Study of the Miller Cycle on a Turbocharged DI Gasoline Engine Regarding Fuel Economy Improvement at Part Load

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Article Study of the Miller Cycle on a Turbocharged DI Gasoline Engine Regarding Fuel Economy Improvement at Part Load

Xuewei Pan 1,2, Yinghua Zhao 1,*, Diming Lou 1,* and Liang Fang 1

1 College of Automotive Studies, Tongji University, Shanghai 200092, China; [email protected] (X.P.); [email protected] (L.F.) 2 SAIC Maxus Automotive Co., Ltd., Shanghai 200438, China * Correspondence: [email protected] (Y.Z.); [email protected] (D.L.)

Received: 17 February 2020; Accepted: 18 March 2020; Published: 22 March 2020

Abstract: This contribution is focused on the fuel economy improvement of the Miller cycle under part-load characteristics on a supercharged DI (Direct Injection) gasoline engine. Firstly, based on the engine bench test, the effects with the Miller cycle application under 3000 rpm were studied. The results show that the Miller cycle has different extents of improvement on pumping loss, combustion and friction loss. For low, medium and high loads, the brake thermal efficiency of the baseline engine is increased by 2.8%, 2.5% and 2.6%, respectively. Besides, the baseline variable valve timing (VVT) is optimized by the test. Subsequently, the 1D CFD (Computational Fluid Dynamics) model of the Miller cycle engine after the test optimization at the working condition of 3000 rpm and BMEP (Brake Mean Effective Pressure) = 10 bar was established, and the influence of the combined change of intake and exhaust valve timing on Miller cycle was studied by simulation. The results show that as the effect of the Miller cycle deepens, the engine’s knocking tendency decreases, so the ignition timing can be further advanced, and the economy of the engine can be improved. Compared with the brake thermal efficiency of the baseline engine, the final result after simulation optimization is increased from 34.6% to 35.6%, which is an improvement of 2.9%.

Keywords: Miller cycle; gasoline engine; eﬃciency; experiment; simulation

1. Introduction As one of the main means of transportation in people’s lives, cars have become an indispensable part of modern life. The dramatic increase in the number of cars poses a huge challenge to the environment and energy, causing increasingly stringent fuel consumption regulations. Facing the pressure of regulations, improving fuel economy has become the top priority of internal combustion engine development. When operating in urban conditions (generally 2000–3000 rpm, medium and low load), the traditional Otto cycle gasoline engine has a higher fuel consumption due to larger pumping loss. Nowadays, gasoline engines have a growing trend toward downsizing and lightweighting, so the displacement is reduced. To ensure that the engine has sufficient output torque at high load, turbocharging technology is often used, which increases the tendency of knock during the high-load operation. To solve this problem, a common solution is to delay the ignition advance angle; another way is to simply reduce the geometric compression ratio of the engine, but these strategic or structural changes make it difficult for the engine to achieve optimal thermal efficiency, which in turn sacrifices the fuel economy of the engine [1]. By either early or late intake valve closing, the Miller cycle can decouple the expansion ratio and the compression ratio of the engine so that the effective compression is significantly lower than the

Energies 2020, 13, 1500; doi:10.3390/en13061500 www.mdpi.com/journal/energies Energies 2020, 13, 1500 2 of 26 geometric compression ratio [2,3]. In 1998, Porsche developed the VarioCam Plus system that combines the best features of shifting the timing phases and changes of cam profile [4]. This system applied in the 911 Turbo uses a hydraulic camshaft phaser with helical teeth instead of changing the chain length. The BMW Valvetronic, well known from the four- and eight-cylinder engines, facilitates low throttle losses as well as efficient partial load control [5], which pioneered this trend and thus put Miller cycle in wider production in 2001 [6,7]. Starting with a 1.6L, four-cylinder engine, VW made an implementation of advanced variable valve timing (VVT) on intake and exhaust to enable effective reduction in compression, while also increasing the exhaust temperature under low-load operation [8]. Besides, the VTEC system (variable valve timing and stroke) of Honda and the Nissan NVCS system of timing changes also the performance well on the engines in production [9]. Many scholars have made some contributions to the research of Miller cycle: Kovács from Braunschweig University of Technology has made a profound study on the potential of Miller cycle in diesel engines [10]; Neher from Karlsruhe University conducted an in-depth research on a naturally aspirated cogeneration lean-burn engine under an external characteristic condition [11,12]; Czech Technical University in Prague made a detailed study on Miller cycle with mixture heating as a means for the fuel economy improvement [13]; Gottschalk conducted a thorough experimental work on the knocking of the Miller cycle engine [14]. Besides, some research over valve train actuation [15] combined with high compression ratio [16] also made remarkable contributions to the valve train techniques. Up to now, the research on the influence of the combined change of intake and exhaust valve timing on the Miller cycle is still insufficient. Besides, the impact on the fuel economy after using the Miller cycle for a gasoline engine equipped with turbocharger also needs further study. For these purposes, based on a supercharged DI gasoline engine, this paper begins with a description of the investigation approach followed by an overview of the experimental setup and test procedure. After that, the results of effects with Miller cycle application and performance improvement are discussed. Finally, the impact of the combined change of intake and exhaust valve timing on the Miller cycle is analyzed by simulation, and thus accomplishes brake thermal efficiency improvement of the baseline engine by experimental and numerical optimization.

2. Experimental Study The experimental test is carried out to achieve an over-expansion cycle of the engine by changing the VVT. The point of operation is selected at speed 3000 rpm, and the BMEP 5, 10, and 15 bar, respectively, to cover the low, medium and high loads of the engine.

2.1. Equipment The research engine is a gasoline engine, equipped with a 350 bar high-pressure fuel injection system and a turbocharger. The speciﬁcations are shown in Table1.

Table 1. Speciﬁcations of the research engine.

Displacement 1.5L Stroke 86.6 mm Bore 74 mm Number of cylinders 4 Rated power 124 kW Rated speed 6000 rpm Maximum torque 250 N m · Speed at maximum torque 1700–4300 rpm BSFC (Brake Speciﬁc Fuel 240 g/kWh Consumption) under rated condition Compression ratio 11.5:1

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To ensure the accuracy and rationality of the data collected by the test, the equipment and measurementTo ensure were the thoroughly accuracy and determi rationalityned. The of layout the data of collectedthe engine by bench the test, test theis shown equipment in Figure and 1.measurement were thoroughly determined. The layout of the engine bench test is shown in Figure1.

FigureFigure 1. 1. TheThe layout ofof the the engine engine bench bench test test (The (The EWG EW inG the in ﬁgure the figure is the abbreviationis the abbreviation of the electrical of the electricalwaste gate). waste gate).

DuringDuring the the data data collection, collection, data data such such as as power, power, torque, torque, and and fuel fuel consumption consumption during during the the test test areare collected collected by by the the engine engine test test bench; bench; the the contro controll parameters parameters during during engine engine operation operation are are read read and and adjustedadjusted by by ETAS-INCA ETAS-INCA (v7.2, (v7.2, ETAS ETAS GmbH, GmbH, Stuttgart, Stuttgart, Germangy); Germangy); the the Kistler Kistler combustion combustion analyzer analyzer (Kistler(Kistler Group, Group, Winterthur,Winterthur, Switzerland) Switzerland) analyzes analyzes the the data data collected collected from from the cylinder the cylinder pressure pressure sensor, sensor,and then and calculates then calculates the combustion the combustion performance performance parameters parameters such as CA50such as during CA50 engine during operation. engine operation.The main equipmentThe main equipment of the test of bench the test and bench measurement and measurement errors are shownerrors are in theshown Table in2 .the Table 2.

TableTable 2. 2. EquipmentEquipment of of the the test test bench. bench.

EquipmentEquipment Purpose Purpose Error Error AVL Dynamometer (AVL List TorqueTorque ＜±0.3N·m< 0.3N m AVL Dynamometer (AVL ± · GmbH, Graz, Austria) Speed < 1r/min List GmbH, Graz, Austria) Speed ＜±1r/min± AVL 735s (AVL List GmbH, AVL 735s (AVL List Fuel flow <0.1kg/h Graz, Austria) Fuel flow ＜0.1kg/h GmbH, Graz, Austria) 2.2. Test Procedure 2.2. Test Procedure The baseline maximum lift of the intake and exhaust valves is 7.5 mm. The intake valve opening The baseline maximum lift of the intake and exhaust valves is 7.5 mm. The intake valve period is 210 CA, and the exhaust valve opening period is 242 CA. Taking the TDC at the end of the opening period◦ is 210 °CA, and the exhaust valve opening period◦ is 242 °CA. Taking the TDC at the exhaust stroke as the reference point, the intake valve closing (IVC) can vary from 151.5 to 223.5 CA, end of the exhaust stroke as the reference point, the intake valve closing (IVC) can vary from 151.5◦ and the exhaust valve closing (EVC) can vary from 1 to 64 CA. to 223.5 °CA, and the exhaust valve closing (EVC) can− vary from◦ −1 to 64 °CA. During the test, the engine is set under the test condition, starting from the baseline working During the test, the engine is set under the test condition, starting from the baseline working VVT. First, ensure that EVC timing is unchanged, and gradually advance the IVC timing to increase VVT. First, ensure that EVC timing is unchanged, and gradually advance the IVC timing to increase intake advance angle so that the Miller cycle effect is deepened. The optimization principle of ignition intake advance angle so that the Miller cycle effect is deepened. The optimization principle of advance angle is to increase ignition advance angle as much as possible under a certain VVT so that the ignition advance angle is to increase ignition advance angle as much as possible under a certain engine has the lowest BSFC (Brake Specific Fuel Consumption) and only a slight knock at the target VVT so that the engine has the lowest BSFC (Brake Specific Fuel Consumption) and only a slight ignition advance angle. According to the experimental experience, under the medium and low load knock at the target ignition advance angle. According to the experimental experience, under the conditions, when CA50 is below 8 CA, the change of the ignition advance angle has little effect on medium and low load conditions, when◦ CA50 is below 8 °CA, the change of the ignition advance combustion. Therefore, under medium and low load, it is reckoned that the ignition advance angle is angle has little effect on combustion. Therefore, under medium and low load, it is reckoned that the adjusted to the optimum when CA50 reaches about 8 CA, while the ignition advance angle is adjusted ignition advance angle is adjusted to the optimum◦ when CA50 reaches about 8 °CA, while the to knock limit under high load. ignition advance angle is adjusted to knock limit under high load. To optimize the throttle angle and EWG (electrical waste gate) opening, the EWG maintains fully To optimize the throttle angle and EWG (electrical waste gate) opening, the EWG maintains open under low load. Only the throttle angle is adjusted to meet the load demand. On the other fully open under low load. Only the throttle angle is adjusted to meet the load demand. On the hand, to achieve the load demand after the VVT change under medium and high loads, the throttle other hand, to achieve the load demand after the VVT change under medium and high loads, the angle is first adjusted, and meanwhile the EWG position changes with the EWG map of the engine. throttle angle is first adjusted, and meanwhile the EWG position changes with the EWG map of the Energies 2020, 13, 1500 4 of 26

engine.When theWhen throttle the isthrottle fully openedis fully andopened the requiredand the loadrequired is still load not is satisﬁed, still not thesatisfied, EWG positionthe EWG is positionmanually is adjusted.manually adjusted. TakingTaking the the point point of of operation BMEP BMEP == 1010 bar bar as as an an example, example, Figure Figure 22 showsshows thethe processprocess ofof adjustingadjusting IVC IVC timing timing to to deepen deepen the the Miller Miller cycle. cycle.

10 Baseline exhaust valve lift 9 Baseline intake valve lift 8 Advanced intake valve lift 7 6 5 4 Advancing IVC timing Valve lift (mm) Valve lift 3

2 Intake advance angle 1 0 -180 -120 -60 0 60 120 180 Crank angle (°CA) FigureFigure 2. Valve liftslifts and and realization realization of theof the Miller Miller cycle cycle (by advancing(by advancing IVC (Intake IVC (Intake Valve Closing)Valve Closing) timing). timing). For comprehensive consideration of the purpose on studying the influence of exhaust retardation angleFor on thecomprehensive engine characteristics consideration under Millerof the cycle, purpose the eff ecton ofstudying deep Miller the cycle influence and the limitationof exhaust of retardationthe VVT range, angle when on the exhaust engine retardation characteristics angle is unde analyzedr Miller at the cycle, points the of effect test operation,of deep Millerintake cycle advance and theangle limitation is fixed underof the deepVVT Millerrange, cycle when position. exhaust Using retard theation method angle above, is analyzed the influence at the points of the exhaustof test operation,retardation intake angle advance on the engine angle characteristicsis fixed under under deep Miller the Miller cycle cycle position. can be Using studied the precisely. method above, Table3 theshows influence the values of the of theexhaust intake retardation advance angle angle of on the the Miller engine cycle characteristics at different points under of operationthe Miller during cycle canthe be test. studied Figure 3precisely. shows the Table process 3 shows of adjusting the values EVC of timing the intake to change advance the exhaustangle of retardation the Miller anglecycle toat differentstudy the points Miller of cycle. operation during the test. Figure 3 shows the process of adjusting EVC timing to change the exhaust retardation angle to study the Miller cycle. Table 3. Values of intake advance angle at different points of operation when exhaust retardation angle Tableis analyzed 3. Values under of Millerintake cycle. advance angle at different points of operation when exhaust retardation angle is analyzed under Miller cycle. Intake Advance Angle Under Baseline Intake Advance Point of Operation Miller Cycle MMiMiller BaselineAngle / CA ◦ Intake Advance AngleCycle / CA Point of Intake ◦ 3000 rpm 5 bar Under Miller Cycle Operation Advance0 25 86.6 mm MMiMiller Cycle/°CA Angle/°CA 3000 rpm 10 bar 3000rpm5bar 0 25 74 mm 0 25 3000 rpm 1586.6 bar mm 0 15 53000rpm10bar 0 25 74 mm 3000rpm15bar 0 15 5

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10 Baseline exhaust valve lift 109 Advanced exhaust valve lift Baseline exhaust valve lift 8 Advanced intake valve lift 9 Advanced exhaust valve lift 78 Advanced intake valve lift 67 56 45 Advancing EVC timing Valve lift (mm) Valve lift 34 Advancing EVC timing Valve lift (mm) Valve lift 23 Exhaust retardation angle 12 Exhaust retardation angle 01 -180 -120 -60 0 60 120 180 0 -180 -120 -60Crank angle 0 (°CA) 60 120 180

FigureFigure 3.3.Method Method of analyzingof analyzing the influencethe influence ofCrank exhaust ofangle retardationex (°CA)haust retardation angle on theangle engine on characteristics the engine characteristics under the Miller cycle (by advancing EVC (Exhaust Valve closing) timing). Figureunder the3. MillerMethod cycle of (by analyzing advancing the EVC influence (Exhaust Valveof exhaust closing) retardation timing). angle on the engine characteristics under the Miller cycle (by advancing EVC (Exhaust Valve closing) timing). 2.3.2.3. Results Results and and Discussion Discussion 2.3. Results and Discussion 2.3.1.2.3.1. Effects Effects with with Miller Miller Cycle Cycle Application Application 2.3.1.ThisThis Effects paper with will Miller elaborateelaborate Cycle theApplicationthe eeffectsffects with with the the Miller Miller cycle cycle application application in threein three aspects, aspects, which which are arepumping pumping loss, loss, combustion combustion and and friction friction loss. loss. ThisTo express paper withoutwill elaborate ambiguity, the effects in this with paper, the the Mi increaseller cycle or application decrease rate in ofthree a parameter aspects, which in the are pumpingTo express loss, without combustion ambiguity, and friction in this paper,loss. the increase or decrease rate of a parameter in the following analysis (such as BSFC, thermal efficiency, etc.) is defined as |A-Abaseline|/Abaseline. followingTo express analysis without (such asambiguity, BSFC, thermal in this efficiency, paper, the etc.) increase is defined or decrease as |A-A ratebaseline of|/A a parameterbaseline. in the • PumpingPumping Loss Loss following• analysis (such as BSFC, thermal efficiency, etc.) is defined as |A-Abaseline|/Abaseline. • EWG position directly affects intake pressure before intercooler. Since the test shows that EWG PumpingEWG position Loss directly affects intake pressure before intercooler. Since the test shows that EWG position and intake pressure before intercooler relative to intake advance angle changes similarly positionEWG and position intake pressuredirectly affects before intercoolerintake pressure relative before to intake intercooler. advance Since angle the changes test shows similarly that underEWG under various points of operation, for the sake of simplicity, the point of operation at 3000 rpm and positionvarious pointsand intake of operation, pressure for before the sake intercooler of simplicity, relative the pointto intake of operation advance atangle 3000 changes rpm and similarly 10 bar is 10 bar is illustrated in detail. This is shown in graphical form in Figure 4 at the point of operation at underillustrated various in detail. points This of operation, is shown infor graphical the sake formof simplicity, in Figure the4 at point the point of operation of operation at 3000 at 3000 rpm rpm and 300010 bar rpm is illustrated and 10 bar. in detail. This is shown in graphical form in Figure 4 at the point of operation at and 10 bar. 100 132 3000 rpm and 10 bar. EWG position 130 10090 Intake pressure before intecooler 132 EWG position 128130 8090 Intake pressure before intecooler 126128 7080 124126 122 6070 124 120122 5060 118120 EWG position (%) EWG position 4050 116118 EWG position (%) EWG position 114116 3040 112114 (kPa) intercooler before pressure Intake 2030 110 112 (kPa) intercooler before pressure Intake 0 5 10 15 20 25 20 110 0 5Intake advance 10 angle 15 (°CA) 20 25 Figure 4. EWG (Electrical waste gate) positionIntake advance and angleintake (°CA) pressure before intercooler versus intake advance angle (point of operation: n = 3000 rpm, BMEP = 10 bar, exhaust retardation angle = 25 °CA, Figure 4. EWG (Electrical (Electrical waste waste gate) gate) position and intake pressure before intercooler versus intake CA50 = 8 °CA). advance angle (point of operation: n = 30003000 rpm, rpm, BM BMEPEP = 1010 bar, bar, exhaust exhaust retardation angle = 2525 °CA,◦CA, CA50 = 88 °CA).◦CA). As shown in Figure 4, intake pressure before intercooler rises dramatically as EWG opening drops.As When shown the in intakeFigure advance4, intake anglepressure is increabeforesed intercooler from 0 (baselinerises dramatically engine) toas 25EWG °CA, opening EWG drops. When the intake advance angle is increased from 0 (baseline engine) to 25 °CA, EWG Energies 2020, 13, 1500 6 of 26

As shown in Figure4, intake pressure before intercooler rises dramatically as EWG opening drops. When the intake advance angle is increased from 0 (baseline engine) to 25 ◦CA, EWG opening is reduced from 91.6% to 29%, and the corresponding intake pressure before intercooler is increased from 112.1 to 129.9 kPa, which can guarantee the required load of Miller cycle engines. On the one hand, since the exhaust gas still contains high energy, the further intervention of the turbocharger is beneficial to the utilization of exhaust energy, thereby improving the thermal efficiency of the engine; on the other hand, the application of turbocharger causes an increase in exhaust backpressure, which, to some extent, mitigates the pumping loss improvement due to the larger throttle angle. In general, attributed to a turbocharger, the output power under the deep Miller cycle is not opening is reduced from 91.6% to 29%, and the corresponding intake pressure before intercooler is worse than that of the baseline engine under medium and high load at 3000 rpm, and making good increased from 112.1 to 129.9 kPa, which can guarantee the required load of Miller cycle engines. use of the exhaust gas has a certain improvement on the thermal efficiency of Miller cycle engines. On the one hand, since the exhaust gas still contains high energy, the further intervention of the turbochargerPumping mean is beneficial effective pressure to the utilizatio (PMEP) isn aof common exhaust parameterenergy, thereby that characterizes improving thethe pumping thermal efficiencyloss in the of engine’s the engine; working on cyclethe other [17]. hand, Figure the5 shows application PMEP relativeof turbocharger to intake causes advance an angle increase under in exhaustbaseline backpressure, EVC, and PMEP which, relative to tosome exhaust extent, retardation mitigates angle the pumping under the loss Miller improvement cycle at different due pointsto the largerof operation throttle during angle. the In test. general, Figure attributed6 illustrates to intake a turbocharger, pressure after the intercooler output power versus under intake the advance deep Millerangle andcycle exhaust is not worse retardation than that angle. of the baseline engine under medium and high load at 3000 rpm, As can be seen in Figure5, due to larger throttle opening, PMEP is increased from 0.44 to 0.29 and making good use of the exhaust gas has a certain improvement on the thermal− efficiency− of Millerbar when cycle intake engines. advance angle changes from 0 (baseline engine) to 25 ◦CA at the point of operation at 3000Pumping rpm and mean 5 bar, effective which conspicuouslypressure (PMEP) reduces is a pumpingcommon lossparameter in the engine’sthat characterizes working cycle; the pumpingwhile under loss medium- in the engine's and high-load working conditions, cycle [17]. theFigure drop 5 ofshows the EWG PMEP opening relative adversely to intake a ffadvanceects the angleexhaust under backpressure baseline EVC, due to and the PMEP increase relative of turbocharging to exhaust retardation effect, which angle weakens under the the improvement Miller cycle at of differentthe throttle points opening of operation on pumping during loss. the As test. a result, Figu PMEPre 6 illustrates under medium- intake pressure and high-load after intercooler conditions versusonly slightly intake improvedadvance angle relative and to exhaust the baseline retardation engine angle. during the Miller cycle deepening process.

BMEP=5bar BMEP=10bar BMEP=15bar -0.20 -0.25 -0.30 -0.35 -0.40 -0.45 0 5 10 15 20 25 Intake advance angle (°CA)

PMEP (bar) PMEP 0.1 0.0 -0.1 -0.2 -0.3 0 5 10 15 20 25 30 Exhaust retardation angle (°CA) Figure 5. PMEPPMEP (Pumping(Pumping Mean Mean E ﬀEffectiveective Pressure) Pressure) versus versus intake intake advance advance angle angle under under baseline baseline EVC, EVC,and PMEP and PMEP versus versus exhaust exhaust retardation retardation angle under angle theunder Miller the cycle Miller at dicycleﬀerent at pointsdifferent of operationpoints of operationduring the during test. the test. Energies 2020, 13, 1500 7 of 26

BMEP=5bar BMEP=10bar BMEP=15bar 200 180 160 140 120 100 0 5 10 15 20 25 Intake advance angle (°CA) 200 180 160 140 Intake pressure after intercooler (bar) 120 100 0 5 10 15 20 25 30 Exhaust retardation angle (°CA) FigureFigure 6. IntakeIntake pressurepressure after after intercooler intercooler versus versus intake intake advance advance angle angle under unde baseliner baseline EVC, andEVC, PMEP and PMEPversus exhaustversus exhaust retardation retardation angle under angle the under Miller cyclethe Miller at diff erentcycle pointsat different of operation points during of operation the test. during the test. The change of exhaust retardation angle also had certain effects on pumping loss under the loadAs characteristics can be seen ofin 3000Figure rpm. 5, due At theto larger point ofthrottle operation opening, at 3000 PMEP rpm is and increased 5 bar, asfrom the − exhaust0.44 to −retardation0.29 bar when angle intake decreases, advance the throttle angle openingchanges anglefrom can0 (baseline be further engine) enlarged. to 25 Consequently, °CA at the whenpoint theof baseline exhaust retardation angle is reduced from 25 to 0 CA, PMEP follows from 0.29 to 0.26 bar, operation at 3000 rpm and 5 bar, which conspicuously◦ reduces pumping loss− in the −engine's workingwhich mitigates cycle; whilepumping under loss. medium- Under medium- and high-l andoad high-load conditions, conditions, the drop due toof thethe faster EWG gas opening flow in adverselythe cylinder, affects the excessivethe exhaust exhaust back retardationpressure due angle to ofthe the increase baseline of engine turbocharging causes the effect, exhaust which back weakenspressure tothe a ffimprovementect PMEP during of the the throttle intake processopening for on a longerpumping time. loss. As As a result, a result, when PMEP the exhaust under medium-retardation and angle high-load is advanced, conditions the pumping only slightly loss ofimpr theoved Miller relative cycle intake to the processbaseline is engine greatly during improved. the MillerFor the cycle high-load deepening condition process. of 3000 rpm with the most obvious improvement, PMEP is improved from 0.28 to 0.06 bar after the baseline exhaust retardation angle 30 CA is adjusted to 5 CA. According to − The change of exhaust retardation angle also had certain◦ effects on pumping◦ loss under the loadthe test, characteristics the improved of 3000 exhaust rpm. retardation At the point angle of evenoperation has a positiveat 3000 rpm effect and on 5 the bar, pumping as the exhaust process. retardationCompared withangle the decreases, results above, the throttle the influence openin ofg the angle throttle can openingbe further and enlarged. the EWG openingConsequently, on the whenMiller the cycle baseline pumping exhaust loss underretard medium-ation angle and is large-loadreduced from conditions 25 to 0 °CA, is relatively PMEP insignificant.follows from −0.29 to −0.26As shownbar, which in Figure mitigates6, the intakepumping pressure loss. afterUnder intercooler medium- remains and high-load low at the conditions, point of operation due to the at faster3000 rpm gas andflow 5 bar;in the under cylinder, medium the andexcessive high load, exha theust intake retardation pressure angle after of intercooler the baseline rises engine quickly causes when thethe exhaust Miller cycle back becomes pressure deeper, to affect which PMEP guarantees during the the intake suffi cientprocess air supply.for a longerMeanwhile, time. As the a result, intake whenpressure the afterexhaust intercooler retardation keeps angle increasing is advanced, with the the decrease pumping of loss the exhaustof the Miller retardation cycle intake angle. process is greatlyCombustion improved. For the high-load condition of 3000 rpm with the most obvious improvement, PMEP• is improved from −0.28 to 0.06 bar after the baseline exhaust retardation angle 30 °CA is adjustedFigure to 75 °CA.shows According the logP versusto the test, logV the diagram improved relative exhaust to intake retardation advance angle angle even under has a baselinepositive effectEVC aton di fftheerent pumping points ofprocess. operation. Compared It should with be notedthe results that, due above, to the the fluctuation influence of of the the test throttle value, openingthe curve and was the smoothed EWG opening to a certain on the extent Miller without cycle decreasing pumping theloss accuracy. under medium- and large-load conditions is relatively insignificant. As shown in Figure 6, the intake pressure after intercooler remains low at the point of operation at 3000 rpm and 5 bar; under medium and high load, the intake pressure after intercooler rises quickly when the Miller cycle becomes deeper, which guarantees the sufficient air supply. Meanwhile, the intake pressure after intercooler keeps increasing with the decrease of the exhaust retardation angle. • Combustion Figure 7 shows the logP versus logV diagram relative to intake advance angle under baseline EVC at different points of operation. It should be noted that, due to the fluctuation of the test value, the curve was smoothed to a certain extent without decreasing the accuracy. Energies 2020, 13, 1500 8 of 26

Intake advance angle=0°CA(3000r/min and 5bar) Intake advance angle=25°CA(3000r/min and 5bar) Intake advance angle=0°CA(3000r/min and 10bar) Intake advance angle=25°CA(3000r/min and 10bar) Intake advance angle=0°CA(3000r/min and 15bar) Intake advance angle=25°CA(3000r/min and 15bar)

Pressure (bar) 1

Over expansion

0.1 1 Volumn/Vmax Figure 7. The logP versus logV diagram relative to intake advance angle under baseline EVC at different Figure 7. The logP versus logV diagram relative to intake advance angle under baseline EVC at points of operation. different points of operation. As shown in Figure7, an obvious e ffect of over expansion occurs when the baseline engine is adjustedAs shown to the in extreme Figure Miller7, an obviou conditions effect at diofff overerent expansion points of operation,occurs when which the baseline confirms engine that the is adjustedapplication to the of the extreme Miller Miller cycle bycondition VVT adjustment at different is points effective. of operation, On this basis, which a further confirms analysis that the on applicationcombustion of can the be Miller performed cycle properly.by VVT adjustment is effective. On this basis, a further analysis on combustionThe change can be of performed intake advance properly. angle has a significant influence on the combustion phase. Figure8 showsThe the change optimal of ignitionintake advance timing, angle CA50 has and a CA10-90significant relative influence to intake on the advance combustion angle phase. under Figurebaseline 8 EVC.shows the optimal ignition timing, CA50 and CA10-90 relative to intake advance angle under baseline EVC. Figure8 first shows the control principle of CA50 during the test: When CA50 is around 8 ◦CA, the ignition timing is considered to be the best. (If knock occurs when CA50 is greater than 8 ◦CA, the minimum ignition timing is set at the knock limitation point). Subsequently, as shown in Figure8, the optimal ignition timing is conspicuously affected by the intake advance angle. As the intake advance angle increases, the effective compression ratio of the Miller cycle decreases, so the knock tendency is suppressed, which could further advance the ignition timing. The load of the engine also has a great influence on the optimal ignition timing: When the intake advance angle is 25 ◦CA, the optimum ignition timings at 3000 rpm and 5 bar, 3000 rpm and 10 bar, and 3000 rpm and 15 bar are at 27.75, 19.5 and 14.25 ◦CA BTDC (Before Top Dead Center), respectively. This is because, as the load rises, more fuel is injected in one working cycle, which augments the possibility of knock. Therefore, as the load rises, the ignition timing must be delayed to suppress the tendency of knock. Energies 2020, 13, 1500 9 of 26

BMEP=5bar BMEP=10bar BMEP=15bar 30 25 20 15 10

0 5 10 15 20 25 Intake advance angle (°CA) Optimal ignition timing (°CA bTDC) timing ignition Optimal 20 18 16 14 12

CA50 (°CA) CA50 10 8 0 5 10 15 20 25 Intake advance angle (°CA)

25 24 23 22 21 CA10-90 (°CA) CA10-90 20 0 5 10 15 20 25 Intake advance angle (°CA) Figure 8. Optimal ignition timing, CA50, and CA10-90 versus intake advance angle under baseline EVC. Figure 8. Optimal ignition timing, CA50, and CA10-90 versus intake advance angle under baseline EVC.The change of CA10-90 relative to intake advance angle indicates the influence of the Miller cycle on combustion duration. It can be found from Figure8 that, at the point of operation at 3000 rpm Figure 8 first shows the control principle of CA50 during the test: When CA50 is around 8 °CA, and 15 bar, due to the ignition timing advance caused by the Miller cycle, combustion duration is the ignition timing is considered to be the best. (If knock occurs when CA50 is greater than 8 °CA, remarkably reduced; the combustion duration under medium load conditions also decreases slightly the minimum ignition timing is set at the knock limitation point). Subsequently, as shown in Figure (about 3.2%) with the increase of intake advance angle. The shortening of the combustion duration also 8, the optimal ignition timing is conspicuously affected by the intake advance angle. As the intake indicates the optimization effect of the Miller cycle on in-cylinder combustion. For low-load conditions, advance angle increases, the effective compression ratio of the Miller cycle decreases, so the knock the Miller cycle has not been found to improve CA10-90 during combustion. tendency is suppressed, which could further advance the ignition timing. The load of the engine Additionally, the test found that the variation of exhaust retardation angle under medium and also has a great influence on the optimal ignition timing: When the intake advance angle is 25 °CA, low load conditions had only little effects on the combustion phase. However, after applying the Miller the optimum ignition timings at 3000 rpm and 5 bar, 3000 rpm and 10 bar, and 3000 rpm and 15 bar cycle, the change in in-cylinder maximum combustion temperature deserves further attention. In this are at 27.75, 19.5 and 14.25 °CA BTDC (Before Top Dead Center), respectively. This is because, as paper, the in-cylinder temperature is calculated as follows. Firstly a calculation of the air density is the load rises, more fuel is injected in one working cycle, which augments the possibility of knock. completed with the help of in-cylinder pressure, specific gas constant and the intake temperature: Therefore, as the load rises, the ignition timing must be delayed to suppress the tendency of knock. The change of CA10-90 relative to intake advancepnorm angle indicates the influence of the Miller ρ0 = (1) cycle on combustion duration. It can be found fromRT Figurein 8 that, at the point of operation at 3000 rpm and 15 bar, due to the ignition timing advance caused by the Miller cycle, combustion duration Then the temperature in the cylinder by the thermal equation of state is calculated: is remarkably reduced; the combustion duration under medium load conditions also decreases slightly (about 3.2%) with the increase of intake advance angle. The shortening of the combustion pin cyclinderVin cyclinder duration also indicates the optimizationT = effect− of the Miller− cycle on in-cylinder combustion. For(2) ρ0V0ve f f R low-load conditions, the Miller cycle has not been found to improve CA10-90 during combustion.

Additionally,In this formula, the V 0testrepresents found that the displacementthe variation ofof theexhaust cylinder retardation while ve ffanglerepresents under the medium volumetric and lowefficiency. load conditions The value 0.9had is only a suggested little effects value on for th thee combustion volumetric ephase.fficiency However, which will after automatically applying the be Millercorrected cycle, by thethe formula change with in thein-cylinder help of the maxi calculatedmum combustion intake pressure. temperature deserves further attention. In this paper, the in-cylinder temperature is calculated as follows. Firstly a calculation of the air density is completed with the help of in-cylinder pressure, specific gas constant and the intake temperature:

pnorm ρ = (1) 0 RT in Then the temperature in the cylinder by the thermal equation of state is calculated:

pin−cyclinderVin−cyclinder T = (2) ρ V v R 0 0 eff Energies 2020, 13, 1500 10 of 26 In this formula, V0 represents the displacement of the cylinder while veff represents the volumetric efficiency. The value 0.9 is a suggested value for the volumetric efficiency which will automaticallyFigure9 illustrates be corrected maximum by the formula combustion with temperaturethe help of the relative calculated to intake intake advance pressure. angle under baselineFigure EVC, 9 illustrates and maximum maximum combustion combustion temperature temper relativeature relative to exhaust to intake retardation advance angle angle under under the baselineMiller cycle EVC, at diandﬀerent maximum points combustion of operation temperatur during thee test. relative to exhaust retardation angle under the Miller cycle at different points of operation during the test.

BMEP=5bar BMEP=10bar BMEP=15bar 2500 2400 2300 2200 2100

0 5 10 15 20 25

Maximum combustion temperature (°C) temperature combustion Maximum Intake advance angle (°CA)

2400

2300

2200

2100

2000 0 5 10 15 20 25 30

Maximum combustion temperature (°C) Exhaust retardation angle (°CA) Figure 9. Maximum combustion temperature versus intake advance angle under baseline EVC, Figure 9. Maximum combustion temperature versus intake advance angle under baseline EVC, and and maximum combustion temperature versus exhaust retardation angle under the Miller cycle at maximum combustion temperature versus exhaust retardation angle under the Miller cycle at different points of operation during the test. different points of operation during the test. It can be seen from Figure9 that the in-cylinder maximum combustion temperature relative It can be seen from Figure 9 that the in-cylinder maximum combustion temperature relative to to intake advance angle varies from loads at 3000 rpm: At the point of operation at 3000 rpm and intake advance angle varies from loads at 3000 rpm: At the point of operation at 3000 rpm and 5 5 bar, the in-cylinder maximum combustion temperature shows a slight downward trend while the bar, the in-cylinder maximum combustion temperature shows a slight downward trend while the in-cylinder maximum combustion temperature shows an opposite trend under medium- and high-load in-cylinder maximum combustion temperature shows an opposite trend under medium- and high- conditions. On the other hand, the in-cylinder maximum combustion temperature relative to exhaust load conditions. On the other hand, the in-cylinder maximum combustion temperature relative to retardation angle under different loads at 3000 rpm shows a much more obvious change: As the exhaust exhaust retardation angle under different loads at 3000 rpm shows a much more obvious change: retardation angle decreases, the in-cylinder maximum combustion temperature drops remarkably. As the exhaust retardation angle decreases, the in-cylinder maximum combustion temperature The dramatic drop occurs under the high load condition at 3000 rpm, and a deep Miller cycle combined drops remarkably. The dramatic drop occurs under the high load condition at 3000 rpm, and a deep with a smaller exhaust retardation angle can reduce the maximum combustion temperature even below Miller cycle combined with a smaller exhaust retardation angle can reduce the maximum the baseline engine under all three test conditions. The reason for this phenomenon is that when the combustion temperature even below the baseline engine under all three test conditions. The reason exhaust retardation angle is reduced, less high-temperature exhaust gas in the intake stroke is returned for this phenomenon is that when the exhaust retardation angle is reduced, less high-temperature to the cylinder, so the temperature of the cylinder gas at the end of the intake stroke is lower, which is exhaust gas in the intake stroke is returned to the cylinder, so the temperature of the cylinder gas at beneficial to the combustion process. the end of the intake stroke is lower, which is beneficial to the combustion process. According to the above analysis, it can be concluded that deep Miller cycle with the appropriate According to the above analysis, it can be concluded that deep Miller cycle with the valve overlap angle can effectively reduce in-cylinder maximum combustion temperature at 3000 rpm: appropriate valve overlap angle can effectively reduce in-cylinder maximum combustion For the most commonly used 3000 rpm and 10 bar condition, the in-cylinder maximum combustion temperature of the baseline engine reaches approximately 2260 ◦C. After applying deep Miller cycle combined with the smaller exhaust retardation angle (30 ◦CA for the intake advance angle and 5 ◦CA for the exhaust retardation angle), the in-cylinder maximum combustion temperature drops to approximately 2240 ◦C, and further reducing the exhaust retardation angle can even lower the maximum combustion temperature to below 2200 ◦C. Energies 2020, 13, 1500 11 of 26 temperature at 3000 rpm: For the most commonly used 3000 rpm and 10 bar condition, the in- cylinderThe maximum reduction combustion of in-cylinder temp maximumerature of combustionthe baseline temperatureengine reaches is approximately a manifestation 2260 of the°C. Afterimprovement applying in deep combustion Miller conditionscycle combined because with a higher the smaller in-cylinder exhaust maximum retardation combustion angle temperature(30 °CA for theproduces intake a advance higher thermal angle and load, 5which °CA for causes the theexha thermalust retardation efficiency angle), deterioration the in-cylinder of the engine; maximum at the combustionsame time, a highertemperature in-cylinder drops maximum to approximately combustion 2240 temperature °C, and doesfurther harm reducing to the lubrication the exhaust and retardationthe durability angle of thecan components; even lower the besides, maximum a higher combustion in-cylinder temperature maximum to combustionbelow 2200 °C. temperature increasesThe thereduction generation of in-cylinder of NOx, which maximum adversely combusti influenceson temperature the engine emission is a manifestation characteristics. of the improvementFriction loss in combustion conditions because a higher in-cylinder maximum combustion temperature• produces a higher thermal load, which causes the thermal efficiency deterioration of the engine;The energy at the lost same in time, one cycle a higher to overcome in-cylinde ther maximum friction also combustion has a significant temperature impact does on harm the fuel to theeconomy. lubrication Friction and mean the effdurabilityective pressure of the (FMEP) components; is a common besides, parameter a higher to characterizein-cylinder themaximum friction combustionloss in the engine temperature working increases cycle. The the main generation influencing of factorNOx, of which FMEP adversely is the average influences speed of the the engine piston emissionduring the characteristics. working cycle and in-cylinder peak pressure. Since the engine speed is always maintained •at 3000Friction rpm during loss the test, the average piston speed does not affect the FMEP. Consequently, only the in-cylinderThe energy peak lost pressure in one determines cycle to overcome the value the of fric FMEP.tion Figuresalso has 10 a andsignificant 11 show impact in-cylinder on the peakfuel economy.pressure and Friction FMEP relativemean effectiv to intakee pressure advance angle(FMEP) under is a baseline common EVC, parameter and in-cylinder to characterize peak pressure the frictionrelative loss to exhaust in the engine retardation working angle cycle. under The the main Miller influencing cycle at difactorfferent of FMEP points is of the operation average during speed ofthe the test. piston during the working cycle and in-cylinder peak pressure. Since the engine speed is alwaysAs maintained shown in Figure at 3000 10 rpm, the during in-cylinder the test, peak the pressure average ispiston remarkably speed does affected not byaffect the the load: FMEP. The Consequently,in-cylinder peak only pressure the in-cylinder increases peak as the pressure load rises, determines which, asthe a value result, of magnifies FMEP. Figures the friction 10 and loss. 11 showAs can in-cylinderbe seen in Figurepeak pressure 11, the fluctuationand FMEP ofrelative FMEP to at 3000intake rpm advance and 10 angle bar seems under to baseline be abnormal, EVC, andbut generallyin-cylinderspeaking, peak pressure the trends relative at different to exha loadsust are retardation consistent angle with theunde relationshipr the Miller between cycle theat differentpeak pressure points and of operation intake advance during angle the test. or exhaust retardation angle.

BMEP=5bar BMEP=10bar BMEP=15bar 90 80 70 60 50 40 30 0 5 10 15 20 25 In-cylinder In-cylinder peak pressure (bar) Intake advance angle (°CA) 80 70 60 50 40 30 0 5 10 15 20 25 30 In-cylinder peak pressure (bar) In-cylinder Exhaust retardation angle (°CA) Figure 10.10. In-cylinderIn-cylinder peak peak pressure pressure versus versus intake intake advance advance angle angle under under baseline baseline EVC, andEVC, in-cylinder and in- cylinderpeak pressure peak pressure versus exhaust versus retardationexhaust retardation angle under angle the under Miller the cycle Miller at di cycleﬀerent at pointsdifferent of operationpoints of operationduring the during test. the test. Energies 2020, 13, 1500 12 of 26

BMEP=5bar BMEP=10bar BMEP=15bar

0.9 0.8 0.7 0.6 FMEP (bar) FMEP 0.5 0.4 0 5 10 15 20 25 Intake advance angle (°CA) 0.8 0.7 0.6 0.5 0.4 FMEP (bar) 0.3 0.2 0 5 10 15 20 25 30 Exhaust retardation angle (°CA) Figure 11. FMEP (Friction Mean Effective Pressure) versus intake advance angle under baseline EVC, Figure 11. FMEP (Friction Mean Effective Pressure) versus intake advance angle under baseline and FMEP versus exhaust retardation angle under the Miller cycle at different points of operation EVC, and FMEP versus exhaust retardation angle under the Miller cycle at different points of during the test. operation during the test. Under high-load conditions, as the Miller cycle is deepened, the in-cylinder peak pressure changes As shown in Figure 10, the in-cylinder peak pressure is remarkably affected by the load: The oppositely from the medium and low load. This is because when the Miller cycle is deepened, the CA50 in-cylinder peak pressure increases as the load rises, which, as a result, magnifies the friction loss. is greatly decreased due to the advance of the ignition timing. Moreover, the in-cylinder combustion As can be seen in Figure 11, the fluctuation of FMEP at 3000 rpm and 10 bar seems to be abnormal, rate is accelerated, which causes the in-cylinder peak pressure and FMEP to rise, so that the friction but generally speaking, the trends at different loads are consistent with the relationship between loss becomes correspondingly larger. Contrary to the intake advance angle characteristic, under the the peak pressure and intake advance angle or exhaust retardation angle. high load condition, as the exhaust retardation angle decreases, the in-cylinder peak pressure decreases Under high-load conditions, as the Miller cycle is deepened, the in-cylinder peak pressure remarkably. When the exhaust retardation angle decreases from 30 to 5 CA, the in-cylinder peak changes oppositely from the medium and low load. This is because when◦ the Miller cycle is pressure drops from 77.32 to 67.11 bar, which is a drop of about 13%. This is followed by a drop in deepened, the CA50 is greatly decreased due to the advance of the ignition timing. Moreover, the FMEP, which dramatically reduces the in-cylinder friction loss, which indicates that a deep Miller cycle in-cylinder combustion rate is accelerated, which causes the in-cylinder peak pressure and FMEP to matching a smaller exhaust retardation angle under the high load condition can achieve an equivalent rise, so that the friction loss becomes correspondingly larger. Contrary to the intake advance angle FMEP to the baseline engine. Meanwhile, the baseline engine is further improved in terms of pumping characteristic, under the high load condition, as the exhaust retardation angle decreases, the in- loss and combustion after the Miller cycle application. cylinder peak pressure decreases remarkably. When the exhaust retardation angle decreases from 302.3.2. to 5 Performance °CA, the in-cylinder Increase ofpeak the pressure Miller Cycle drops from 77.32 to 67.11 bar, which is a drop of about 13%. This is followed by a drop in FMEP, which dramatically reduces the in-cylinder friction loss, whichAfter indicates the Miller that cyclea deep application, Miller cycle the matching above effects a smaller directly exhaust affect the retardation BSFC under angle diff erentunder loads. the highFigure load 12 showscondition BSFC can relative achieve to intakean equivalent advance FMEP angle underto the baseline baseline EVC, engine. and BSFCMeanwhile, relative the to baselineexhaust retardationengine is further angle improved under the Millerin terms cycle of pumping at different loss points and combustion of operation after during the the Miller test. cycle application.It can be seen from Figure 12 that, as the intake advance angle increases, the continuous deepening of the Miller cycle has a significant influence on BSFC. At the point of operation at 3000 rpm and 5 bar, the BSFC is reduced from 265.1 g/kW h of the baseline engine to 258.9 g/kW h, with an improvement 2.3.2. Performance Increase of the Miller· Cycle · of about 2.3%. For low-load conditions, the reduction in BSFC is mainly due to the significant improvementAfter the Miller in PMEP cycle as application, the Miller cycle the above is deepened, effects directly which is affect beneficial the BSFC to the under fuel economy.different loads.For medium Figureload 12 shows conditions, BSFC when relative the Millerto intake cycle advance is deepened, angle the under influence baseline of PMEP EVC, is and negligible, BSFC relativewhile the to combustion exhaust retardation and FMEP angle are slightlyunder the improved. Miller cycle As a result,at different when points the intake of operation advance angleduring is theadjusted test. from baseline VVT to Miller cycle VVT, BSFC drops from 236.3 to 233.1 g/kW h. · Energies 2020, 13, 1500 13 of 26

BMEP=5bar BMEP=10bar BMEP=15bar 265 260 255 250 245 240 Minimum Value

BSFC (g/kW·h) 235 230 0 5 10 15 20 25 Intake advance angle (°CA) 260 255 250 245 240 235 Minimum Value BSFC (g/kW·h) BSFC 230 0 5 10 15 20 25 30 Exhaust retardation angle (°CA) Figure 12. BSFC (Brake Specific Fuel Consumption) versus intake advance angle under baseline EVC, Figure 12. BSFC (Brake Specific Fuel Consumption) versus intake advance angle under baseline and BSFC versus exhaust retardation angle under the Miller cycle at different points of operation during EVC, and BSFC versus exhaust retardation angle under the Miller cycle at different points of the test. operation during the test. In addition, it can be inferred from Figure 12 that the engine load also determines BSFC considerably. The riseIt can of enginebe seen load from can Figure accelerate 12 that, the as in-cylinder the intake air advance flow speed angle and increases, the speed the of flamecontinuous front, deepeningwhich is beneficial of the Miller to reduce cycle the ha BSFC.s a significant However, asinfluence the engine on loadBSFC. rises, At thethegas point flow of in operation the cylinder at 3000rpmis strengthened and 5bar, and the the BSFC heat transferis reduced loss from is also 265.1 increased. g/kW·h Therefore,of the baseline the optimal engine BSFCto 258.9 is achievedg/kW·h, withwithin an the improvement range from mediumof about to 2.3%. high For load low-load at 3000 rpm conditions, (the above the two reduction effects arein BSFC balanced is mainly at medium due toand the high significant load). Theimprovement exhaust retardation in PMEP as angle the alsoMille hasr cycle a certain is deepened, impact onwhich BSFC. is beneficial Under low to load the fuelconditions, economy. since For the medium pumping load loss, conditions, combustion when and the friction Miller loss cycle vary is deepened, slightly with the the influence change of PMEPexhaust is retardationnegligible, while angle, the the combustion BSFC change and is alsoFMEP insignificant. are slightly BSFCimproved. reaches As 257.9a result, g/kW whenh when the intake advance angle is adjusted from baseline VVT to Miller cycle VVT, BSFC drops from ·236.3 to the exhaust retardation angle is 5 ◦CA at 3000 rpm and 5 bar, which is slightly lower than the IVC 233.1optimization g/kW·h. result by 1g/kW h. At the point of operation of 3000 rpm and 10 bar, the change of · exhaustIn addition, retardation it anglecan be still inferred has a potential from Figure for the 12 BSFC that reduction. the engine The load reason also is determines that the pumping BSFC considerably.loss and combustion The rise are of improvedengine load to somecan accelera extentte with the thein-cylinder advance air of theflow exhaust speed retardationand the speed angle. of flame front, which is beneficial to reduce the BSFC. However, as the engine load rises, the gas flow Therefore, compared with the IVC optimization results, when the exhaust retardation angle is 5 ◦CA, inthe the BSFC cylinder can reach is strengthened 230.5 g/kW h, and which the isheat about transfer 2.4% lowerloss isthan also theincreased. BSFC of Therefore, the baseline the engine. optimal · BSFCAt is theachieved point within of operation the range at 3000 from rpm medium and 15 to bar, high the load BSFC at 3000 has anrpm obvious (the above minimum two effects value are in balancedFigure 10 .at To medium explain itand clearly, high Figure load). 13 The shows exhaust BSFC retardation and related angle effects also relative has toa thecertain intake impact advance on BSFC.angle underUnder baseline low load EVC, conditions, and BSFC andsince related the pump effectsing relative loss, tocombustion the exhaust and retardation friction angle loss undervary slightly with the change of exhaust retardation angle, the BSFC change is also insignificant. BSFC the Miller cycle at the point of operation 3000 rpm and 10 bar. reaches 257.9 g/kW·h when the exhaust retardation angle is 5 °CA at 3000 rpm and 5 bar, which is At the point of operation at 3000 rpm and 15 bar, as Miller cycle is deepened, the PMEP variation slightly lower than the IVC optimization result by 1g/kW·h. At the point of operation of 3000 rpm range is relatively small, and the optimal intake advance angle is about 10 to 15 ◦CA; the ignition and 10 bar, the change of exhaust retardation angle still has a potential for the BSFC reduction. The timing and CA50 are significantly advanced, which results in a large improvement in combustion. reason is that the pumping loss and combustion are improved to some extent with the advance of However, due to the rise of in-cylinder peak pressure, the deteriorated friction loss offsets some of the the exhaust retardation angle. Therefore, compared with the IVC optimization results, when the improvement due to the benefit in combustion. As can be seen in Figure6, the intake pressure after exhaust retardation angle is 5 °CA, the BSFC can reach 230.5 g/kW·h, which is about 2.4% lower intercooler is also the maximum at the intake advance angle of 15 ◦CA. The relationship between the than the BSFC of the baseline engine. above three achieves the best overall effect when the intake advance angle is 15 ◦CA. Compared to At the point of operation at 3000 rpm and 15 bar, the BSFC has an obvious minimum value in 237.0 g/kW h BSFC of the baseline engine, the deep Miller cycle can be used to reduce it to 230.9/kW h, Figure 10. ·To explain it clearly, Figure 13 shows BSFC and related effects relative to the intake· which is about 2.6%. advance angle under baseline EVC, and BSFC and related effects relative to the exhaust retardation angle under the Miller cycle at the point of operation 3000 rpm and 10 bar. Energies 2020, 13, 1500 14 of 26

PMEP CA50 240 In-cylinder peak pressure 20 86 -0.28 84 238 18 82 80 236 16 78 -0.30 76 74

234 14 In-cylinder (bar) peak pressure 72 232 -0.32 12 70 68 CA50 (°CA) 230 (bar) PMEP 10 66 0 5 10 15 20 25 Intake advance angle (°CA) 0.1 20 78 BSFC (g/kW·h) 234 76 0.0 18 74 232 -0.1 72 16 70 -0.2 68 230 14 66 -0.3 64 228 12 5 1015202530 Exhaust retardation angle (°CA) Figure 13. Figure 13. BSFC and related effects effects versus intake advance angle under baseline EVC, and BSFC and related effects versus exhaust retardation angle under the Miller cycle (point of operation: n = 3000 rpm, related effects versus exhaust retardation angle under the Miller cycle (point of operation: n = 3000 BMEP = 15 bar, exhaust retardation angle = 25 CA, CA50 = 8 CA). rpm, BMEP = 15 bar, exhaust retardation angle ◦= 25 °CA, CA50◦ = 8 °CA). Similarly, the optimal BSFC at 15 CA exhaust retardation angle is the consequence of the At the point of operation at 3000 ◦rpm and 15 bar, as Miller cycle is deepened, the PMEP integrated effects of pumping loss, combustion and friction loss. As the exhaust retardation angle variation range is relatively small, and the optimal intake advance angle is about 10 to 15 °CA; the decreases, friction loss is well improved. At the same time, due to the knock limit, the CA50 deteriorates ignition timing and CA50 are significantly advanced, which results in a large improvement in remarkably, and besides, when the exhaust retardation angle continues to decrease from 15 CA, combustion. However, due to the rise of in-cylinder peak pressure, the deteriorated friction ◦loss the in-cylinder maximum combustion temperature (not shown) also drops by only a small amount. offsets some of the improvement due to the benefit in combustion. As can be seen in Figure 6, the Due to poor combustion, when the exhaust retardation angle is adjusted excessively, the BSFC saved by intake pressure after intercooler is also the maximum at the intake advance angle of 15 °CA. The pumping loss and friction loss improvement is insufficient to overcome the loss caused by combustion relationship between the above three achieves the best overall effect when the intake advance angle deterioration, resulting in degradation of engine fuel economy. Therefore, the optimal exhaust is 15 °CA. Compared to 237.0 g/kW·h BSFC of the baseline engine, the deep Miller cycle can be used retardation angle is set to 15 CA, and the BSFC at this time is 229.2 g/kW h. to reduce it to 230.9/kW·h, which◦ is about 2.6%. · By calculating the brake thermal efficiency, the improvement of the brake thermal efficiency after Similarly, the optimal BSFC at 15 °CA exhaust retardation angle is the consequence of the the application of the Miller cycle can be obtained. When the brake thermal efficiency is calculated, integrated effects of pumping loss, combustion and friction loss. As the exhaust retardation angle 44000 J/kg is used as the low calorific value of gasoline. Figure 14 shows the comparison of brake decreases, friction loss is well improved. At the same time, due to the knock limit, the CA50 thermal efficiency among the baseline engine and different Miller cycle optimization stages at different deteriorates remarkably, and besides, when the exhaust retardation angle continues to decrease points of operation. The Miller cycle optimization is defined as the settings of IVC optimization and from 15 °CA, the in-cylinder maximum combustion temperature (not shown) also drops by only a EVC optimization. small amount. Due to poor combustion, when the exhaust retardation angle is adjusted excessively, The brake thermal efficiency reaches 31.7%, 35.5%, and 35.7%, which has a 2.8%, 2.5%, and the BSFC saved by pumping loss and friction loss improvement is insufficient to overcome the loss 2.6% improvement compared to the baseline engine, respectively, at three points of operation. caused by combustion deterioration, resulting in degradation of engine fuel economy. Therefore, Therefore, regardless of the load at 3000 rpm, the brake thermal efficiency is prominently improved the optimal exhaust retardation angle is set to 15 °CA, and the BSFC at this time is 229.2 g/kW·h. after the Miller cycle application. The test shows that under the condition of 3000 rpm and 5 bar, due to By calculating the brake thermal efficiency, the improvement of the brake thermal efficiency slower airflow velocity and the higher pumping loss caused by smaller throttle opening, the brake after the application of the Miller cycle can be obtained. When the brake thermal efficiency is thermal efficiency at low load is about 4% inferior to that at medium and high load, which can indicate calculated, 44000 J/kg is used as the low calorific value of gasoline. Figure 14 shows the comparison that the optimal brake thermal efficiency of Miller cycle engine is achieved in the relatively high-load of brake thermal efficiency among the baseline engine and different Miller cycle optimization stages working range at 3000 rpm. at different points of operation. The Miller cycle optimization is defined as the settings of IVC optimization and EVC optimization. Energies 2020, 13, 1500 15 of 26

38 Abbreviations: Baseline engine IVC (Intate Valve Closing) IVC optimization(Experiment) EVC (Exhaust Valve Closing) IVC+EVC optimization(Experiment) BMEP (Brake Mean Effective Pressure) 36

32 Brake thermal efficiency (%)

30 BMEP=5bar BMEP=10bar BMEP=15bar Figure 14. Comparison of brake thermal efficiency. Figure 14. Comparison of brake thermal efficiency. 3. Numerical Study The brake thermal efficiency reaches 31.7%, 35.5%, and 35.7%, which has a 2.8%, 2.5%, and 2.6% Sinceimprovement the step compared size of the to intake the baseline advance engine and exhaust, respectively, retardation at three angle points during of operation. the test is Therefore,5 ◦CA, it isregardlessstill possible of the toload obtain at 3000 a higherrpm, th brakee brake thermal thermal e efficiencyfficiency withinis prominently the 5 ◦ CAimproved range. afterMoreover, the Miller further cycle research application.on the The combined test shows change that of under intake the and condition exhaust of VVT 3000is rpm still and relatively 5 bar, dueinsu toffi cient.slowerTherefore, airflow velocity this paper andwill the furtherhigher pumping study the loss point caused of operation by smaller 3000 throttle rpm and opening, 10 bar andthe brakethe VVT thermal settings efficiency at this point at low through load is numerical about 4% analysis. inferior to that at medium and high load, which can indicate that the optimal brake thermal efficiency of Miller cycle engine is achieved in the 3.1. 1D CFD Model relatively high-load working range at 3000 rpm. The 1D CFD model was established in GT-Power (v6.2, Gamma Technologies Inc., Westmont, IL, 3.United Numerical States) Study [18]. The model contains the combustion chamber, intake and exhaust and turbocharger modules. During the simulation, the point of operation is maintained at 3000 rpm and 10 bar. Since the step size of the intake advance and exhaust retardation angle during the test is 5 °CA, After changing the intake and exhaust VVT, the EWG opening degree is automatically adjusted to the it is still possible to obtain a higher brake thermal efficiency within the 5 °CA range. Moreover, required load by PID (Proportion Integration Differentiation) control. The heat transfer is calculated by further research on the combined change of intake and exhaust VVT is still relatively insufficient. the Woshini model [19]. The completed 1D CFD model is shown in Figure 15. Therefore, this paper will further study the point of operation 3000 rpm and 10 bar and the VVT settings at this point through numerical analysis.

3.1. 1D CFD Model The 1D CFD model was established in GT-Power (v6.2, Gamma Technologies Inc., Westmont, IL, United States) [18]. The model contains the combustion chamber, intake and exhaust and turbocharger modules. During the simulation, the point of operation is maintained at 3000rpm and 10bar. After changing the intake and exhaust VVT, the EWG opening degree is automatically adjusted to the required load by PID (Proportion Integration Differentiation) control. The heat transfer is calculated by the Woshini model [19]. The completed 1D CFD model is shown in Figure 15.

Figure 15. 1D CFD (Computational Fluid Dynamics) model of the engine. Figure 15. 1D CFD (Computational Fluid Dynamics) model of the engine.

50 Experiment data SIWiebe simulation SITurb simulation 40

Heat release rate (J/°CA) release Heat 10

-25 0 25 50 75 100 Crank angle (°CA)

Energies 2020, 13, 1500 16 of 26

3.2. Calibration Since several parameters need to be controlled during the simulation process, this paper considers the predictive requirements of the model. In addition to the more commonly used SIWiebe model to analyze the thermal efficiency of the engine, a more accurate quasi-three-dimensional predictable model SITurb is also introduced to study the performance of Miller cycle, which can consider factors such as the geometryFigure inside 15. 1D the CFD cylinder, (Computational ignition Fluid timing, Dynamics) airflow model andfuel of the properties engine. more precisely. Compared with the commonly used SIWiebe model, the SITurb model can more accurately predict the 3.2.influence Calibration of heat release rate of the in-cylinder combustion related to geometrical compression ratio, VVT operation, air:fuel ratio and ignition timing, which help analyze the suppression of knock versus Since several parameters need to be controlled during the simulation process, this paper Miller cycle parameters [20]. considers the predictive requirements of the model. In addition to the more commonly used According to the results of the Miller cycle optimization during the test, the SIWiebe model and SIWiebe model to analyze the thermal efficiency of the engine, a more accurate quasi-three- the SITurb model are calibrated by the measured in-cylinder pressure data. Figures 16 and 17 show the dimensional predictable model SITurb is also introduced to study the performance of Miller cycle, heat release rate and the in-cylinder pressure by simulation using SIWiebe and SITurb model and the which can consider factors such as the geometry inside the cylinder, ignition timing, airflow and test in-cylinder pressure. fuel properties more precisely. Compared with the commonly used SIWiebe model, the SITurb As can be seen from Figure 16, the error between both simulation data and test data at the point of model can more accurately predict the influence of heat release rate of the in-cylinder combustion operation 3000 rpm and 10 bar is below 3% (the error of BSFC and torque is below 1%). Figure 17 also related to geometrical compression ratio, VVT operation, air:fuel ratio and ignition timing, which confirms the accuracy of the simulation models compared to the test engine. Therefore, the established help analyze the suppression of knock versus Miller cycle parameters [20]. GT-Power model has good consistency with the baseline engine after test optimization, which is According to the results of the Miller cycle optimization during the test, the SIWiebe model a prerequisite for further analysis and optimization. In addition, the SIWiebe model was established and the SITurb model are calibrated by the measured in-cylinder pressure data. Figures 16 and 17 for the baseline VVT at 3000 rpm and 10 bar to compare the intake and exhaust flow of the baseline show the heat release rate and the in-cylinder pressure by simulation using SIWiebe and SITurb engine and Miller cycle optimization results during the test. model and the test in-cylinder pressure.

50 Experiment data SIWiebe simulation SITurb simulation 40

Heat release rate (J/°CA) release Heat 10

-25 0 25 50 75 100 Crank angle (°CA) Figure 16. Heat release rate by simulation using SIWiebe and SITurb models in comparison of measured

data by experiment (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 ◦CA, intake advance angle = 25 ◦CA, exhaust retardation angle = 5 ◦CA).

Figure 16. Heat release rate by simulation using SIWiebe and SITurb models in comparison of measured data by experiment (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 °CA,

intake advance angle = 25 °CA, exhaust retardation angle = 5 °CA). Figure 16. Heat release rate by simulation using SIWiebe and SITurb models in comparison of 70 Energiesmeasured2020, 13, 1500data by experiment (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 °CA,17 of 26 Experiment data intake advance angle = 25 °CA, exhaust retardation angle = 5 °CA). 60 SIWiebe simulation SITurb simulation 5070 Experiment data 4060 SIWiebe simulation SITurb simulation 3050

2040 In-cyclinder pressure (bar) 1030

200 In-cyclinder pressure (bar) 10 -100 -50 0 50 100 150 200 250 300 Crank angle (°CA) 0

Figure 17. In-cylinder pressure-100 by -50simulation 0 using 50 100SIWiebe 150 and 200 SITurb 250 model 300 in comparison of measured data by experiment (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 °CA, Crank angle (°CA) intake advance angle = 25 °CA, exhaust retardation angle = 5 °CA). Figure 17. In-cylinder pressure by simulation using SIWiebe and SITurb model in comparison of Figure 17. In-cylinder pressure by simulation using SIWiebe and SITurb model in comparison of measured data by experiment (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 CA, Asmeasured can be data seen by from experiment Figure (point16, the of erroroperatio betweenn: n = 3000 both rpm, simulation BMEP = data10 bar, and CA50 test = data8 °CA,◦ at the intake advance angle = 25 CA, exhaust retardation angle = 5 CA). pointintake of operation advance angle3000 =rpm 25 °CA, ◦and exhaust 10 bar retardationis below 3% angle (the = 5 error °CA).◦ of BSFC and torque is below 1%). Figure3.3. Results 17 also and Discussionconfirms the accuracy of the simulation models compared to the test engine. Therefore,As can the be established seen from GT-PowerFigure 16, model the error has betweengood consistency both simulation with the databaseline and enginetest data after at testthe optimization,point3.3.1. Massof operation Flow which 3000 is a prerequisiterpm and 10 forbar further is below analysis 3% (the and error optimization. of BSFC and In addition, torque is the below SIWiebe 1%). modelFigure was17 alsoestablished confirms for the the accuracy baseline ofVVT the at si3mulation000 rpm andmodels 10 barcompared to compare to the the testintake engine. and Figure 18 is the logP versus logV diagram of the baseline VVT and Miller cycle VVT at the point of exhaustTherefore, flow the of established the baseline GT-Power engine and model Miller has cycle good optimization consistency results with the during baseline the test.engine after test operation 3000 rpm and 10 bar. After Miller cycle application, the engine gains a smaller compression optimization, which is a prerequisite for further analysis and optimization. In addition, the SIWiebe ratio than that of the baseline VVT, while the expansion ratio remains the same owing to the same 3.3.model Results was and established Discussion for the baseline VVT at 3000 rpm and 10 bar to compare the intake and geometric compression ratio. The over expansion in the Miller cycle can also be clearly observed, exhaust flow of the baseline engine and Miller cycle optimization results during the test. 3.3.1.as shown Mass in Flow Figure 18. 3.3. Results and Discussion 3 Baseline VVT 3.3.1. Mass Flow Miller cycle VVT

23 Baseline VVT Miller cycle VVT

2 Pressure (bar) 1

Over expansion Pressure (bar) 1

0.1Over expansion 1 Volumn/Vmax FigureFigure 18. The logP versusversus logVlogV diagram diagram of of the the baseline baseline VVT VVT (Variable (Variable Valve Valve Timing) Timing) and Millerand Miller cycle 0.1 1 cycleVVT (pointVVT (point of operation: of operation: n = 3000 n = 3000 rpm, rpm, BMEP BMEP= 10 = bar, 10 CA50bar, CA50= 8 ◦ =CA). 8 °CA). Volumn/Vmax The research [20] shows that with the increase of the intake advance angle (which means the Figure 18. The logP versus logV diagram of the baseline VVT (Variable Valve Timing) and Miller deepening of the Miller cycle), the in-cylinder volumetric eﬃciency of the engine decreases. In the cycle VVT (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 °CA). simulation study of this paper, the in-cylinder volumetric eﬃciency of the baseline engine (whose VVT is close to Otto cycle) is 81.5% at the point of operation 3000 rpm and 10 bar, and the in-cylinder

Figure 18 is the logP versus logV diagram of the baseline VVT and Miller cycle VVT at the point of operation 3000 rpm and 10 bar. After Miller cycle application, the engine gains a smaller compression ratio than that of the baseline VVT, while the expansion ratio remains the same owing to the same geometric compression ratio. The over expansion in the Miller cycle can also be clearly observed, as shown in Figure 18. The research [20] shows that with the increase of the intake advance angle (which means the deepeningEnergies 2020, of13, the 1500 Miller cycle), the in-cylinder volumetric efficiency of the engine decreases. In18 ofthe 26 simulation study of this paper, the in-cylinder volumetric efficiency of the baseline engine (whose VVTvolumetric is close eﬃ tociency Otto ofcycle) deep is Miller 81.5% cycle at the engine, point which of operation is optimized 3000 by rpm test procedure,and 10 bar, approaches and the in- to cylinder80.5%. Figure volumetric 19 is a comparisonefficiency of of deep the massMiller ﬂow cycl ratee engine, of the baselinewhich is VVT optimized and Miller by cycletest procedure, VVT at the approachespoint of operation to 80.5%. 3000 Figure rpm and19 is10 a comparison bar. of the mass flow rate of the baseline VVT and Miller cycle VVT at the point of operation 3000 rpm and 10 bar.

Baseline exhaust Miller cycle exhaust Baseline intake Miller cycle intake 8 7 6 5 4 3 2

Valve lift (mm) Valve lift 1 0 -180 -120 -60 0 60 120 180 Crank angle (°CA) (a)Timing diagram of valve lifts 0.06 0.05 0.04 0.03 Mass flow increase 0.02 0.01 Reverse flow 0.00 Mass flow (kg/s) flow Mass -0.01 -180 -120 -60 0 60 120 180 Crank angle (°CA) (b)Timing diagram of mass flow Figure 19. The timing diagram and mass flow rate of the baseline VVT and Miller cycle VVT (point of Figure 19. The timing diagram and mass flow rate of the baseline VVT and Miller cycle VVT (point operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 CA). of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8◦ °CA). It is found from Figure 19 that the exhaust process is clearly advanced due to the decrease of the It is found from Figure 19 that the exhaust process is clearly advanced due to the decrease of exhaust retardation angle. The change of the intake mass flow after the Miller cycle application is more the exhaust retardation angle. The change of the intake mass flow after the Miller cycle application significant than that of the exhaust mass flow. Therefore, this paper mainly analyzes the influence of is more significant than that of the exhaust mass flow. Therefore, this paper mainly analyzes the the Miller cycle on the intake mass flow, and determines the reason for the slight decrease in in-cylinder influence of the Miller cycle on the intake mass flow, and determines the reason for the slight volumetric efficiency at the point of operation 3000 rpm and 10 bar. decrease in in-cylinder volumetric efficiency at the point of operation 3000 rpm and 10 bar. After the Miller cycle application, as the intake advance angle becomes larger, the IVO is also After the Miller cycle application, as the intake advance angle becomes larger, the IVO is also advanced, so the engine is still at the end of the exhaust stroke when the intake valve is opened. advanced, so the engine is still at the end of the exhaust stroke when the intake valve is opened. The The in-cylinder pressure is much larger than that at the baseline IVO. At the same time, the EVC is in-cylinder pressure is much larger than that at the baseline IVO. At the same time, the EVC is also also advanced relative to the baseline engine. The earlier EVC timing causes a part of the residual advanced relative to the baseline engine. The earlier EVC timing causes a part of the residual exhaust gas in the cylinder to flow back into the intake valve, resulting in the reverse flow effect shown exhaust gas in the cylinder to flow back into the intake valve, resulting in the reverse flow effect in Figure 19. This reverse flow phase hinders the engine to properly intake as expected during the shown in Figure 19. This reverse flow phase hinders the engine to properly intake as expected initial IVO timing, which reduces the in-cylinder volumetric efficiency of the Miller cycle. during the initial IVO timing, which reduces the in-cylinder volumetric efficiency of the Miller Meanwhile, as mentioned above, when the baseline engine is optimized to the deep Miller cycle, cycle. the throttle opening is boosted from 46.7% to 100% at 3000 rpm and 10 bar; consequently, the intake Meanwhile, as mentioned above, when the baseline engine is optimized to the deep Miller pressure is also enlarged from 12 to about 30 kPa. The impact in throttle opening and intake pressure cycle, the throttle opening is boosted from 46.7% to 100% at 3000 rpm and 10 bar; consequently, the change on intake mass flow can also be seen in Figure 19. When the intake valve is fully opened, intake pressure is also enlarged from 12 to about 30 kPa. The impact in throttle opening and intake the maximum intake mass flow of the Miller cycle is notably larger than that of the baseline engine, pressure change on intake mass flow can also be seen in Figure 19. When the intake valve is fully which is about 0.05 kg/s and 0.03 kg/s respectively. This improvement compensates for the adverse effect due to the deterioration in the in-cylinder volumetric efficiency caused by the reverse flow effect. In the case where both of them act simultaneously, the Miller cycle has an in-cylinder volumetric efficiency of 80.5% at 3000 rpm and 10 bar, which is only 1% lower than that of the baseline engine.

opened, the maximum intake mass flow of the Miller cycle is notably larger than that of the baseline engine, which is about 0.05 kg/s and 0.03 kg/s respectively. This improvement compensates for the adverse effect due to the deterioration in the in-cylinder volumetric efficiency caused by the reverse flow effect. In the case where both of them act simultaneously, the Miller cycle has an in-cylinder volumetricEnergies 2020, 13efficiency, 1500 of 80.5% at 3000 rpm and 10 bar, which is only 1% lower than that of19 ofthe 26 baseline engine. 3.3.2. The Analysis of Two Parameters 3.3.2. The Analysis of Two Parameters After experimental research and optimization, to go a step further, the simulation studied severalAfter eﬀ ectsexperimental based on theresearch variation and of optimization, two parameters. to go Figure a step 20 further,shows the the volumetric simulation e ﬃstudiedciency, severalPMEP, in-cylindereffects based maximum on the variationcombustion of two temperature parameters. and Figure in-cylinder 20 shows peak the pressure volumetric relative efficiency, to the PMEP,combined in-cylinder change ofmaximum intake advance combustion angle temperature and exhaust and retardation in-cylinder angle peak in pressure the vicinity relative of the to VVT the combinedsettings by change experimental of intake optimization advance angle at 3000 and rpm exhaust and 10 retardation bar. angle in the vicinity of the VVT settings by experimental optimization at 3000 rpm and 10 bar.

Volumetric efficiency 10 0.807 0.806 0.808 0.806 0.805 0.807 8 0.806 0.806 0.804 6 0.806 0.803 0.805 4 0.806 0.804 2 0.803 0.803 0 0.803 0.802 20 22 24 26 28 30 Exhaust retardation angleExhaust (°CA) Intake advance angle (°CA) (a) PMEP(bar) -0.22 10 -0.26 -0.25 -0.22 8 -0.25 -0.23 -0.24 6 -0.24 -0.24 4 -0.23 -0.25 2 -0.24 -0.25 -0.22 0 -0.26 -0.27 20 22 24 26 28 30 Exhaust retardation angle (°CA)Exhaust retardation angle Intake advance angle (°CA) (b) Maximum combustion temperature(°C) 10 2257.4 2251.4 2251.4 8 2245.3 2245.3 2239.3 6 2239.3 2233.2 4 2233.2 2227.2 2 2227.2 2221.1 2221.1 0 2215.1 2215.1 2209.0 20 22 24 26 28 30 Exhaust retardation angle (°CA) Intake advance angle (°CA) (c) In-cycliner peak pressure(bar) 61.9 63.7 10 61.6 63.4 8 63.1 6 62.2 62.8 62.5 4 62.5 62.8 62.2 2 63.1 61.9 0 61.6 63.4 61.2 20 22 24 26 28 30 Exhaust retardation angle (°CA) Intake advance angle (°CA) (d) Figure 20. Volumetric efficiency, PMEP (Pumping Mean Effective Pressure), in-cylinder maximum Figure 20. Volumetric efficiency, PMEP (Pumping Mean Effective Pressure), in-cylinder maximum combustion temperature and in-cylinder peak pressure versus the combined change of intake advance combustion temperature and in-cylinder peak pressure versus the combined change of intake angle and exhaust retardation angle (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 CA). advance angle and exhaust retardation angle (point of operation: n = 3000 rpm, BMEP = 10◦ bar, CA50From = Figure 8 °CA). 20a, at the point of operation 3000 rpm and 10 bar, the in-cylinder volumetric efficiency relative to the combined change of intake advance angle and exhaust retardation angle showsFrom a similar Figure trend 20a, to at that the ofpoint the analysisof operation discussed 3000 above.rpm and At the10 bar, same the time, in-cylinder as the Miller volumetric cycle is efficiencydeepened, relative when the to intakethe combined valve is opened,change of more intake intake advance air is pushedangle and back exhaust to the intakeretardation port byangle the showspiston, a which similar causes trend a to higher that of reverse the analysis flow loss discusse and lowerd above. volumetric At the same efficiency. time, Inas addition,the Miller it cycle can beis seen from Figure 20a that a suitable valve overlap angle also has a certain positive influence on the in-cylinder volumetric efficiency. Figure 20b indicates that the simulation results are consistent with the experimental results obtained above, that is, at the point of operation at 3000 rpm and 10 bar, the larger the intake advance

deepened, when the intake valve is opened, more intake air is pushed back to the intake port by the piston, which causes a higher reverse flow loss and lower volumetric efficiency. In addition, it can be seen from Figure 20a that a suitable valve overlap angle also has a certain positive influence on the in-cylinder volumetric efficiency. EnergiesFigure2020, 13 20b, 1500 indicates that the simulation results are consistent with the experimental 20results of 26 obtained above, that is, at the point of operation at 3000 rpm and 10 bar, the larger the intake advance angle with an appropriate valve overlap angle, the lower the pumping loss that can be angle with an appropriate valve overlap angle, the lower the pumping loss that can be achieved. achieved. The lower-right corner area in Figure 20b is an area where the pumping loss is relatively The lower-right corner area in Figure 20b is an area where the pumping loss is relatively optimal, optimal, whose PMEP is about 0.04 bar higher than that in the upper-left corner area. It can be seen whose PMEP is about 0.04 bar higher than that in the upper-left corner area. It can be seen from from Figure 20c that the combustion effect of the Miller cycle at 3000 rpm and 10 bar is mainly Figure 20c that the combustion effect of the Miller cycle at 3000 rpm and 10 bar is mainly determined by determined by the exhaust retardation angle. Decreasing the exhaust retardation angle greatly the exhaust retardation angle. Decreasing the exhaust retardation angle greatly reduces the in-cylinder reduces the in-cylinder maximum combustion temperature, thus reducing the heat load loss of the maximum combustion temperature, thus reducing the heat load loss of the Miller cycle engine and Miller cycle engine and also improving the lubrication and emission characteristics of the Miller also improving the lubrication and emission characteristics of the Miller cycle engine to some extent; cycle engine to some extent; the trend of the in-cylinder maximum combustion temperature relative the trend of the in-cylinder maximum combustion temperature relative to the intake advance angle to the intake advance angle deviates from the test result, which may be due to the lack of deviates from the test result, which may be due to the lack of consideration of vortex and tumble consideration of vortex and tumble flow in 1D CFD model. Overall, the combustion situation is flow in 1D CFD model. Overall, the combustion situation is relatively better for the lower part of the relatively better for the lower part of the distribution map (about 35 °C lower than the upper part), distribution map (about 35 C lower than the upper part), so it is expected to contribute to the fuel so it is expected to contribute◦ to the fuel economy of the Miller cycle engine. economy of the Miller cycle engine. It can be concluded from Figure 20d that the in-cylinder peak pressure is mainly affected by It can be concluded from Figure 20d that the in-cylinder peak pressure is mainly affected by the the exhaust retardation angle. At the point of operation 3000 rpm and 10 bar, as the exhaust exhaust retardation angle. At the point of operation 3000 rpm and 10 bar, as the exhaust retardation retardation angle decreases by 10 °CA, the in-cylinder peak pressure shows an increase of about angle decreases by 10 CA, the in-cylinder peak pressure shows an increase of about 1.5bar, which also 1.5bar, which also has◦ the same trend as the test results above. Meanwhile, when the exhaust has the same trend as the test results above. Meanwhile, when the exhaust retardation angle is relatively retardation angle is relatively small, increasing the intake advance angle can slightly reduce the in- small, increasing the intake advance angle can slightly reduce the in-cylinder peak pressure. The FMEP cylinder peak pressure. The FMEP is calculated by the Chen-Flynn formula [21], which shows that is calculated by the Chen-Flynn formula [21], which shows that the distribution law of FMEP is exactly the distribution law of FMEP is exactly the same as the distribution law of the in-cylinder peak the same as the distribution law of the in-cylinder peak pressure relative to the combined change pressure relative to the combined change of intake advance angle and exhaust retardation angle in of intake advance angle and exhaust retardation angle in Figure 20d. So we can conclude that the Figure 20d. So we can conclude that the lower-left corner area in Figure 20d is the high friction loss lower-left corner area in Figure 20d is the high friction loss zone during the Miller cycle operation. zone during the Miller cycle operation. Knock is a limitation of the engine’s work and is one of the keys to the Miller cycle to improve Knock is a limitation of the engine’s work and is one of the keys to the Miller cycle to improve the fuel economy of the engine. The predictable knock phenomenological model of the gasoline the fuel economy of the engine. The predictable knock phenomenological model of the gasoline engine [20] and the output parameter Knock index can be used to analyze knock. When the Knock engine [20] and the output parameter Knock index can be used to analyze knock. When the Knock index is greater than 200, it is considered that knock has occurred. Figure 21 shows the Knock index index is greater than 200, it is considered that knock has occurred. Figure 21 shows the Knock index relative to the combined change of intake advance angle and exhaust retardation angle in the vicinity relative to the combined change of intake advance angle and exhaust retardation angle in the of the VVT settings by experimental optimization at 3000rpm and 10bar. The result is based on fixed vicinity of the VVT settings by experimental optimization at 3000rpm and 10bar. The result is based ignition timing. on fixed ignition timing. Knock index 283.5 10 172.5 154.0 265.0 191.0 8 246.5

228.0 6 209.5 4 191.0

2 228.0 209.5 172.5

Exhaust retardation angle (°CA) retardation Exhaust 246.5 154.0 0 265.0 135.5 20 22 24 26 28 30 Intake advance angle (°CA) Figure 21. Knock index versus the combined change of intake advance angle and exhaust retardation Figure 21. Knock index versus the combined change of intake advance angle and exhaust angle (point of operation: n = 3000 rpm, BMEP = 10 bar, ignition timing = 18.75 CA BTDC (Before Top retardation angle (point of operation: n = 3000 rpm, BMEP = 10 bar, ignition◦ timing = 18.75 °CA Dead Center)). BTDC (Before Top Dead Center)).

It can be seen from Figure 21 that when the intake advance angle is reduced from 25 ◦CA, the engine will knock at the same ignition timing (indicating that the Knock index is over 200). Therefore, to ensure the normal operation of the engine, the ignition timing can only be delayed. On the other hand, as the Miller cycle continues to deepen, knock is suppressed, so that the optimal ignition timing can be Energies 2020, 13, 1500 21 of 26 further advanced, thereby improving BSFC. This result is consistent with the experimental results above. The reason is that the intake advance angle greatly affects the effective compression ratio of the Miller cycle, which is, a larger intake advance angle produces a smaller effective compression ratio. Therefore, the occurrence of knock is well suppressed. The exhaust retardation angle also has a certain impact on knocking. As the exhaust retardation angle becomes larger, knock is suppressed. This is because the exhaust gas is discharged more completely as the exhaust valve is closed later. Therefore, the temperature in the cylinder is lower during the subsequent intake stroke. In addition, the EWG opening at the point of operation at 3000 rpm and 10 bar is analyzed in the simulation procedure. Figure 22 shows the back pressure relative to the combined change of intake advance angle and exhaust retardation angle in the vicinity of the VVT settings by experimental optimization at 3000 rpm and 10 bar; Figure 23 shows the EWG opening relative to the combined change of intake advance angle and exhaust retardation angle in the vicinity of the VVT settings by It can be seen from Figure 21 that when the intake advance angle is reduced from 25 °CA, the experimental optimization at 3000 rpm and 10 bar. Note that when the Miller cycle is applied in this engine will knock at the same ignition timing (indicating that the Knock index is over 200). condition, the throttle opening has reached 100%. Therefore, to ensure the normal operation of the engine, the ignition timing can only be delayed. On As shown in Figure 22, the back pressure at the lower-right corner is lower than that at the the other hand, as the Miller cycle continues to deepen, knock is suppressed, so that the optimal upper-left corner. After the application of the Miller cycle, a lower back pressure reduces the resistance ignition timing can be further advanced, thereby improving BSFC. This result is consistent with the during exhausting, which reduces the total loss. experimental results above. The reason is that the intake advance angle greatly affects the effective When the throttle is fully open, the load is completely satisfied by a turbocharger. compression ratio of the Miller cycle, which is, a larger intake advance angle produces a smaller Therefore, after the Miller cycle application at 3000 rpm and 10 bar, a relatively small valve timing effective compression ratio. Therefore, the occurrence of knock is well suppressed. The exhaust change can cause a dramatic EWG position change. The result is consistent with the conclusions from retardation angle also has a certain impact on knocking. As the exhaust retardation angle becomes the test results above. It is worth noting that when the Miller cycle is deepened to an intake advance larger, knock is suppressed. This is because the exhaust gas is discharged more completely as the angle of 30 ◦CA and an exhaust retardation angle of around 0 ◦CA, the EWG opening is less than 7%. exhaust valve is closed later. Therefore, the temperature in the cylinder is lower during the During the test, it was found that the EWG opening will also generate about 5% fluctuation under fixed subsequent intake stroke. working conditions, which is due to the output fluctuation caused by the engine’s working cycle and In addition, the EWG opening at the point of operation at 3000 rpm and 10 bar is analyzed in the normal influence caused by PID control, etc., but it also means if the throttle and EWG remain fully the simulation procedure. Figure 22 shows the back pressure relative to the combined change of open during the actual operation of the engine, their fluctuations will have certain adverse effects on intake advance angle and exhaust retardation angle in the vicinity of the VVT settings by the output characteristics. Therefore, in this paper, the EWG position after the Miller cycle application experimental optimization at 3000 rpm and 10 bar; Figure 23 shows the EWG opening relative to is regarded as one of the limitations for optimizing VVT under 3000rpm and 10bar condition. When the the combined change of intake advance angle and exhaust retardation angle in the vicinity of the deep Miller cycle is applied to improve thermal efficiency, the EWG opening must be kept above 7% to VVT settings by experimental optimization at 3000 rpm and 10 bar. Note that when the Miller cycle ensure the stability of the engine under this condition. is applied in this condition, the throttle opening has reached 100%. Back pressure(bar)

1.033 10 1.032 1.033 1.032

8 1.032 1.032

1.032 6

1.032 1.032

4 1.032 1.032

2 1.031 1.031 Exhaust retardation angle (°CA) angle retardation Exhaust 1.031 0 1.031 1.031 20 22 24 26 28 30 Intake advance angle (°CA) Figure 22. Back pressure versus the combined change of intake advance angle and exhaust retardation Figure 22. Back pressure versus the combined change of intake advance angle and exhaust angle (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 CA, throttle opening = 100%). retardation angle (point of operation: n = 3000 rpm, BMEP = 10 bar,◦ CA50 = 8 °CA, throttle opening = 100%).

Energies 2020, 13, 1500 22 of 26

EWG opening(%) 22.9 10 20.7 20.7 8 16.2 18.4 18.4 11.7 16.2 6 14.0 14.0 4 11.7

2 9.5

Exhaust retardation angle (°CA)Exhaust 9.5 7.2 7.2 0 5.0 20 22 24 26 28 30

Intake advance angle (°CA) Figure 23. EWG opening versus the combined change of intake advance angle and exhaust retardation Figure 23. EWG opening versus the combined change of intake advance angle and exhaust angle (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 CA, throttle opening = 100%). retardation angle (point of operation: n = 3000 rpm, BMEP = 10 bar,◦ CA50 = 8 °CA, throttle opening = The100%). BSFC of the Miller cycle engine is the main research object of this paper. From the above analysis, it can be concluded that a deeper Miller cycle suppresses the occurrence of knock, and accordingly, the ignitionAs shown timing in canFigure be further 22, the advanced. back pressure Therefore, at the in lower-right the simulation corner research, is lower combined than that with at the the aboveupper-left experimental corner. experienceAfter the andapplication simulation of resultsthe Miller above, cycle, based a onlower the testback optimized pressure VVT reduces settings the andresistance the ignition during timing exhausting, of the optimization which reduces result, the thetotal ignition loss. timing is appropriately adjusted under When the throttle is fully open, the load is completely satisfied by a turbocharger. Therefore, diﬀerent VVT settings to ensure that the CA50 reaches 8 ◦CA. Under this condition, the combustion ofafter the Millerthe Miller cycle cycle engine application is nearly at optimized. 3000 rpm Afterand 10 the bar, CA50 a relatively is controlled, small thevalve BSFC timing relative change to the can combinedcause a dramatic change of EWG intake position advance change. angle andThe exhaustresult is retardation consistent angle.with theAt theconclusionssame time, fromsimilar the test to theresults simulation above. hypothesis It is worth made noting in thethat above when analysis the Miller of combustion, cycle is deepenedto simplify to anthe intake simulation advance process, angle CA10-90of 30 °CA remains and an unchanged. exhaust retardation Figure 24 showsangle of the around BSFC relative0 °CA, the to the EWG combined opening change is less of than intake 7%. advanceDuring anglethe test, and it was exhaust found retardation that the EWG angle openin in theg vicinitywill also of generate the VVT about settings 5% byfluctuation experimental under fixed working conditions, which is due to the output fluctuation caused by the engine's working optimization at 3000 rpm and 10 bar. cycle and the normal influence caused by PID control, etc., but it also means if the throttle and EWG remain fully open during the actual operation of the engine,BSFC(g/kW·h) their fluctuations will have certain adverse effects on the output characteristics. Therefore, in this paper, 231.3the EWG position after the 10 Miller cycle application is regardedPMEP-Combustion as one of degradation the limitations for optimizing231.1 VVT under 3000rpm and 10bar condition. When the deep Miller cycle is applied to improve thermal efficiency, the EWG 8 230.9 opening must be kept above 7% to ensure the230.6 stability of the engine under this condition. 230.7 The BSFC of the Miller6 cycle engine is the main research object of this paper. From the above analysis, it can be concluded that a deeper Mille230.4r cycle suppresses the230.6 occurrence of knock, and FMEP-Knock accordingly, the ignition4 timing can be further advanced. Therefore, in the simulation research, degradation 230.4 230.9 combined with the above experimental experience230.2 and simulation results above, based on the test optimized VVT settings2 and the ignition timing of the230.0 optimization result,230.2 the ignition timing is

Exhaust retardation angle (°CA) retardation Exhaust 230.7 Low BSFC appropriately adjusted under different VVT settings to ensure that the CA50230.0 reaches 8 °CA. Under 0 region EWG this condition, the combustion of the Miller cycle enginelimitation is nearly optimized.229.8 After the CA50 is controlled, the BSFC relative20 to 22the 24combined 26 change 28 30 of intake advance angle and exhaust retardation angle. At the same time,Intake similar advance to angle the (°CA) simulation hypothesis made in the above analysis of combustion, to simplify the simulation process, CA10-90 remains unchanged. Figure 24 shows Figure 24. BSFC versus the combined change of intake advance angle and exhaust retardation angle the BSFCFigure relative 24. BSFC to versus the combined the combined change change of intaof intakeke advance advance angle angle andand exhaustexhaust retardation retardation angle angle in (point of operation: n = 3000 rpm, BMEP = 10 bar, CA50 = 8 ◦CA). the vicinity(point of of operation: the VVT n settings= 3000 rpm, by experimentalBMEP = 10 bar, optimization CA50 = 8 °CA). at 3000 rpm and 10 bar. It can be seen from Figure 24 that, based on the test optimization results, there is still potential for improvementIt can be seen in from further Figure enhancing 24 that, the based brake on thermal the test e ffioptimizationciency of the results, Miller there cycle engine.is still potential For the for improvement in further enhancing the brake thermal efficiency of the Miller cycle engine. For the convenience of description, four regions are marked from Figure 24. The brake thermal efficiency of the upper PMEP-combustion degradation zone (with large exhaust retardation angle) is lower than the test optimization result, and the results can be explained by the above simulation analysis: As can be inferred from Figure 20b, the upper area of the PMEP distribution is the area where the pumping loss is larger, and the pumping loss of the upper area is about 15% larger than that of the lower right corner. In addition, as shown in Figure 20c, the in-cylinder maximum combustion temperature in this region is higher, thereby increasing the heat load. In summary, poor combustion and large pumping losses result in a decrease in the brake thermal efficiency of the PMEP-Combustion degradation zone. The brake thermal efficiency of the FMEP-Knock degradation zone (with small intake advance angle and exhaust retardation angle) at the lower-left corner is also lower than the test optimization result but different from the upper area, this area is mainly affected by FMEP and Knock limitation: The lower-left corner of Figures 20d and 21 are the areas of larger in-cylinder peak pressure and Knock index. The larger in-cylinder peak pressure leads to an increase in FMEP during the operation of the Miller cycle engine; in addition, the Knock index obtained through the fixed ignition timing is much larger than 200. If the ignition timing by test optimization is maintained, the engine will knock and fail to operate normally while working in this area. Therefore, when the CA50 is determined to be 8 °CA to avoid knock during the use of the SIWiebe model, the ignition timing has been already delayed. In summary, greater friction loss and delayed ignition timing deteriorate the brake thermal efficiency of the FMEP-Knock degradation zone. The BSFC at the low BSFC region in Figure 24 (with large intake advance angle and small exhaust retardation angle) is further reduced than that of test optimization result, owing to an integrated consequence of pumping loss, combustion, friction loss and knock. It is obvious that the pumping loss in this area is relatively small from the PMEP distribution, and the Knock index is also optimal in this area. Therefore, the ignition timing can be further advanced when the Miller cycle engine works in this area. Secondly, from the in-cylinder peak pressure distribution, increasing the intake advance angle can compensate for the increase in FMEP caused by faster combustion. Low pumping losses, advanced ignition timing and low friction loss act comprehensively to improve the brake thermal efficiency in this region. The brake thermal efficiency in the EWG limitation zone is even higher, but as mentioned above, since the EWG in this region is almost fully open, load fluctuation affects the normal operation of the engine. As a result, Miller cycle engines should avoid operating in this region. Energies 2020, 13, 1500 23 of 26 convenience of description, four regions are marked from Figure 24. The brake thermal efficiency of the upper PMEP-combustion degradation zone (with large exhaust retardation angle) is lower than the test optimization result, and the results can be explained by the above simulation analysis: As can be inferred from Figure 20b, the upper area of the PMEP distribution is the area where the pumping loss is larger, and the pumping loss of the upper area is about 15% larger than that of the lower right corner. In addition, as shown in Figure 20c, the in-cylinder maximum combustion temperature in this region is higher, thereby increasing the heat load. In summary, poor combustion and large pumping losses result in a decrease in the brake thermal efficiency of the PMEP-Combustion degradation zone. The brake thermal efficiency of the FMEP-Knock degradation zone (with small intake advance angle and exhaust retardation angle) at the lower-left corner is also lower than the test optimization result but different from the upper area, this area is mainly affected by FMEP and Knock limitation: The lower-left corner of Figures 20d and 21 are the areas of larger in-cylinder peak pressure and Knock index. The larger in-cylinder peak pressure leads to an increase in FMEP during the operation of the Miller cycle engine; in addition, the Knock index obtained through the fixed ignition timing is much larger than 200. If the ignition timing by test optimization is maintained, the engine will knock and fail to operate normally while working in this area. Therefore, when the CA50 is determined to be 8 ◦CA to avoid knock during the use of the SIWiebe model, the ignition timing has been already delayed. In summary, greater friction loss and delayed ignition timing deteriorate the brake thermal efficiency of the FMEP-Knock degradation zone. The BSFC at the low BSFC region in Figure 24 (with large intake advance angle and small exhaust retardation angle) is further reduced than that of test optimization result, owing to an integrated consequence of pumping loss, combustion, friction loss and knock. It is obvious that the pumping loss in this area is relatively small from the PMEP distribution, and the Knock index is also optimal in this area. Therefore, the ignition timing can be further advanced when the Miller cycle engine works in this area. Secondly, from the in-cylinder peak pressure distribution, increasing the intake advance angle can compensate for the increase in FMEP caused by faster combustion. Low pumping losses, advanced ignition timing and low friction loss act comprehensively to improve the brake thermal efficiency in this region. The brake thermal efficiency in the EWG limitation zone is even higher, but as mentioned above, since the EWG in this region is almost fully open, load fluctuation affects the normal operation of the engine. As a result, Miller cycle engines should avoid operating in this region. As the intake advance angle is further increased compared with the test optimization result and the corresponding exhaust retardation angle is further reduced by 1–4 ◦CA, the brake thermal efficiency continues to improve. Meanwhile, the EWG opening can be guaranteed to be greater than 7%. Therefore, after considering the thermal efficiency and working stability of the Miller cycle engine, the intake advance angle and exhaust retardation angle are finally optimized to 28 and 2 ◦CA. Figure 25 shows the results of BSFC and brake thermal efficiency improvement after the baseline engine accomplishes each optimization stage at 3000 rpm and 10 bar. Through the research of experiments and simulations, the final optimization results ensure that the engine’s stable work at 3000rpm and 10bar, and the brake thermal efficiency is improved from the baseline engine of 34.6% to 35.6% (a relative improvement of 2.9%), which greatly enhances the fuel economy. In addition, the final optimization result was confirmed by a final experiment. The BSFC of the experimental result is 226.8 g/kW h, which is 1.3% lower than the final optimization result. ·

As the intake advance angle is further increased compared with the test optimization result and the corresponding exhaust retardation angle is further reduced by 1–4 °CA, the brake thermal efficiency continues to improve. Meanwhile, the EWG opening can be guaranteed to be greater than 7%. Therefore, after considering the thermal efficiency and working stability of the Miller cycle engine, the intake advance angle and exhaust retardation angle are finally optimized to 28 and 2 °CA.Energies Figure2020, 13 ,25 1500 shows the results of BSFC and brake thermal efficiency improvement after24 ofthe 26 baseline engine accomplishes each optimization stage at 3000 rpm and 10 bar.

36.0 The line chart represents the BSFC. The bar chart represents the efficiency. 35.8 236 0.3% 35.6 35.4 1.1% 35.2 234 35.0 1.4% 34.8 232 BSFC (g/kW.h) 34.6

Brake thermal efficiency (%) efficiency thermal Brake 34.4 34.2 230 34.0 Baseline IVC IVC+EVC IVC+EVC engine optimization optimization optimization (Experiment) (Simulation) Figure 25. BSFC and brake thermal efficiency eﬃciency improvementimprovement after the baseline engine accomplishes

each optimization stage (point of operation: n = 30003000 rpm, rpm, BMEP BMEP == 1010 bar, bar, CA50 CA50 == 88 °CA).◦CA).

4. ConclusionsThrough the research of experiments and simulations, the final optimization results ensure that theThis engine's paper proposes stable work the ideaat 3000rpm of Miller and cycle 10bar, application and the on brake conventional thermal enginesefficiency to is improve improved the fromthermal the e ffibaselineciency and engine thus reduceof 34.6% fuel to consumption. 35.6% (a relative Miller cycleimprovement can decouple of 2.9%), the expansion which greatly ratio of enhancesthe engine the from fuel the economy. compression In addition, ratio. For the the commonfinal optimization operating result conditions was ofconfirmed the gasoline by engine,a final experiment.the intake valve The canBSFC be of closed the experimental early with the result appropriate is 226.8 valveg/kW·h, overlap which angle, is 1.3% combined lower than with the a larger final optimizationthrottle opening result. and turbocharger application and the appropriate ignition timing, thereby reducing the BSFC of the engine. The main conclusions of this paper are as follows: 4. ConclusionsFor the engine’s common speed of 3000 rpm, based on the engine bench test data, the Miller cycleThis of thepaper baseline proposes engine the underidea of di Millerfferent cycle loads application at 3000 rpm on wasconventional first studied. engines The to pumpingimprove theloss, thermalcombustion efficiency and friction and thus loss reduce were comparedfuel consumption. and analyzed. Miller cycle The results can decouple show that the after expansion Miller ratiocycle of application, the engine thefrom pumping the compressionloss is improved ratio. For due the tocommon the increase operating of the conditions throttle openingof the gasoline under engine,constant the load; intake the Millervalve can cycle be under closed medium early with and the low appropriate load can reduce valve theoverlap combustion angle, combined duration. withThe in-cylinder a larger throttlemaximum opening combustion and turbocharger temperature application is lower thanand the that appropriate of the baseline ignition engine timing, after therebyMiller cycle reducing application the BSFC with of the engine. appropriate The ma valvein conclusions overlap angle. of this The paper friction are as loss follows: increases with the riseFor ofthe the engine's load; when common the Millerspeed cycleof 3000 deepens, rpm, base thed friction on the lossengine increases bench attest high data, loads, the butMiller as cyclethe exhaust of the baseline valve closes engine earlier, under the different friction lossloads decreases at 3000 rpm significantly, was first which studied. also The demonstrates pumping loss, the combustionimportance ofand matching friction exhaustloss were valve compared timing. and analyzed. The results show that after Miller cycle application,According the topumping the improvement loss is improved effect of due the Millerto the cycleincrease on theof the BSFC throttle under opening different under loads constantat 3000 rpm load;during the Miller the test, cycle the under VVT medium angle was and adjusted low load based can onreduce the baselinethe combustion engine. duration. The test Theoptimization in-cylinder was maximum completed combustion after adjusting temperature the intake is lower advance than angle that of to the 25, baseline 25 and 15engine◦CA, after and, Millercorrespondingly, cycle application the exhaust with retardation the appropriate angle tovalve 5, 5 andoverlap 15 ◦CA angle. at the The point friction of operation loss increases of 3000 with rpm theand rise 5, 10 of and the15 load; bar, when respectively. the Miller cycle deepens, the friction loss increases at high loads, but as the exhaustTo further valve study closes the pointearlier, of operationthe friction at loss 3000 decreases rpm and 10significantly, bar, whichis which quite commonlyalso demonstrates used by thegasoline importance engines, of basedmatching on theexhaust test optimizedvalve timing. VVT results, the 1D CFD analysis of the Miller cycle was carried out. On the one hand, combined with the results obtained during the test, the influence of the combined change of intake advance angle and exhaust retardation angle on the Miller cycle is studied by simulation. On the other hand, based on the test optimization results, a higher brake thermal efficiency is found to achieve better fuel economy. For the results that have been obtained in the test, 1D simulation is consistent with the experimental results, and the simulation results further illustrate the influence intensity and coupling relationship on each parameter. The simulation also analyzed the effect of the Miller cycle on knocking. The engine has better brake thermal efficiency as Energies 2020, 13, 1500 25 of 26 the Miller cycle deepens. The main reason is that the engine’s knock tendency decreases as the intake advance angle increases, so the ignition timing can be further advanced. At the point of operation 3000 rpm and 10 bar, further increasing the intake advance angle and appropriately reducing the exhaust retardation angle compared to the test optimization result make the Miller cycle engine work in the low BSFC region mentioned in the paper, which can achieve the optimal fuel economy while ensuring stable engine operation. The maximum brake thermal efficiency is increased by 0.23% relative to the test optimization result, and 2.71% relative to the baseline engine.

Author Contributions: Conceptualization, D.L.; Data curation, X.P.; Formal analysis, X.P. and Y.Z.; Funding acquisition, D.L.; Methodology, L.F.; Project administration, L.F.; Resources, D.L.; Software, Y.Z.; Supervision, D.L.; Validation, Y.Z.; Writing—original draft, X.P. and Y.Z. All authors have read and agreed to the published version of the manuscript. Funding: This project was supported by the National Key R&D Program of China (2018YFB0105802). Conﬂicts of Interest: The authors declare no conﬂict of interest.

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