A Study on the Effect of Ignition Timing on Residual Gas, Effective Release Energy, and Engine Emissions of a V-Twin Engine

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Article A Study on the Effect of Ignition Timing on Residual Gas, Effective Release Energy, and Engine Emissions of a V-Twin Engine

Quach-Nhu Yhcmute 1, Nguyen-Xuan Khoa 1,2 and Ocktaeck Lim 1,*

1 School of Mechanical Engineering, University of Ulsan, San 29, Mugeo2-dong, Nam-gu, Ulsan 44610, Korea; [email protected] (Q.-N.Y.); [email protected] (N.-X.K.) 2 Faculty of Automobile Technology, HaNoi University of Industry, No.298, Cau Dien Street, Bac Tu Liem District, Ha Noi 100000, Vietnam * Correspondence: [email protected]; Tel.: +82-10-7151-8218

Abstract: The ignition timing of an SI engine is a critical parameter. The influence on residual gas, effective release energy, and emissions characteristics of ignition timing for the V-twin engine is investigated in this research. For this purpose, an experiment system was built with a dynamometer, and a model of the simulation was created. In this research, the ignition timing was varied from 10 to 45 degrees BTDC under full load operating conditions, with engine speeds ranging from 3000 to 10,000 rpm. Based on the output data, ignition timing has a major impact on the proportion of residual gas, efficient release energy, performance of the engine, and the emission characteristics. The smallest proportion of residual gas was 0.07% at 8000 rpm and ignition timing of 10 ◦CA. At 15 ◦CA of ignition timing, the highest efficient release energy was 0.817 kJ at 4000 rpm, while at 8000 rpm and 25 ◦CA of ignition timing, it was 0.8305 kJ. At 6000 rpm, the greatest braking torque of the engine Citation: Yhcmute, Q.-N.; Khoa, was 21.57 Nm, while the minimal BSFC was 343.821 g/kWh. The nitrogen oxide emission and HC N.-X.; Lim, O. A Study on the Effect emission increase with the advanced ignition timing, but CO emission decreases. of Ignition Timing on Residual Gas, Effective Release Energy, and Engine Keywords: ignition timing; emissions; effective release energy; residual gas; engine Emissions of a V-Twin Engine. Energies 2021, 14, 4523. https:// doi.org/10.3390/en14154523

Academic Editor: Andrzej 1. Introduction Teodorczyk Global population growth and technological advancements have resulted in a significant increase in the number and use of road vehicles that are powered by fossil fuels. Received: 7 June 2021 Problems associated with the depletion of fossil fuels and increasing environmental pollu- Accepted: 13 July 2021 tion are on the rise due to this rapid development [1]. Due to these issues, better thermal Published: 27 July 2021 efficiency and lower emissions from engines are increasingly mandated [2]. Many factors influence the emission characteristics and performance of SI engines [3,4]; ignition timing Publisher’s Note: MDPI stays neutral is among the most important [5]. Ignition timing is strongly influenced by flame speed, with regard to jurisdictional claims in which has a significant impact on the SI engine performance [6]. published maps and institutional affil- The thermodynamics of an engine with high ignition retardation was studied by Chan iations. and Zhu, who focused on the impact of a retarded spark on pressure distribution in the engine cylinder. Temperature of the gas inside the cylinder and trapped mass were also calculated under different spark timing conditions [7]. To analyze the effect on the burning rate and pressure of the gas in the cylinder of the spark timing and the composition of Copyright: © 2021 by the authors. the fuel and equivalence ratio for a natural gas engine, a model for Zero-D two zone Licensee MDPI, Basel, Switzerland. thermodynamics was built by Soylu and Van Gerpen [8]. Under a variety of operating This article is an open access article conditions, characteristics of burning rates are analyzed to clarify the initiation, propagation, distributed under the terms and and termination of the flame. Hedfi et al. studied bioethanol combustion and discovered conditions of the Creative Commons that retarded spark timing leads to higher temperature and pressure of the gas mixture Attribution (CC BY) license (https:// in the cylinder [9]. The optimal injection and ignition timing for a methanol engine was creativecommons.org/licenses/by/ investigated by Li et al. They discovered that these timings had a major impact on methanol 4.0/).

Energies 2021, 14, 4523. https://doi.org/10.3390/en14154523 https://www.mdpi.com/journal/energies Energies 2021, 14, 4523 2 of 18

engine combustion and exhaust emissions [10]. A fuel injection system with a high-pressure injector and with a hollow cone spray was studied by Beatrice et al. [11,12]. They report that high injector pressure reduces peak firing pressure, increasing fuel efficiency. In past years, the battery ignition manifold’s centrifugal and vacuum regulators were used to control ignition timing. Manufacturers are being forced to produce increasingly complicated engines, and their parts as emissions regulations tighten. There are a number of options for meeting the emission standards. One method involves incorporating con- temporary structural components: injection systems with high pressure, all electronically controlled ignition systems, etc. [13]. To better satisfy user requirements and operating conditions, electronic controls are increasingly replacing mechanical controls. Factors such as speed, load, temperature, and change of fuel type in an engine are the operational conditions, particularly given the penchant for biomass fuels or biofuels in recent years. Alcohol-based biofuels (for SI engines) have a higher octane rating compared with gasoline because they contain oxygen [14]. The higher octane number of the fuel, in addition to its improved anti-knocking properties, can change the timing of ignition required to reach decreased emissions. At high loads, an earlier ignition timing helps to decrease the temperature of exhaust gas [15,16]. As a result, adding fuel to the mixture in order to prevent engine parts from rising in temperature is unnecessary. The engine would run more efficiently if the dose injected were reduced [17]. Lower fuel consumption is expressed in quality of ignition control based on conditions of operation and impacts overall efficiency, leading to increased engine power [18]. The quality of the combustion is a reflection of the engine’s efficiency. In addition, the change in concentration of exhaust gases such as NOx, CO, and unburned HC is due to the impact of changing ignition timing [19]. In spark ignition engines, the formation of flame, premature burning, and behaviors of emissions are all influenced by ignition timing [20]. The impact of timing for spark on the combustion of an engine fueled with hydrogen was studied by Shi et al. [21]. They discovered that with the advanced spark ignition angle, the brake thermal efficiency initially increases followed by a decline. NOX, HC, and CO emissions also decreased with retarded spark timing. NOX emissions increase as the fraction of hydrogen volume increases, whereas HC and carbon monoxide are reduced. The impact of adding a high percentage of hydrogen on the performance of engines fueled by hydrogen–gasoline blends was studied by Elsemary et al. [22]. The finding was that at an ignition timing of 30 ◦CA BTDC, the consumption of fuel is reduced, and the thermal efficiency improves [23]. Zhang et al. studied the impact of a spark timing hydrogen/methanol engine on combustion and emissions with the coefficient of excess air at 1.20 and discovered that with increased spark advance angles, after the initial increase, the indicated thermal efficiency begins to decline. Along with the more advanced ignition timing, the flame production time lengthens while the flame propagation period shortens. After hydrogen is added, unburned hydrocarbon and carbon monoxide fall [24]. Syed Yousu Fuddin investigated the impact of the spark timing and compression ratio on engine-fueled hydrogen–ethanol. The researchers showed that by increasing the fraction of hydrogen, the thermal efficiency and mean effective pressure improved for a given ignition time. Ignition timing is a critical operating parameter that affects the performance and emission characteristics of SI engines. However, little work has been reported addressing the impact of ignition timing on the proportion of residual gas and efficient release energy for motorcycles. Thus, it is necessary to study the impact of ignition timings on V-twin engine performance and emissions characteristics, as well as to assess the optimal ignition timing for maximum brake torque and minimum BSFC.

2. Experiment Setup and Material 2.1. Experiment Setup Schematic representation of setup for experiments and the engine testing system are depicted in Figures1 and2, respectively. There are 22 components to this experiment system. The torque resistance for the experiment was supplied by a controller for dynamo Energies 2021, 14, x FOR PEER REVIEW 3 of 17

2. Experiment Setup and Material 2.1. Experiment Setup Energies 2021, 14, x FOR PEER REVIEW 3 of 17 Schematic representation of setup for experiments and the engine testing system are depicted in Figures 1 and 2, respectively. There are 22 components to this experiment system.2. Experiment The torque Setup resistance and Material for the experiment was supplied by a controller for dynamo testing2.1. Experiment systems Setup (1) and a dynamometer from the AVL company (MCA325MO2) (2). En- Energies 2021, 14, 4523 3 of 18 gineSchematic and dynamometer representation are of connectedsetup for expe byriments couplin andg the(3). engine An encoder testing system autonics are (6) and sensor fordepicted temperature in Figures (7)1 and were 2, re mountedspectively. Thereon the are flywh 22 componentseel (5) and to thethis experimentengine (4), sys- respectively. The fueltem. travelsThe torque to theresistance fuel filter for the (10) experime from thent was fuel supplied tank (8) by and a controller then to forthe dynamo fuel pump (9), where testingittesting is injected systems systems (1)into (1) and and the a dynamometera fuel dynamometer injectors from from(11). the the The AVL AVL companybox company to filter (MCA325MO2) (MCA325MO2) the air (15) (2). and Engine(2). En-the throttle allow and dynamometer are connected by coupling (3). An encoder autonics (6) and sensor for airgine to and enter dynamometer the intake are pipe connected (16). byThe couplin air heatg (3).er An regulated encoder autonics the temperature (6) and sensor of the intake air temperaturefor temperature (7) were (7) were mounted mounted on the on flywheel the flywh (5)eel and (5) theand engine the engine (4), respectively. (4), respectively. The fuel The travels(17).fuel travels The to the analyzerto fuel the filterfuel filterfor (10) exhaust from(10) from the fuelgasesthe fuel tank and tank (8) and tran(8) and thensducer then to the to for the fuel cylinder fuel pump pump (9), pressure (9), where where it (19), ECU (20), isdatait injectedis injected collection into into the the(21), fuel fuel injectorsand injectors computer (11). (11). The The (22) box box receiv to to filter filteres the thesignals air air (15) (15) from and and the thethe throttle throttlesensor allow allowthat detects oxygen airqualityair toto enterenter (12), thethe intake sensorintake pipe pipe for (16). (16).temperature TheThe airair heaterheat ofer regulatedexhaustregulated the gasthetemperature temperature (13), sensor of of the forthe intake intakecylinder air air pressure (14), (17).angle(17). TheThe of analyzeranalyzer the throttle, forfor exhaustexhaust and gasessensorgases andand for transducertran measuringsducer forfor cylindercylinderair mass pressurepressure (18). (19),(19), ECUECU (20),(20), datadata collection collection (21), (21), and and computer computer (22) (22) receives receives signals signals from from the the sensor sensor that that detects detects oxygen oxygen qualityquality (12),(12), sensorsensor forfor temperaturetemperature ofof exhaustexhaust gasgas (13),(13), sensorsensor forfor cylindercylinder pressurepressure (14),(14), angleangle ofof thethe throttle, throttle, and and sensor sensor for for measuring measuring air air mass mass (18). (18).

Figure 1. The structure of the experiment setup. FigureFigure 1. 1.The The structure structure of the of experiment the experiment setup. setup.

Figure 2. The experiment engine setup system

The engine used in the experiments is a V-twin engine (4-strokes, four intake valves, FigureFigureand four 2. 2.The exhaustThe experiment experiment valves). engine The engine setup engine system. setup is inst systemalled for the motorcycle. This engine has a system of injection on the intake manifold and a piston with a flat top. Two camshafts are The engine used in the experiments is a V-twin engine (4-strokes, four intake valves, used in the experimental engine, one for inlet valve control and one for exhaust valve and fourThe exhaust engine valves). used in The the engine experiments is installed is for a theV-tw motorcycle.in engine This (4-strokes, engine has four a intake valves, control. The conditions of experiment were defined with a volume fraction of 11.8:1.0, the systemand four of injection exhaust on valves). the intake The manifold engine and is a inst pistonalled with for a flat the top. motorcycle. Two camshafts This engine has a air-to-fuel coefficient is 13.6, and the temperature of the testing system surroundings are used in the experimental engine, one for inlet valve control and one for exhaust valve systemshould be of between injection 29.5 on and the 30 intakedegrees. manifold Air is used and as a acoolant; piston the with lubrication a flat top. tempera- Two camshafts are control. The conditions of experiment were defined with a volume fraction of 11.8:1.0, the usedture was in keptthe atexperimental 80 degrees. The engine, fuel supply one system for inlet holds valve the fuel control pressure and between one for333 exhaust valve air-to-fuelcontrol. coefficientThe conditions is 13.6, and of experiment the temperature were of the def testingined system with surroundingsa volume fraction should of 11.8:1.0, the be between 29.5 and 30 degrees. Air is used as a coolant; the lubrication temperature was keptair-to-fuel at 80 degrees. coefficient The fuel is supply 13.6, systemand the holds temp the fuelerature pressure of the between testing 333 kPasystem and surroundings 363should kPa. Asbe the between engine was 29.5 running, and 30 the degrees. throttle opening Air is used angle wasas a set coolant; to completely the lubrication open. tempera- Underture was different kept speeds at 80 ofdegrees. the engine, The the fuel state supply of the experiment system holds is steady thewith fuel constant pressure between 333 spark advance. Before the analysis, all of the equipment was calibrated. The specifications of the engine are provided in Table1

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Table 1. The engine speciﬁcations.

Parameter Unit Value Engine model - Four stroke, Spark ignition Number of cylinder - 2(V-Twin) Compression ratio - 11.8:1 Bore mm 57 Stroke mm 53.8 Connecting rod mm 107.9 Intake valve - 2 Exhaust valve - 2 Cooling system - Air cooled

2.2. Simulation Model Setup It would be very difficult to calculate residual air remaining and the effective release energy in our experiment. Simulation is, for that reason, a very strong and useful technique for addressing this issue. AVL-boost is a simulation program for engine cycles and gas exchange that allows one to create a model of the whole engine by selecting elements from a toolbox and linking them with pipe elements. In the automobile industry, AVL-boost is used to simulate engines accurately. It can model a large variety of types and sizes of engines, such as two-cycle, four-cycle, diesel, and gasoline. The simulation elements reflect the experimental engine’s components in Figure3. The simulation elements are employed to investigate the properties of engine components. For the conditions of the experiment, the element E1 represents that the engine has been established in a steady state or in a transient state. Another factor was that the display screen MNT1 is absent in the engine under consideration. Through this monitor feature, investigators can see desired information including torque, ratio of residual gas, peak fire pressure, and effective release energy. The two factors SB1 and SB2 reflect the borders of the inlet and exhaust tubes, respectively. The factor CL1 in the experiment plays a role as a system air filter. The element TH1 was established from the throttle opening angle, which was completely opened during engine operation. The element restrictions R1, R2, and R3 take on reducing the pressure of the air stream through the intake and exhaust manifold (from 1 to 21). The contiguous elements from J1 to J6 take charge of allocating the gathered airflow in the tube. Some of the main indicators of airflow include velocity, temperature, and mass flow rate; they can be measured directly through MP1 and MP2. Fuel from the injectors I1 and I2 supply the cylinders C1 and C2, respectively. To measure characteristics of heat release, the simulated model of the engine was built using the combustion function by Vibe:

dx a = (m + 1)ym e−a.y(m+1) (1) dα ∆α dQ α − α dx = ; y = 0 Q ∆αc where: Q is total fuel heat input (W); α is the crank angle (deg); α0 is the start of combustion (deg); ∆α0 is combustion duration (deg); m is shaped parameter (–). Energies 2021, 14, 4523 5 of 18

The integral of the Vibe feature provided the mass faction burned χ:

Z dx x = dα = 1 − e−a.y(m+1) (2) dα Woschni’s model was used to determine how much heat was transferred in the combustion chamber: QT = A.qcoe f f (Tc − Tw) (3) where: 2 QT is heat lost to the wall (W/m ); A is the total surface area of the cylinder head, piston, and cylinder (m2); 2 qcoeff is heat transfer coefﬁcient (W/m K); Tc is combustion gas temperature (K); Tw is the wall temperature of the cylinder (K). Engine efﬁciency is calculated through Equation (4):

1 η = 1 − (4) th rγ−1

Cp γ = Cv where: ηth is theoretical engine efficiency; r is the compression ratio; γ is the specific heat ratio (depending on monatomic gas or diatomic gas); Cp is specific heat capacity with constant pressure (J/kg K); Cv is specific heat capacity with constant volume (J/kg K). Equation (4) was used to measure the residual gas fraction:

R V ρ xcp dV cp Vc xSOC = (5) R V ρ dV Vc

Equations (6) and (7) are used to determine the concentration of N2O and the rate of NO production, respectively:

− 9471.6 q = 6 0.6125 T CN2O 1.1802.10 T e CN2 PO2 (6)

h 2i r1 r4 rNO = 2 Cp Ck 1 − λ (7) 1 + λ R1 1 + R2

CNO,act 1 r1 r4 λ = ; R1 = ; R2 = CNO,equ Cp r2 + r3 r5 + r6 Equation (8) is used to calculate CO production:

rCO = CConst (r1 + r2)(1 − θ) (8) C θ = NO,act CNO,equ Equation (9) was used to measure the mass of unburned HC:

Pc Vcrevice M mHC = (9) RTpiston Energies 2021, 14, x FOR PEER REVIEW 6 of 17

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Figure 3. GV-300S simulation engine model.

3. Results and Discussions 3.1. Model Validation

TheFigure aspects 3. GV-300S and simulation robustness engine model. of our research were demonstrated by both theoretical Figure 3. GV-300S simulation engine model. and experimental3. Results and Discussions tests, represented by red and black lines in Figure 4. Calibration and validation3. Results3.1. Model and equipment Validation Discussions (simulated model) help avoid imprecise measurements in estimating ignitionThe aspectstime. The and robustnessfactors that of our are research taken were into demonstrated account in by the both assessment theoretical of simulated 3.1. Modeland experimental Validation tests, represented by red and black lines in Figure4. Calibration and model include values of engine specification such as bore, stroke, path length, and diam- Thevalidation aspects equipment and robustness (simulated model)of our help research avoid imprecise were demonstrated measurements in by estimating both theoretical eterand experimentalofignition intake time. and The tests, exhaust factors represented that pipe, are taken as bywell into red accountas and a list inblack theof assessmentvalues lines in in Figure of experimental simulated 4. Calibration model conditions. and include values of engine speciﬁcation such as bore, stroke, path length, and diameter of validationintake equipment and exhaust pipe,(simulated as well as model) a list of valueshelp avoid in experimental imprecise conditions. measurements in estimating ignition time. The factors that are taken into account in the assessment of simulated model include45 values of engine specification such as bore, stroke, path length, and diameter of intake and exhaust pipe, as well as a list of values in experimental conditions. Experimment - ignition timing Simulation - Ignition timing 40

45 35

Experimment - ignition timing Simulation - Ignition timing 40 30

35 25

Ignition timing (Deg - BTDC) - (Deg timing Ignition 30 20 2000 4000 6000 8000 10000 25 Engine Speed (rpm) Ignition timing (Deg - BTDC) - (Deg timing Ignition

Figure 4. Ignition timing versus engine speed. 20 Figure 4. Ignition2000 timing 4000 vers 6000us engine 8000 speed. 10000 Engine Speed (rpm) Overall, the ignition time values in both investigations (simulation and experiment) areFigure relatively 4. Ignition consistent timing vers withus engine each speed. other (Figure 4), denoting confidence that the result for the index was verified. The result is, moreover, a good sign for the other indicators that Overall, the ignition time values in both investigations (simulation and experiment) will be implemented in the upcoming figures. are relatively consistent with each other (Figure 4), denoting confidence that the result for The air mass flow in the simulated model was adjusted by means of changing major the index was verified. The result is, moreover, a good sign for the other indicators that coefficientswill be implemented known asin theintake upcoming pipe, throttle,figures. intake ports, and exhaust ports. Results from The air mass flow in the simulated model was adjusted by means of changing major coefficients known as intake pipe, throttle, intake ports, and exhaust ports. Results from

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simulation and experimental tests are presented in Figure 5. their values are compatible Energies 2021, 14, x FOR PEER REVIEW 7 of 17 with insignificantOverall, the differences. ignition time The values highest in both different investigations value (simulation is a very and small experiment) number (1.42%) at 6000are rpm, relatively which consistent is acceptable with each when other seen (Figure as4 ),an denoting average confidence value of that the the air result mass for flow during thethe experiment. index was verified. The result is, moreover, a good sign for the other indicators that simulationwill be implementedand experimental in the upcoming tests are figures. presented in Figure 5. their values are compatible The air mass flow in the simulated model was adjusted by means of changing major with coefficientsinsignificant known differences. as intake pipe, The throttle,highest intake different ports, value and exhaust is a very ports. small Results number from (1.42%) at 6000simulation rpm, which and experimental is acceptable tests when are presented seen as inan Figure average5. their value values of arethe compatible air mass flow dur- with24 insignificant differences. Experiment-Airmass The highest different value is a very small number (1.42%) at ing the experiment. Simulation-Airmass 6000 rpm, which is acceptable when seen as an average value of the air mass flow during the22 experiment.

24 Experiment-Airmass 18 Simulation-Airmass

22 16

20 Air-mass flow (g/s) flow Air-mass 14

18 12

16 2000 4000 6000 8000 10000 Air-mass flow (g/s) flow Air-mass 14 Engine Speed (rpm)

Figure 5. Air12 mass flow versus engine speed.

2000 4000 6000 8000 10000 Errors in measurementEngine of Speedthe engine (rpm) brake torque and power were restricted by sup-

port fromFigure the 5. Air accurate mass ﬂow ratio versus between engine speed. air-fuel and engine friction. On the other hand, the mFigure parameter 5. Air massin Equation flow versus (3) isengine used speed.as a key in reducing errors. Figure 6 shows the results for brake torqueErrors in indicator measurement in both of the cases engine (simulation brake torque and and experiment). power were restrictedThe numerical by sim- ulationErrorssupport values in from measurementare the similar accurate with ratio of theexperimental between engine air-fuel brake da and tatorque at engine each and friction. velocity. power On were theInterestingly, other restricted hand, at by 7000 sup- the m parameter in Equation (3) is used as a key in reducing errors. Figure6 shows the rpm,port thefromresults brake the for brake accuratetorque torque reached ratio indicator between the in bothpeak air-fuel cases at 21.8 (simulation and Nm. engine andThe experiment). friction.difference On The between the numerical other theoretical hand, the andm parameter experimentalsimulation in values Equation cases are was similar (3) approximately is withused experimental as a key 0.9%. in reducing data at each errors. velocity. Figure Interestingly, 6 shows at the results for brake7000 rpm,torque the brakeindicator torque in reached both cases the peak (simulation at 21.8 Nm. Theand difference experiment). between The theoretical numerical sim- ulationand values experimental are similar cases was with approximately experimental 0.9%. data at each velocity. Interestingly, at 7000

rpm, the brake torque reached the peak at 21.8 Nm. The difference between theoretical

and experimental24 cases was Simulation approximately 0.9%. Experiment

24 Simulation Experiment

20 Brake Torque (Nm) Torque Brake

2000 4000 6000 8000 10000 Brake Torque (Nm) Torque Brake Engine Speed (rpm)

12 Figure 6. Brake torque versus engine speed. Figure 6. Brake torque versus engine speed. 2000 4000 6000 8000 10000 Engine power results areEngine demonstrated Speed (rpm) in Figure 7. The best engine power was 18.11 kW at 9000 rpm. The phenomenon in both cases was not out of our prediction, with the differencesFigure 6. Brake being torque less versus at each engine speed speed. of the engine and only a 0.59% difference being reported in the indicator as the highest value difference. Engine power results are demonstrated in Figure 7. The best engine power was 18.11 kW at 9000 rpm. The phenomenon in both cases was not out of our prediction, with the differences being less at each speed of the engine and only a 0.59% difference being reported in the indicator as the highest value difference.

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Energies 2021, 14, x FOR PEER REVIEW 8 of 17 Engine power results are demonstrated in Figure7. The best engine power was 18.11 kW at 9000 rpm. The phenomenon in both cases was not out of our prediction, with the differences being less at each speed of the engine and only a 0.59% difference being reported in the indicator as the highest value difference.

20 Simulation Experiment

Power (KW)

4 2000 4000 6000 8000 10000 Engine Speed (rpm)

Figure 7. Engine power versus engine speed. Figure 7. Engine power versus engine speed. By making good use of the many compatibilities, experimental and simulation results have been meticulously formulated and linked together to establish a stable model, making Byit amaking precise method good usefuluse of to the efficiently many predict compatibilities, the engine performance. experimental Furthermore, and simulation results implementing the model makes it possible to learn more about the effect of air residue on have beenthe engine meticulously performance, formulated as well as on engine and emission linked characteristics. together to establish a stable model, making it a precise method useful to efficiently predict the engine performance. Furthermore, implementing3.2. Results the model makes it possible to learn more about the effect of air residue on The results show the effect of ignition timing on residual gas, effective real energy, the engineand parameters performance, such as BSFC,as well IMEP, as BSFC, on engine peak firing emission temperature, characteristics. brake torque, and engine power. 3.2. ResultsFigure 8 shows the impact of ignition time on the residual air indicator. At an engine speed of 8000 rpm, increasing ignition time (from 10 to 45 degrees) results in the residual Theair going results up from show 0.07% the to 0.26%.effect On of the ignition other hand, timi reducingng on engine residual speed atgas, 6000 effective rpm real energy, and parameterscauses the residual such air as rate BSFC, to increase IMEP, by a very BSFC, small peak number firing (about temperature, 1–1.05%). However, brake torque, and at a lower engine speed (4000 rpm), this ratio fluctuated slightly. The minimum ratio was engine0.54% power. under the specific ignition time condition (30 ◦CA). This can be attributed to the Figureaverage 8 and shows higher the speed impact of the piston, of ignition which leads time to a on faster the and residual cleaner moving air indicator. out of At an engine the residual air. From the process, we need to consider the upcoming intake stroke, and it speed predictsof 8000 a reductionrpm, increasing in fresh air flow ignition entering time the combustion (from 10 chamber. to 45 degrees) results in the residual air going Figureup from9 presents 0.07% the to effect 0.26%. of ignition On the timing other on peak hand, firing reducing temperature. engine As the speed at 6000 rpm causesignition the residual timing increases, air rate the to peak increase firing temperature by a very increases. small number It can be explained (about that1–1.05%). However, when ignition timing is increased, with the lateness of the ignition timing, there will be at a lowermore timeengine for the speed mixture (4000 to blend rpm), better. this This rati leadso fluctuated to the mixture slightly. being more The evenly minimum ratio was 0.54% mixed,under thereby the specific leading to aignition complete combustion.time condit Theion maximum (30 °CA). peak firing This temperature can be attributed to the ◦ averagewas and 2993.26 higher K at 45 speedCA of ignitionof the timingpiston, as enginewhich speed leads was to 6000 a faster rpm. and cleaner moving out of the residual air. From the process, we need to consider the upcoming intake stroke, and it predicts a reduction in fresh air flow entering the combustion chamber.

3.75

3.00 4000 6000 2.25 8000

1.50 Residualratiogas (%) 0.75

0.00 10 20 30 40 50 Ignition timing (CA)

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20 Simulation Experiment

Power (KW)

4 2000 4000 6000 8000 10000 Engine Speed (rpm)

Figure 7. Engine power versus engine speed.

By making good use of the many compatibilities, experimental and simulation results have been meticulously formulated and linked together to establish a stable model, making it a precise method useful to efficiently predict the engine performance. Furthermore, implementing the model makes it possible to learn more about the effect of air residue on the engine performance, as well as on engine emission characteristics.

3.2. Results The results show the effect of ignition timing on residual gas, effective real energy, and parameters such as BSFC, IMEP, BSFC, peak firing temperature, brake torque, and engine power. Figure 8 shows the impact of ignition time on the residual air indicator. At an engine speed of 8000 rpm, increasing ignition time (from 10 to 45 degrees) results in the residual air going up from 0.07% to 0.26%. On the other hand, reducing engine speed at 6000 rpm causes the residual air rate to increase by a very small number (about 1–1.05%). However, at a lower engine speed (4000 rpm), this ratio fluctuated slightly. The minimum ratio was 0.54% under the specific ignition time condition (30 °CA). This can be attributed to the average and higher speed of the piston, which leads to a faster and cleaner moving out of the residual air. From the process, we need to consider the upcoming intake stroke, and it predicts a reduction in fresh air flow entering the combustion chamber. Energies 2021, 14, 4523 9 of 18

3.75

3.00 Energies 2021, 14, x FOR PEER REVIEW 9 of 17 4000 6000 2.25 8000 Figure 8. Residual gas ratio versus ignition timing.

1.50 Figure 9 presents the effect of ignition timing on peak firing temperature. As the ig-

nition Residualratiogas (%) timing increases, the peak firing temperature increases. It can be explained that when ignition0.75 timing is increased, with the lateness of the ignition timing, there will be more time for the mixture to blend better. This leads to the mixture being more evenly mixed, 0.00thereby leading to a complete combustion. The maximum peak firing temperature was 2993.26 K at10 45 °CA of 20 ignition 30 timing as 40 engine speed 50 was 6000 rpm. Ignition timing (CA)

Figure 8. Residual gas ratio versus ignition timing.

3000

2800

2600 4000 6000 8000 2400 Peak firing temperature (K)

2200

10 20 30 40 50 Ignition timing (CA)

Figure 9. Peak ﬁring temperature versus ignition timing.

Figure 9.Figures Peak firing 10 and temperature 11 illustrate the versus impact ignition of ignition timing. timing on peak firing pressure and peak pressure rise. The peak firing pressure and peak pressure rise show the same trend withFigures the peak 10 firingand temperature11 illustrate when the increasing impact theof timingignition of ignition. timing The on peakpeak firing firing pressure and peakpressure pressure and peakrise. pressure The peak rise reachedfiring thepressure maximum and when peak the revolutionpressure of rise the engine show the same trend was 6000 rpm and the timing of ignition was 45 ◦CA. The maximum values are 108.36 bar withand the 9.71 peak bar/deg, firing respectively. temperature when increasing the timing of ignition. The peak firing pressure and peak pressure rise reached the maximum when the revolution of the engine was 6000 rpm and the timing of ignition was 45 °CA. The maximum values are 108.36 bar and 9.71 bar/deg, respectively.

105

63 4000 6000 8000 42 Peak firing pressure (bar) pressure firing Peak

10 20 30 40 50 Ignition timing (CA)

Figure 10. Peak firing pressure versus ignition timing.

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Figure 8. Residual gas ratio versus ignition timing.

Figure 9 presents the effect of ignition timing on peak firing temperature. As the ignition timing increases, the peak firing temperature increases. It can be explained that when ignition timing is increased, with the lateness of the ignition timing, there will be more time for the mixture to blend better. This leads to the mixture being more evenly mixed, thereby leading to a complete combustion. The maximum peak firing temperature was 2993.26 K at 45 °CA of ignition timing as engine speed was 6000 rpm.

3000

2800

2600 4000 6000 8000 2400 Peak firing temperature (K)

2200

10 20 30 40 50 Ignition timing (CA)

Figure 9. Peak firing temperature versus ignition timing.

Figures 10 and 11 illustrate the impact of ignition timing on peak firing pressure and peak pressure rise. The peak firing pressure and peak pressure rise show the same trend with the peak firing temperature when increasing the timing of ignition. The peak firing pressure and peak pressure rise reached the maximum when the revolution of the engine was 6000 rpm and the timing of ignition was 45 °CA. The maximum values are 108.36 bar and 9.71 bar/deg, respectively. Energies 2021, 14, 4523 10 of 18

105

63 4000 6000 8000 42 Peak firing pressure (bar) pressure firing Peak

21 Energies 2021, 14, x FOR PEER REVIEW 10 of 17

10 20 30 40 50 Ignition timing (CA)

Figure 10. Peak ﬁring pressure versus ignition timing. Figure 10. Peak firing pressure versus ignition timing.

10 4000 6000 8000 8

Peak pressure rise (bar/deg) rise pressure Peak 2

0 10 20 30 40 50 Ignition timing (CA)

Figure 11. Peak pressure rise versus ignition timing.

Figure 11.The Peak impact pressure of ignition rise timing versus on ignition effective release timing. energy is presented in Figure 12. When the engine revolution speed was 6000 rpm and 8000 rpm, the values of effective release energy trend upward. In particular, when the revolution speed was 6000 rpm, the Theeffective impact release energyof ignition increased timing quickly before on reachingeffective its peak. release Following energy that, a declineis presented in Figure 12. Whenwas the found. engine When revolution the engine speed speed was 4000 was rpm, 6000 the effective rpm releaseand energy8000 decreasedrpm, the values of effective sharply before steadily increasing. The maximum efﬁcient energy release is an important releasemetric energy that cannot trend be ignoredupward. in any In engineering particular, test. Inwh ouren study, the the revolution coefﬁcients change speed was 6000 rpm, the effectivecontinuously release as energy the engine increased speed changes. quickly Two hypotheses before reaching are assumed. its Firstly, peak. the Following that, a decline wasignition found. time occurs When late, resultingthe engine in too shortspeed of awas burning 4000 time, rpm, so the the fuel cannoteffective be release energy decreased sharply before steadily increasing. The maximum efficient energy release is an important metric that cannot be ignored in any engineering test. In our study, the coefficients change continuously as the engine speed changes. Two hypotheses are assumed. Firstly, the ignition time occurs late, resulting in too short of a burning time, so the fuel cannot be completely converted into heat. In contrast, if the burning time is long because of the advanced ignition time, energy in the form of heat will be significantly lost. This explains why the efficient release energy is sharply increased at the first stage and is then controlled by ignition time. Secondly, the combustion reacts in a short time since energy occurs simultaneously at TDC. In order to optimize the thermal energy for efficiency, the total heat release must be concluded in a short time. This leads to the increased frequency of the energy release to be close with the Otto cycle. Moreover, the range of engine velocity plays a role in driving the total energy release “heat” [25]. The best effective release energy at 4000 rpm and 15 °CA of ignition timing was 0.81725 kJ. Additionally, the best effective release energy at 6000 rpm and 8000 rpm was 0.83374 kJ and 0.83407 kJ at 45 °CA and 25 °CA of ignition timing, respectively.

Energies 2021, 14, 4523 11 of 18

completely converted into heat. In contrast, if the burning time is long because of the advanced ignition time, energy in the form of heat will be significantly lost. This explains why the efficient release energy is sharply increased at the first stage and is then controlled by ignition time. Secondly, the combustion reacts in a short time since energy occurs simultaneously at TDC. In order to optimize the thermal energy for efficiency, the total Energies 2021, 14, x FOR PEER REVIEW heat release must be concluded in a short time. This leads to the increased frequency of 11 of 17 the energy release to be close with the Otto cycle. Moreover, the range of engine velocity plays a role in driving the total energy release “heat” [25]. The best effective release energy at 4000 rpm and 15 ◦CA of ignition timing was 0.81725 kJ. Additionally, the best effective release energy at 6000 rpm and 8000 rpm was 0.83374 kJ and 0.83407 kJ at 45 ◦CA and 25 ◦CA of ignition timing, respectively.

0.832

0.806

0.780

0.754 4000 6000

Effective release energy (kJ) Effective 8000

0.728

10 20 30 40 50 Ignition timing (CA) Figure 12. Effective release energy versus ignition timing. Figure 12.The Effective effects of ignition release time energy on BMEP versus and IMEP ignition are presented timing. in Figures 13 and 14, respectively. The phenomenon can be observed under two stages. In the ﬁrst stage, BMEP and IMEP rapidly increase to a clear peak value. In the next stage, the trend reverses steadily.The effects At each of revolution ignition speed, time the optimalon BMEP values and of BMEP IMEP and IMEPare presented during the in Figures 13 and 14, respectively.same time the The ignition phenomenon time reached its can peak be were observed reported. Thisunder can betwo explained stages. by In the first stage, BMEP the phenomenon that the moment of ignition occurs simultaneously with the release and ofIMEP energy. rapidly Some noteworthy increase values to a areclear the peak ignition value. time benchmark In the next (optimal stage, value) the trend reverses steadily. Atobtained each at revolution 9.49 bar on BMEP speed, and 12.09 the bar optimal on IMEP atvalu 4000es rpm—20 of BMEP◦CA. Under and otherIMEP during the same time conditions (6000 rpm—30 ◦CA), the optimal value was 9.87 and 12.82 bar on BMEP and the ignitionIMEP, respectively. time reached Increasing atits 8000 peak rpm were and 30 ◦reported.CA of the ignition This time, can BMEP be explained and by the phenomenon IMEPthat reached the moment 9.62 and 12.12 of bar.ignition This denotes occurs that thesimu ignitionltaneously time clearly with depends the on release of energy. Some these conditions. noteworthy values are the ignition time benchmark (optimal value) obtained at 9.49 bar on BMEP and 12.09 bar on IMEP at 4000 rpm—20 °CA. Under other conditions (6000 rpm—30 °CA), the optimal value was 9.87 and 12.82 bar on BMEP and IMEP, respectively. Increasing at 8000 rpm and 30 °CA of the ignition time, BMEP and IMEP reached 9.62 and 12.12 bar. This denotes that the ignition time clearly depends on these conditions.

7 BMEP (bar) BMEP 4000 6 6000 8000

10 20 30 40 50 Ignition timing (CA)

Figure 13. BMEP versus ignition timing.

Energies 2021, 14, x FOR PEER REVIEW 11 of 17

0.832

0.806

0.780

0.754 4000 6000

Effective release energy (kJ) Effective 8000

0.728

10 20 30 40 50 Ignition timing (CA)

Figure 12. Effective release energy versus ignition timing.

The effects of ignition time on BMEP and IMEP are presented in Figures 13 and 14, respectively. The phenomenon can be observed under two stages. In the first stage, BMEP and IMEP rapidly increase to a clear peak value. In the next stage, the trend reverses steadily. At each revolution speed, the optimal values of BMEP and IMEP during the same time the ignition time reached its peak were reported. This can be explained by the phenomenon that the moment of ignition occurs simultaneously with the release of energy. Some noteworthy values are the ignition time benchmark (optimal value) obtained at 9.49 bar on BMEP and 12.09 bar on IMEP at 4000 rpm—20 °CA. Under other conditions (6000 rpm—30 °CA), the optimal value was 9.87 and 12.82 bar on BMEP and IMEP, respectively. Increasing at 8000 rpm and 30 °CA of the ignition time, BMEP and IMEP reached 9.62 and 12.12 bar. This denotes that the ignition time clearly depends on these conditions. Energies 2021, 14, 4523 12 of 18

7 BMEP (bar) BMEP 4000 6 6000 8000

5 Energies 2021, 14, x FOR PEER REVIEW 12 of 17

10 20 30 40 50 Ignition timing (CA) Figure 13. BMEP versus ignition timing. Figure 13. BMEP versus ignition timing.

10 IMEP (bar)

9 4000 6000 8000 8

10 20 30 40 50 Ignition timing (CA)

Figure 14. IMEP versus ignition timing.

Figure 14. TheIMEP effect versus of ignition ignition time timing. on BSFC is shown in Figure 15. The trend is opposite to that of BMEP and IMEP. During the ﬁrst stage, BSFC went down continuously until it met the Theminimum effect value, of ignition and then thetime trend on reversed. BSFC Oneis shown interesting in phenomenonFigure 15. toThe note trend is that is opposite to the trends of BMEP and BSFC run in opposite directions [26]. At each engine speed, a set that ofof BMEP numbers and including IMEP. ignition During time the and first fuel consumptionstage, BSFC is reported.went down The minimal continuously value until it met the minimumof BSFC was value, reported and with then 401.89 the g/kWH trend atrevers 4000 rpmed. andOne 20 interesting◦CA. At the same phenomenon speed to note is ◦ that the(6000 trends rpm) andof BMEP increasing and at 25 BSFCCA of run the ignitionin oppo time,site BSFC directions reached 343.82 [26]. g/kWh.At each If engine speed, the set of numbers is increased to 8000 rpm and 35 ◦CA, the BSFC value is 371.69 g/kWh. a set of numbers including ignition time and fuel consumption is reported. The minimal value of BSFC was reported with 401.89 g/kWH at 4000 rpm and 20 °CA. At the same speed (6000 rpm) and increasing at 25 °CA of the ignition time, BSFC reached 343.82 g/kWh. If the set of numbers is increased to 8000 rpm and 35 °CA, the BSFC value is 371.69 g/kWh.

750

4000 675 6000 8000

600

525 BSFC (g/kWh) BSFC 450

375

10 20 30 40 50 Ignition timing (CA)

Figure 15. BSFC versus ignition timing.

Figures 16 and 17 also highlight the relationship of ignition time on power and braking torque, respectively. When the ignition time had an increase on BMEP, the trend of engine brake torque was consistent with the engine power trend. Their trends clearly divide into two stages. In the early stage, they rose steadily until reaching their optimal

Energies 2021, 14, x FOR PEER REVIEW 12 of 17

10 IMEP (bar)

9 4000 6000 8000 8

10 20 30 40 50 Ignition timing (CA)

Figure 14. IMEP versus ignition timing.

The effect of ignition time on BSFC is shown in Figure 15. The trend is opposite to that of BMEP and IMEP. During the first stage, BSFC went down continuously until it met the minimum value, and then the trend reversed. One interesting phenomenon to note is that the trends of BMEP and BSFC run in opposite directions [26]. At each engine speed, a set of numbers including ignition time and fuel consumption is reported. The minimal value of BSFC was reported with 401.89 g/kWH at 4000 rpm and 20 °CA. At the same speed (6000 rpm) and increasing at 25 °CA of the ignition time, BSFC reached 343.82 g/kWh. If the set of numbers is increased to 8000 rpm and 35 °CA, the BSFC value is 371.69 g/kWh. Energies 2021, 14, 4523 13 of 18

750

4000 675 6000 8000

600

525 BSFC (g/kWh) BSFC 450

375

Energies 2021, 14, x FOR PEER REVIEW 10 20 30 40 50 13 of 17 Ignition timing (CA)

Figure 15. BSFC versus ignition timing. Figure 15. BSFC versus ignition timing. values. TheFigures second 16 and stage 17 also can highlight be observed the relationship with of reverse ignition timetrends on power that andoccurred braking at the same ignitiontorque, time, respectively. while the When ignition the ignition time timecontinuously had an increase went on BMEP,up. Furthermore, the trend of engine at each revolu- brakeFigures torque 16 wasand consistent 17 also withhighlight the engine the powerrelationsh trend.ip Their of trendsignition clearly time divide on power into and brak- tion speed, the specific values of the combustion time and optimal performance are pre- ing torque,two stages. respectively. In the early stage, When they the rose ignition steadily untiltime reaching had an their increase optimal on values. BMEP, The the trend of sented. At 4000 rpm and 20 °CA, the maximum engine braking torque was 20.74 Nm, enginesecond brake stage torque can be observedwas consistent with reverse with trends the thatengine occurred power at the trend. same ignition Their time,trends clearly di- while whilethe maximum the ignition timepower continuously was 8.96 went kW. up. At Furthermore, 6000 rpm, at eachtheir revolution values speed,were the21.57 Nm and vide into two stages. In the early stage, they rose steadily until reaching their optimal 13.55 speciﬁckW under values 30 of the°CA combustion ignition time time, and optimalrespectively. performance At arehigher presented. speed At 4000(8000 rpm rpm) and 35 and 20 ◦CA, the maximum engine braking torque was 20.74 Nm, while the maximum °CA, 21.01power Nm was 8.96and kW. 17.6 At 6000kW rpm,were their respectively values were reported. 21.57 Nm and 13.55 kW under 30 ◦CA ignition time, respectively. At higher speed (8000 rpm) and 35 ◦CA, 21.01 Nm and 17.6 kW were respectively reported.

Brake Torque (Nm) Torque Brake 4000 6000 12 8000

10 20 30 40 50 Ignition timing (CA) Figure 16. Brake torque versus ignition timing. Figure 16. Brake torque versus ignition timing.

15 4000 6000 8000

Engine power(kW) Engine 9

10 20 30 40 50 Ignition timing (CA)

Figure 17. Engine power versus ignition timing.

Figure 18 presents the impact of ignition timing on nitrogen oxide emission. The nitrogen oxide emission increases with the advanced ignition timing owing to the increased peak firing temperature. The nitrogen oxides are more affected by the peak firing temperature. The retarded ignition timing is said to decline nitrogen oxide emissions by lowering the peak firing temperature, and the reverse also holds.

Energies 2021, 14, x FOR PEER REVIEW 13 of 17

values. The second stage can be observed with reverse trends that occurred at the same ignition time, while the ignition time continuously went up. Furthermore, at each revolution speed, the specific values of the combustion time and optimal performance are presented. At 4000 rpm and 20 °CA, the maximum engine braking torque was 20.74 Nm, while the maximum power was 8.96 kW. At 6000 rpm, their values were 21.57 Nm and 13.55 kW under 30 °CA ignition time, respectively. At higher speed (8000 rpm) and 35 °CA, 21.01 Nm and 17.6 kW were respectively reported.

Brake Torque (Nm) Torque Brake 4000 6000 12 8000

10 20 30 40 50 Ignition timing (CA)

Figure 16. Brake torque versus ignition timing. Energies 2021, 14, 4523 14 of 18

15 4000 6000 8000

Engine power(kW) Engine 9

10 20 30 40 50 Ignition timing (CA)

Figure 17. Engine power versus ignition timing.

Energies 2021, 14, x FOR PEER REVIEWFigure 17. FigureEngine 18 powerpresents versus the impact ignition of ignition timing. timing on nitrogen oxide emission. The14 nitro- of 17 gen oxide emission increases with the advanced ignition timing owing to the increased Figurepeak firing 18 temperature.presents the The impact nitrogen oxidesof ignition are more timing affected on by thenitrogen peak firing oxide tempera- emission. The ni- ture. The retarded ignition timing is said to decline nitrogen oxide emissions by lowering trogenthe oxide peak firingemission temperature, increases and the with reverse the also advanced holds. ignition timing owing to the increased peak firing temperature. The nitrogen oxides are more affected by the peak firing temperature. The28 retarded ignition timing is said to decline nitrogen oxide emissions by lowering the peak firing temperature, and the reverse also holds. 4000 21 6000 8000

14 NOx (g/kWh) NOx

0 10 20 30 40 50 Ignition timing (CA)

Figure 18. NOx emissions versus ignition timing.

Figure 18. NOxFigure emissions 19 presents versus the ignition effect of timing. ignition timing on the carbon monoxide emission. At 4000 rpm, the CO emission declined, while the ignition timing increased from 10 to 35 ◦CA Figureof ignition 19 presents timing. the After effect this, of the ignition CO emission timing increases. on the carbon At 6000 monoxide rpm and 8000 emission. rpm, the At 4000 rpm,CO emissionthe CO trendemission ﬂuctuates declined, when increasingwhile the ignitionignition timing timing from increased 10 to 45 ◦ CAfrom of ignition10 to 35 °CA oftiming. ignition This timing. denotes After that this, the the residual CO emission gas plays increases. a key role At in the6000 characteristics rpm and 8000 of rpm, CO the COemission. emission In trend particular, fluctuates when thewhen residual increa gassing proportion ignition increases, timing from it leads 10 to to a reduction45 °CA of ignition timing. This denotes that the residual gas plays a key role in the characteristics of CO emission. In particular, when the residual gas proportion increases, it leads to a reduction of the fresh air on the next charge. Hence, the growth of the residual gas directly increases CO emission.

500

4000 400 6000 8000

300

200 CO (g/kWh) CO

100

0 10 20 30 40 50 Ignition timing (CA)

Figure 19. CO emissions versus ignition timing.

Figure 20 presents the impact of ignition timing on HC emission. In this research, HC emission increases as the ignition timing value increases because with the delayed ignition

Energies 2021, 14, x FOR PEER REVIEW 14 of 17

4000 21 6000 8000

14 NOx (g/kWh) NOx

0 10 20 30 40 50 Ignition timing (CA)

Figure 18. NOx emissions versus ignition timing.

Figure 19 presents the effect of ignition timing on the carbon monoxide emission. At 4000 rpm, the CO emission declined, while the ignition timing increased from 10 to 35 °CA of ignition timing. After this, the CO emission increases. At 6000 rpm and 8000 rpm, the CO emission trend fluctuates when increasing ignition timing from 10 to 45 °CA of ignition timing. This denotes that the residual gas plays a key role in the characteristics of CO emission. In particular, when the residual gas proportion increases, it leads to a re- Energies 2021, 14, 4523 15 of 18 duction of the fresh air on the next charge. Hence, the growth of the residual gas directly increases CO emission. of the fresh air on the next charge. Hence, the growth of the residual gas directly increases CO emission.

500

4000 400 6000 8000

300

200 CO (g/kWh) CO

100

Energies 2021, 14, x FOR PEER REVIEW 15 of 17 0 10 20 30 40 50 Ignition timing (CA)

timing,Figure the 19.mixtureCO emissions will versus be more ignition evenly timing. mixed, thereby leading to a more complete combustion. Figure 19. FigureCO emissions 20 presents versus the impact ignition of ignition timing. timing on HC emission. In this research, HC emission increases as the ignition timing value increases because with the delayed ignition Figuretiming, the 20 mixture presents will bethe more impact evenly mixed,of ignition thereby leadingtiming to on a more HCcomplete emission. combustion In this. research, HC

emission 5increases as the ignition timing value increases because with the delayed ignition

2 HC (g/kWh) 4000 1 6000 8000

0 10 20 30 40 50 Ignition timing (CA)

Figure 20. HC emissions versus ignition timing. Figure 20. HC emissions versus ignition timing.

4. Conclusions This paper presents an approach using a combination of experimentation and simulation to determine the ignition timing’s impact on performance and exhaust characteristics of V-twin engines. The effect of ignition timing was completely presented in a V-twin engine. With an engine testing speed band ranging from 3000 rpm to 10,000 rpm, the ignition timing increased from a 10-degree to 45-degree crank angle. Through our approach, the disadvantages of the experimental method have been resolved, providing a reliable method for determining effective energy and residual gas. Some of the conclusions are summarized as follows: (1) This paper denotes a major impact on residual gas ratio, effective release energy, performance, and characteristics of emissions resulting from ignition timing. The minimal residual gas proportion was 0.07% with a revolution speed of 8000 rpm and ignition timing of 10 °CA. Moreover, at 15 °CA of ignition timing, the highest effective release energy was 0.817 kJ at 4000 rpm, while at 8000 rpm, it was 0.8305 kJ at 25 °CA. At 6000 rpm, the highest braking torque was 21.57 Nm, while the minimal BSFC was 343.821 g/kWh. (2) The peak firing temperature and peak pressure rise increase until achieving a maximum value when ignition timing increases. The peak firing pressure and peak pressure rise reached a maximum with a revolution of the engine at 6000 rpm and timing of ignition at 45 °CA. The maximum values are 108.36 bar and 9.71 bar/deg, respectively. (3) At various speeds of the engine, an optimum ignition timing value exists, and the engine will provide the best performance at that value. The optimum ignition timing was found to be 20 °CA at 4000 rpm in this analysis. The optimum ignition timing is 30 °CA and 35 °CA at 6000 rpm and 8000 rpm, respectively. (4) BMEP and IMEP rapidly increase to a peak. In the next stage, the trend reverses steadily. At each revolution speed, the optimal values of BMEP and IMEP at the same time as the ignition timing were reported. (5) The residual gases play a key role in the characteristics of carbon monoxide emission. With increased residual gas, a reduction of the fresh air on the next charge will result.

Energies 2021, 14, 4523 16 of 18

4. Conclusions This paper presents an approach using a combination of experimentation and simulation to determine the ignition timing’s impact on performance and exhaust characteristics of V-twin engines. The effect of ignition timing was completely presented in a V-twin engine. With an engine testing speed band ranging from 3000 rpm to 10,000 rpm, the ignition timing increased from a 10-degree to 45-degree crank angle. Through our approach, the disadvantages of the experimental method have been resolved, providing a reliable method for determining effective energy and residual gas. Some of the conclusions are summarized as follows: (1) This paper denotes a major impact on residual gas ratio, effective release energy, performance, and characteristics of emissions resulting from ignition timing. The minimal residual gas proportion was 0.07% with a revolution speed of 8000 rpm and ignition timing of 10 ◦CA. Moreover, at 15 ◦CA of ignition timing, the highest effective release energy was 0.817 kJ at 4000 rpm, while at 8000 rpm, it was 0.8305 kJ at 25 ◦CA. At 6000 rpm, the highest braking torque was 21.57 Nm, while the minimal BSFC was 343.821 g/kWh. (2) The peak ﬁring temperature and peak pressure rise increase until achieving a maximum value when ignition timing increases. The peak ﬁring pressure and peak pressure rise reached a maximum with a revolution of the engine at 6000 rpm and timing of ignition at 45 ◦CA. The maximum values are 108.36 bar and 9.71 bar/deg, respectively. (3) At various speeds of the engine, an optimum ignition timing value exists, and the engine will provide the best performance at that value. The optimum ignition timing was found to be 20 ◦CA at 4000 rpm in this analysis. The optimum ignition timing is 30 ◦CA and 35 ◦CA at 6000 rpm and 8000 rpm, respectively. (4) BMEP and IMEP rapidly increase to a peak. In the next stage, the trend reverses steadily. At each revolution speed, the optimal values of BMEP and IMEP at the same time as the ignition timing were reported. (5) The residual gases play a key role in the characteristics of carbon monoxide emission. With increased residual gas, a reduction of the fresh air on the next charge will result. Hence, the growth of the residual gases directly increases carbon monoxide emission. The nitrogen oxide emission and HC emission increase with the advanced ignition timing.

Author Contributions: Conceptualization, Q.-N.Y. and N.-X.K.; methodology, Q.-N.Y. and N.-X.K.; software, Q.-N.Y.; validation, Q.-N.Y. and N.-X.K. formal analysis, Q.-N.Y.; investigation, N.-X.K.; resources, Q.-N.Y.; data curation, Q.-N.Y.; writing—original draft preparation, Q.-N.Y.; writing— review and editing, Q.-N.Y. and N.-X.K.; supervision, O.L.; project administration, O.L. All authors have read and agreed to the published version of the manuscript. Funding: This research is financially supported by the individual basic research project by the National Research Foundation of Korea (NRF-2021R1F1A1048238, Reliability Improvement of Am- monia-Diesel Dual-Fuel Combustion Model regarding Optimized Combustion Strategy for Im- proved Combustion Efficiency and Emission Characteristics). This research is financially supported by the Shipbuilding and Offshore Industry Core Technology Development Business by the Ministry of Trade, Industry and Energy (MOTIE, Korea) [Develop-ment of Low Print Point Alternative Fuel Injection System for Small and Medium Vessel Engines for Ships Hazardous Emission Reduce]. (20013146). Institutional Review Board Statement: Not applicable. Informed Consent Statement: Not applicable. Data Availability Statement: Not applicable. Conflicts of Interest: The authors declare no conflict of interest. Energies 2021, 14, 4523 17 of 18

Nomenclature

TDC Top dead center A Total surface area of cylinder head, piston and cylinder (m2) 2 BTDC Before top dead center qcoeff Heat transfer coefficient (W/m K) ATDC After top dead center K Ratio of specific heats (-) 3 BSFC Brake specific fuel consumption (g/kWh) VD Displacement volume (m ) BMEP Brake mean effective pressure (bar) Teff Engine effective torque (Nm) IMEP Indicated mean effective pressure (bar) n Engine speed (rpm) FMEP Friction mean effective pressure (bar) mair Air mass flow (kg/s) dm SMEP Scavenging mean effective pressure (bar) dt The air mass flow rate HC Hydrocarbon Tc The combustion gas temperature (K) CA Crank angle, deg Tw The wall temperature of the cylinder (K) SI-engine Spark ignition engine Aeff The effective flow area (-) 2 CI-engine Compression ignition engine QT Heat lost to the wall (W/m ) m Shape parameter (-) a Vibe parameter Qh Total fuel input (W)

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