC was oxidized at −8 ◦CA ATDC. As seen in Figure 16b, the NG combustion the NG combustion can be clearly promoted by using the pilot‐main injection strategy. can be clearly promoted by using the pilot-main injection strategy. Some previous studies Some previous studies have found that the main source of HC emission is the incomplete have found that the main source of HC emission is the incomplete combustion of dual fuel. combustion of dual fuel. The combustion phase was also advanced with the increase of The combustion phase was also advanced with the increase of SODI1. Thus, more HC will SODI1.be oxidized. Thus, Asmore seen HC in will Figure be oxidized.16c, it can As be seen seen in that Figure the NG 16c, combustion it can be seen rate that decreases the NG combustionwith the main-post rate decreases injection with strategy. the main As the‐post post injection injection strategy. diesel enters As the the post cylinder, injection the dieselHC emission enters the will cylinder, appear the a second HC emission small peak will appear at the back.a second It is small due topeak the at fact the thatback. the It isreduction due to the in fact injection that the mass reduction leads to in the injection decrease mass of in leads cylinder to the temperature decrease of inin thecylinder main temperaturecombustion stagein the and main the combustion increase of incompletelystage and the burned increase NG. of incompletely Figure 16d shows burned that NG. the Figurepilot-main-post 16d shows injection that the strategy pilot‐ improvesmain‐post the injection combustion strategy and improves accelerates the the combustion combustion andrate ofaccelerates NG, but the the post combustion injection dieselrate of also NG, forms but athe second post HC injection emission diesel peak. also In addition,forms a second HC emission peak. In addition, Figure 17 shows the combustion rate of CH4 un‐ Figure 17 shows the combustion rate of CH4 under different injection strategies. It can be derseen different that the pilot-maininjection strategies. injection and It can pilot-main-post be seen that injection the pilot strategies‐main injection can advance and thepi‐ lot‐main‐post injection strategies can advance the combustion of CH4 and make it burn

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more fully at TDC. On the contrary, the main‐post injection is not conducive to the combustion of CH4 and make it burn more fully at TDC. On the contrary, the main-post morecombustion fully atof TDC.CH4 in On cylinder. the contrary, the main‐post injection is not conducive to the injectioncombustion is not of conduciveCH4 in cylinder. to the combustion of CH4 in cylinder.

2.5 1.0 2.5 1.0 D100 2.0 D100D75 0.8 2.0 D75D60 ) 0.8 ) ％ )

％ D60D50 ) ％

％ 1.5 D50D40 0.6 1.5 D40D25 0.6 D25

1.0 0.4 1.0 0.4 HCFraction( Mass Main injection HC MassFraction(

HCFraction( Mass MainSODI1=-24°CA injection ATDC HC MassFraction( 0.5 0.2 SODI1=-24°CASODI1=-30°CA ATDCATDC 0.5 0.2 SODI1=-30°CASODI1=-38°CA ATDCATDC SODI1=-38°CASODI1=-40°CA ATDCATDC SODI2=-8°CA ATDC SODI1=-40°CA ATDC 0.0 0.0 SODI2=-8°CA ATDC -90 -60 -30 0 30 60 90 -90 -60 -30 0 30 60 90 0.0 Crank angle(°CA) 0.0 -90 -60 -30 0 30 60 90 -90 -60 -30Crank an 0gle(°CA 30) 60 90 Crank angle(°CA) Crank angle(°CA) (a) Mixing ratio. (b) Pilot‐main injection. 1.0 (a) Mixing ratio. 1.0 (b) Pilot‐main injection. Main injection 1.0 1.0 Main injection MainSODI1,3=-30/+13°CA injection ATDC MainSODI3=+10°CA injection ATDC SODI1,3=-30/+13°CASODI1,3=-32/+17°CA ATDCATDC 0.8 SODI3=+10°CASODI3=+16°CA ATDCATDC 0.8 SODI1,3=-32/+17°CASODI1,3=-36/+21°CA ATDCATDC SODI3=+24°CA ATDC SODI1,3=-40/+25°CA ATDC 0.8 SODI3=+16°CA ATDC ) 0.8 SODI1,3=-36/+21°CA ATDC ) SODI3=+30°CA ATDC SODI2=-8°CA ATDC

SODI3=+24°CA ATDC ％ SODI1,3=-40/+25°CA ATDC ) ％ ) SODI3=+30°CASODI2=-8°CA ATDC ATDC SODI2=-8°CA ATDC ％％ 0.6 SODI2=-8°CA ATDC 0.6 0.6 0.6

0.4 0.4 0.4 0.4 HCMassFraction( HC Mass Fraction( HCMass HCMassFraction(

HC Mass Fraction( HCMass 0.2 0.2 0.2 0.2

0.0 0.0 -90 -60 -30 0 30 60 90 -90 -60 -30 0 30 60 90 0.0 0.0 -90 -60 -30Crank an 0gle(°CA) 30 60 90 -90 -60 -30Crank an 0gle(°CA) 30 60 90 Crank angle(°CA) Crank angle(°CA) (c) Main‐post injection. (d) Pilot‐main‐post injection. (c) Main‐post injection. (d) Pilot‐main‐post injection. Figure 16. HC emission for different NG mixing ratios and injection strategy. Figure 16. HC emission for different NG mixing ratios and injection strategy. Figure 16. HC emission for different NG mixing ratios and injection strategy. 100 100

0 ꞏs) 3 0 ꞏs) 3 (kg/m -1004 (kg/m -1004

-200 -200 Main injection -300 MainPilot-main injection injection -300 Pilot-mainMain-post injection

Combustion Rate of CH of Rate Combustion Main-postPilot-main-post injection injection

Combustion Rate of CH of Rate Combustion Pilot-main-post injection -400 -400-30 -20 -10 0 10 20 30 -30 -20 -10Crank angle(°CA) 0 10 20 30 Crank angle(°CA) Figure 17. CH4 combustion rate with different injection strategies. Figure 17. CH4 combustion rate with different injection strategies. Figure 17. CH4 combustion rate with different injection strategies.

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4.2.4. CO Emission 4.2.4. CO Emission CO is a kind of intermediate product, which is mainly caused by the incomplete CO is a kind of intermediate product, which is mainly caused by the incomplete combustion due to the low temperature and lack of oxygen in some areas during combustion due to the low temperature and lack of oxygen in some areas during inin‐cylinder combustion. Figure 18 shows the CO emission under different NG mixing cylinder combustion. Figure 18 shows the CO emission under different NG mixing ratios ratios and injection strategies. It can be seen that the CO emission is significantly reduced and injection strategies. It can be seen that the CO emission is signiﬁcantly reduced with with the increase of NG mixing ratio in the early stage. It is due to the fact that the higher the increase of NG mixing ratio in the early stage. It is due to the fact that the higher concentration of NG is helpful to combustion, the higher temperature in the cylinder, and concentration of NG is helpful to combustion, the higher temperature in the cylinder, the more abundant positive gas are conducive to the conversion of CO into CO2. Figure and the more abundant positive gas are conducive to the conversion of CO into CO2. 18b–d show that the pilot‐main injection, main‐post injection, and pilot‐main‐post injec‐ Figure 18b–d show that the pilot-main injection, main-post injection, and pilot-main-post injectiontion strategies strategies of can of canfurther further reduce reduce the the CO CO emission emission in in the the early early stage. stage. It It is is due due to thethe improved combustion caused by thethe NG inin dualdual fuel.fuel. At TDC,TDC, COCO mainlymainly distributeddistributed inin the outer region of diesel injectioninjection trajectory, wherewhere thethe fuelfuel concentrationconcentration waswas highhigh andand the oxygen was poor.poor. In addition, when thethe injectioninjection advanceadvance angleangle decreases,decreases, thethe COCO emission increases. It is due to the fact that thethe fuelfuel injectedinjected laterlater doesdoes notnot havehave enoughenough time to oxidize.

21 21 Main injection D100 SODI1=-24°CA ATDC 18 D75 18 SODI1=-30°CA ATDC D60 SODI1=-38°CA ATDC D50 SODI1=-40°CA ATDC 15 15 D40 SODI2=-8°CA ATDC D25 12 SODI2=-8°CA ATDC 12

9 9

6 CO Mass Fraction(‰) CO Mass

CO Mass Fraction(‰) Mass CO 6

3 3

0 -90 -60 -30 0 30 60 90 0 Crank angle(°CA) -90 -60 -30 0 30 60 90 Crank angle(°CA) (a) Mixing ratio. (b) Pilot‐main injection. 21 21 Main injection Main injection SODI1,3=-30/+13°CA ATDC SODI3=+10°CA ATDC 18 18 SODI1,3=-32/+17°CA ATDC SODI3=+16°CA ATDC SODI1,3=-36/+21°CA ATDC SODI3=+24°CA ATDC SODI1,3=-40/+25°CA ATDC 15 SODI3=+30°CA ATDC 15 SODI2=-8°CA ATDC SODI2=-8°CA ATDC

12 12

9 9

CO Mass Fraction(‰) Mass CO 6 6 CO Mass Fraction(‰) Mass CO

3 3

0 0 -90 -60 -30 0 30 60 90 -90 -60 -30 0 30 60 90 Crank angle(°CA) Crank angle(°CA) (c) Main‐post injection. (d) Pilot‐main‐post injection. Figure 18. CO emission for different NG mixing ratios and injection strategy. Figure 18. CO emission for different NG mixing ratios and injection strategy.

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5. Conclusions With the continuous deterioration of energy crisis [48–55] and environmental problems [56–62], finding effective methods to optimize the combustion of internal combustion engine and reduce emission has become the research focus. However, the fuel injection strategies play an important role in improving performance and reducing emission. There- fore, this paper studied the mixing ratio and the different injection strategies (i.e., main injection, pilot-main injection, main-post injection, and pilot-main-post injection) on the combustion and emission characteristics of diesel NG dual-fuel engine. The main conclusions are as follows: (1) With the increase of NG mixing ratio, the max. cylinder pressure is reduced, but the cylinder temperature is increased. In addition, the NOx and CO emission are reduced. When the NG mixing ratio is 50%, the NOx and CO emission are reduced by 47% and 45%, respectively. However, the HC emission increases. (2) Compared with the single main injection, the pilot-main injection strategy can significantly improve the cylinder pressure and HRR. When the SODI1 is −30 ◦CA ATDC, the cylinder pressure increases by 19.6% and the cylinder temperature also increases by 4.6%. In terms of emission, the pilot-main injection can significantly reduce HC and CO in the cylinder. The soot emission firstly decreases and then increases, but ◦ NOx emission increases, when SODI1 is −38 CA ATDC. (3) Compared with the single main injection, the main-post injection strategy can reduce the cylinder pressure and HRR. However, the cylinder temperature is reduced in the main injection stage. The NOx and CO emissions is reduced. When the SODI3 is ◦ 24 CA ATDC, the NOx emission is reduced by 29.6%. (4) Compared with the single main injection, the pilot-main-post injection strategy slightly increase the cylinder pressure, HRR, and cylinder temperature. In terms of emission, it can effectively reduce the HC and CO emission. Due to the effects of pilot injection and post injection of diesel, the soot emission in cylinder increases, but the NOx emission in cylinder firstly decreases and then increases.

Author Contributions: Conceptualization, J.L., Z.Z. and S.C.; methodology, H.C, Z.Z. and S.C.; software, Z.L., J.W.; validation, J.L., Z.L. and B.Q.; formal analysis, J.L., Z.Z. and S.C.; investigation, J.L., H.C. and S.C.; resources, J.L. and S.C.; data curation, J.L.; writing—original draft preparation, J.L., Z.L. and S.C.; writing—review and editing, J.L., Z.L., H.C. and S.C.; visualization, J.L. and S.C.; supervision, J.L. and S.C.; project administration, Z.Z. and S.C.; funding acquisition, Z.Z. All authors have read and agreed to the published version of the manuscript. Funding: This work is supported by the Guangxi Basic Ability Improving Program of 2020KY08015, the Natural Science Foundation of Guangxi under the research grant 2018GXNSFAA294122; the Guangxi Special Program for Young Talents Guike(AD20159066) and the Guangxi University of Science and Technology Doctoral Fund under the research grants of 20Z22 and 20S04. Data Availability Statement: All data used to support the findings of this study are included within the article. Conflicts of Interest: The authors declare that they have no conflict of interest regarding the publica- tion of this paper.

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