1-D Simulation of a Four Cylinder DirectInjection Supercharged Diesel EngineEquipped with VVT Mechanism

Cristian Soimaru, Anghel Chiru and Daniel Buzea

Abstract Present paper exhibits the influences that variable valve timing concepthas on a four cylinder turbocharged diesel engine. In this scope, using Amesim 101-D simulation software the engine was created and after two main strategies ofvalve timing were studied in order to observe the behaviour of intake mass flowrate and pumping losses diagram during 100 cycles.

Keywords Engine � Diesel � Valve timing � Mass flow � Pumping losses

1 Introduction

Automotive engineers, especially the ones who deal with internal combustionengines, in the last 20 years have been more and more pressured to develop newsolutions that would improve energetically and environmental performanceparameters like: reduction of CO, HC and NOx emissions; lowering fuel con-sumption; higher power and torque. Among the solutions developed there may belisted: common-rail high pressure fuel injection systems, variable geometry tur-bochargers, after treatment systems, variable compression ratio engine, variablevalve timing and lift etc.

F2012-A06-039

C. Soimaru (&) � A. Chiru � D. BuzeaDepartment D02–High Tech Products for Automotive,Transilvania University of Bras�ov, Bras�ov, Romaniae-mail: cristian.soimaru@unitbv.ro

SAE-China and FISITA (eds.), Proceedings of the FISITA 2012 WorldAutomotive Congress, Lecture Notes in Electrical Engineering 190,DOI: 10.1007/978-3-642-33750-5_21, � Springer-Verlag Berlin Heidelberg 2013

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From solutions presented above, variable valve timing represents the focus ofthis paper. This strategy has been studied for a long time and its theoreticalapproach is presented in Atkinson and Miller cycles. Paper [1] presents theprinciple of obtaining variable valve actuation in two ways: by modifying theexhaust valve closing, according to Atkinson cycles, and modifying the intakevalve closing, Miller cycles. Both of the techniques have the same mainimprovement, expansion stoke bigger than compression stroke so that highertorque could be obtained and a decreasing value of pumping looses also.

Within last 20 years many studies have been conducted on spark ignitionengines equipped with variable valve timing system so that the influence that itmanifests over the energetically and environmental parameters could be deter-mined. Scientific papers like [2–4] present the influences that the variable valvetiming involves over engine’s energetically performances, emissions and new fuelburning concepts like CAI—Controlled Auto Ignition.

Regarding diesel engines the variable event valve timing has not been appropriateto be studied due to reasons like: wide area spreading of turbochargers due to theirimprovements to engine’s efficiency, internal and external gas recycling efficiencyand improvements for internal combustion engines, and the third one huge solutionlike high pressure fuel injection systems and injection characteristics [5].

2 Paper’s Objective

The objective of this paper is to present the influence that dimension of valveoverlapping has on the internal combustion engine’s intake mass flow rate and oncylinder pressure, by modifying the moment of intake valve opening. Thisobjective was possible to achieve by using 1-D Amesim 10 software simulationtool.

3 Methodology

3.1 Theoretical Approach of the Simulation

The theoretical approach of the simulation is supplied by [1, 6–8].

3.1.1 Volumetric Efficiency

gv ¼2 � _ma

qa � Vh � Nð1Þ

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_ma ¼ air mass flow inducted into the engine, qa ¼ air density, N ¼ enginespeed

Vh ¼ engine displacement.

3.1.2 Pseudo: Flow Velocity

vps ¼1

Am� dV

dh¼ p � B2

4 � Am� ds

dhð2Þ

V = cylinder volume, B = cylinder bore, s = distance between crank axis andwrist pin, Am = valve area.

3.1.3 Air Flow Rate

_m ¼ CD � AR � p0

R � T0ð Þ1=2� pT

p0

� �2 � cc� 1

� 1� pT

p0

� � c�1cð Þ

" #( )1=2

ð3Þ

CD = discharge coefficient, p0 = upstream stagnation pressure, T0 = upstreamstagnation temperature, qT = pressure at the restriction, AR = reference area ofthe valve, c = adiabatic coefficient.

3.1.4 Equivalence Ratio

/ ¼A=F

� �acutal

A=F

� �teoretic

ð4Þ

A=F ¼ Air � fuelratio

3.1.5 Intake Mass Speed

patm � pc ¼X

Dpj ¼ qa � �S2p �X

nj �Ap

Aj

� �2

ð5Þ

patm = atmospheric pressure, pc = cylinder pressure, �Sp ¼ mean piston speed,Ap = piston area, Aj = component minimum flow area, Dpj ¼ total quasy steady

1-D Simulation of a Four Cylinder Direct Injection Supercharged Diesel Engine 1091

nj ¼ resistance coefficient for that component which depends on its geometricdetails

Ap = piston area, Aj = component minimum flow area, Dpj ¼ total quasy steadypressure loss.

3.1.6 Mass Flow Rate

dm

dt¼ Aeff � p0I �

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi2 � w

R0 � T0I

sð6Þ

dm/dt = mass flow rate, Aeff = effective flow area, p0I = port upstream staticpressure, T0I = port upstream static temperatrue, R0 = gas constant

3.1.7 Combustion Model

dx

da¼ a

Dac� mþ 1ð Þ � ym � e�a�y� mþ1ð Þ ð7Þ

dx ¼ dQ

Q

y ¼ a� ac

Dac

Q = total heat amount received, a = Crank Angle Degree, a0 = correspondingangle for begining of combustion, Dac = combustion duration, m = form coeffi-cient, a = 6.9, Wiebe coefficient

3.1.8 Burnt Fuel Quantity

x ¼ 1� e�a�y�ðmþ1Þ ð8Þ

3.2 Realizing Virtual Model of the Engine

Main parameters of the engine, used in simulation, are presented in Table 1.In Fig. 1 is presented the virtual model of the engine which consists in picto-

grams that have a specific function beginning with fluid properties, engine defi-nition and finishing with in cylinder parameters definition. In this simulation thevariability characteristic of the turbine was not used, also the EGR was kept at zerovalue.

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Table 1 Simulation’s parameters

Parameter Value

Stroke 86.5 mmBore 86.5 mmCompression ratio 18Connecting rod length 150 mmEngine speed 3,000 rpmEngine load 100 %Injection duration 0.0016 sInjection timing 15� CADInjection pressure 1,000 barCombustion model WiebeIntake valve opening BTDC 46.5� CADIntake valve opening duration 277� CADExhaust valve closing ATDC 39.5� CADExhaust valve opening duration 257� CADTurbocharging pressure 1.35 barTurbocharger speed 148,500 rpmAtmospheric temperature 21 �CAir temperature after the intercooler 85 �CAtmospheric pressure 1 bar

Fig. 1 Virtual model of the four cylinder turbocharged diesel engine (Adapted from [7])

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3.3 Presenting Valve Timing Modification

In Fig. 2 is presented normal valve lift for exhaust valve, discontinuous black line,and valve lift for intake valve, continuous red line. In Fig. 3 is presented normalsituation, as in Fig. 2, but with two curves in plus which have the next meaning:normally the intake valve opens 46.5� CAD BTDC—red line, green line indicatesopening of intake valve at 0� CAD BTDC and the last line, blue one indicatesintake valve opening -37.5� CAD BTDC. It may be observed that by delaying theintake valve opening the valves overlap gets smaller and smaller.

4 Results and Conclusions

4.1 Parameter Influences

4.1.1 Intake Mass Flow Rate

There are three cases:First case:Due to the fact that intake valve opens 46.5� CAD BTDC (Before Top Dead

Centre), when exhaust process it is not finished yet, burnt gases pressure p2 issuperior to p1 pressure of the air, this is the reason for which it may be observed abackflow of gases from cylinder to the intake pipe. This trend of gases is stopped

Fig. 2 Normal valve lift

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when the piston reaches Top Dead Centre, after when it goes towards BottomDead Centre there appear positive values of the intake flow within the cylinderwhich reach maxim value of approximately 80 [g/s].

Positive values are encountered even in compression stroke but only untilpressure in cylinder has the same values as pressure p1 of fresh air. When p2 getsbigger than p1 then it may be observed a second tendency of fresh air to backflow,red curve, which will stop when the intake valve will be completely closed.

Second case:In this case the green curve shows the evolution of intake mass flow when the

intake valve opens with 0� CAD BTDC. Due to the fact that there is no advancefor the valve to open there will be no backflow. Maxim value of the intake massflow close to 90 [g/s] and has the next reason: by opening the intake valve right inTDC, air pressure is equal with pressure in the cylinder, but only for a momentbecause now the piston goes to BDC which decreases p2 in such a manner thatmass flow increases rapidly. Another reason would be air speed.

In Fig. 4 it may be observed that by the end of intake stroke, which is extendlong in compression stroke than first case, mass flow decreases dramaticallybecause of the backflow. Beside this, the area under the curve is smaller than in thefirst case, this shows that inducted air is smaller in quantity.

Third case:In this case intake valve opens late, with 31.5� CAD ATDC. Due to this large

delay there is no backflow, but the mass flow reaches its maximum flowapproximately 135 [g/s]. This is motivated by the depression caused from pistonmotion towards BDC. When the piston turns towards TDC the backflow appearseven sooner due to late extension of intake process in the compression stroke. Thearea under the blue curve is the smallest, hence a proportional inducted fresh air.

Fig. 3 Modification of opening timing for intake valve

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4.2 Pressure in Cylinder

Pressure in cylinder is proportional with volumetric efficiency and it is demon-strated by the next formula:

P ¼ N � Vh � qa � F=A

� �� QHV � gf � gv ð9Þ

P = engine power, N = engine speed, Vh = cylinder displacementqa = air density, (F/A) = fuel–air ratio, QHV ¼ fuel lower heating value,gf = combustion efficiency, gv = volumetric efficiencyVolumetric efficiency depends on intake mass flow. Thus, in order to keep air–

fuel ratio equal to 1, for a less quantity of air inducted in the cylinder there is aproportional quantity of fuel which leads to less pressure in the cylinder and brakepower. This is how the descending trend of pressure from approximately 130 bar,to 110 bar and 70 bar may be motivated (Fig. 5).

This statement is supported also by pumping losses presented in Fig. 6. Themain argument according to which pumping losses area increases with the lateopening of the intake valve is that the dimension of the backflow occurring incompression stroke needs a proportional work effectuated by the piston.

Fig. 4 Intake mass flow rate

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Fig. 5 Pressure in cylinder

Fig. 6 Pumping losses

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4.2.1 Conclusion

It is obvious the fact that by opening late the intake valve the air inducted in thecylinder is also influenced. Thus, the best strategy to obtain an efficient engine isopen late the intake valve but also to close it near BDC so that will not be extrapumping work and hence less power available for the engine.

References

1. Pulkrabek WW (2003) Engineering fundamentals of the internal combustion engine. PrenticeHall, New Jersey

2. Bohac SV, Assanis DN (2006) Effect of exhaust valve timing on gasoline engine performanceand hydrocarbon emissions. SAE international 2006, variable vale train systems technology,pp 19–29

3. Leroy T, Chauvin J, Petit N (2009) Motion planning for experimental air path control of avariable valve timing spark ignition engine. Control engineering practice, vol 17. Elsevier,pp 1432–1439

4. Milovanovic N, Chen R, Turner J (2004) Influence of variable valve timings on the gasexchange process in a controlled auto–ignition engine. IMechE Part D, 2004, pp 567–583

5. Lancefield T, Methley I, Räse U, Kuhn T (2006) The application of variable event valve timingto a modern diesel engine. SAE international 2006, variable vale train systems technology,pp 229–241

6. Heywood JB (1988) Internal combustion engine fundamentals. McGraw Hill, New York7. Amesim 10, user guide pdf book8. Boost 2010, user guide pdf book

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