Review Article Comparison of Engine Simulation Software ...downloads.hindawi.com/journals/mse/2013/401643.pdf · e in-house package, Matlab Engine Simulation Code (MESC), has been

Hindawi Publishing CorporationModelling and Simulation in EngineeringVolume 2013, Article ID 401643, 21 pageshttp://dx.doi.org/10.1155/2013/401643

Review ArticleComparison of Engine Simulation Software for Development ofControl System

KinYip Chan, Andrzej Ordys, Konstantin Volkov, and Olga Duran

School ofMechanical andAutomotive Engineering, Faculty of Science, Engineering andComputing, KingstonUniversity, Friars Avenue,Roehampton, London SW15 3DW, UK

Correspondence should be addressed to KinYip Chan; [email protected]

Received 14 March 2013; Accepted 29 May 2013

Academic Editor: Luis Carlos Rabelo

Copyright © 2013 KinYip Chan et al. This is an open access article distributed under the Creative Commons Attribution License,which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

Most commonly used commercial engine simulation packages generate detailed estimation of the combustion and gas flowparameters. These parameters are required for advanced research on fluid flow and heat transfer and development of geometriesof engine components. However, engine control involves different operating parameters. Various sensors are installed into theengine, the combustion performance is recorded, and data is sent to engine control unit (ECU). ECU computes the new set ofparameters to make fine adjustments to actuators providing better engine performance. Such techniques include variable valvetiming, variable ignition timing, variable air to fuel ratio, and variable compression ratio. In the present study, two of the commercialpackages, Ricardo Wave and Lotus Engine Simulation, have been tested on the capabilities for engine control purposes. Thesepackages are compared with an in-house developed package and with reference results available from the literature. Differentnumerical experiments have been carried out from which it can be concluded that all packages predict similar profiles of pressureand temperature in the engine cylinder. Moreover, those are in reasonable agreement with the reference results while in-housedeveloped package is possible to run simulations with changing speed for engine control purpose.

1. Introduction

Lately, control engineers in the automotive industry havebeen challenged with the task of improving fuel consump-tion and engine performance while simultaneously reducingpollutant emissions. Engines fitted with electronic controlprovide flexibility to adjust the engine parameters to achievecertain performance. However, it has to be noted that thoserequirements can lead to contradictory constrains to besatisfied by the control algorithms.

The main component of the electronic control system isthe engine control unit (ECU), an electronic device which isconnected to the engine sensors and actuators to calculatethe best parameters needed to perform efficient combustion,and hence provide cleaner exhaust gas and better engineperformance [1].

Sensors connected to ECU in Figure 1 may be used inopen loop strategies, such as air-flow sensor, air temperaturesensor, and throttle position sensor. In open loop configura-tion, the ECU adjusts the engine parameters according to the

sensors outputs with the help of look-up tables. Other types ofsensors, such as lambda sensor, engine temperature sensor, orknock sensor are used on closed loop control configuration.The engine parameters, such as the ignition timing, arethen continuously updated according to the feedback of theparticular sensor.

In order to develop suitable controls strategies, simula-tion has been a valuable tool to predict and optimise keyparameters in the engine control system. It allows flexible andlow cost development of control algorithms without the needof engine bed. Once the algorithm has been developed andtested, it may be installed into the ECU.

Various commercial packages have been developed andare available to solve engineering problems related to designand optimisation of internal combustion engines (ICEs).There are four primary engine simulation commercial pack-ages used in the automotive industry today: Ricardo Wave(RW), Lotus Engine Simulation (LESoft), AVL fire, and GT-Power. These packages are similar in purpose and func-tionality. They require detailed input parameters to simulate

2 Modelling and Simulation in Engineering

AirThrottle

Throttlepositionsensor

Fuel

Fuel

Fuel

E.C.U.

Air flow/temperature

sensor

Fuel

Knock sensor

Feedback circuitsOpen-loop

Engine speed and

Engine temperaturesensor

position sensorstankinjector

Pressureregulator

ValveIgnition

Coil BatteryIgnitionValve timingprofile

timing

Mixture Engine Exhaust gas

Lambdaexhaust gas

oxygen sensor+ Exhaust

distributor

Figure 1: Block diagram of engine control system.

the engine operation in an integrated manner rather thanusing different subsystems.

LESoft is an in-house code developed by Lotus Engi-neering. The package processes engine simulation in twomodules, the data module and the solver module [2]. Thedata module allows user to input the engine dimension data.The solver module is a built in combustion and heat transferzero-dimensional equations and fuel/gas composition solveraccording to user input data in data module. The code is ableto predict gas flow, combustion and overall performance ofICEs.

RW is an engine simulation package designed to analyzethe dynamics of pressure waves, mass flows, energy lossesin ducts, plenums, and manifolds of various systems andmachines [3]. RW provides simulation of time-dependentfluid dynamics and thermodynamics using two-zone model.

However, software costs generally prohibit use in smallorganizations, primarily making them industry-specific soft-ware packages. Moreover, commercial software packages arebased on computation fluid dynamics (CFD) since they aredesigned to improve mechanical aspects of the engine. Whilemechanical models are used to calculate the moving partsof the engine in order to obtain the engine torque andacceleration, control models are used to allow calculations infeedback control schemes to optimise engine performance,such as variable valve timing, ignition timing, air to fuel ratio,and other variable engine geometries systems. Open sourcepackages have been developed by small research groups tosolve specific issues.

The in-house package, Matlab Engine Simulation Code(MESC), has been developed in the Kingston Universityfor solution of optimisation problems related to develop-ment and operation of ICEs. The code solves a full enginecycle with compression, combustion, expansion, exhaust, andintake strokes in a single cylinder model. User is able to

specify cylinder geometry and input data such as air andfuel composition, valve and ignition timing, engine speed,parameters for heat transfer, and heat release submodels.Thesolver is based on two-zone model and uses the first law ofthermodynamics and real-gas law to calculate pressure, tem-perature, heat flux, and other functions. On top compared tocommercial software packages, the code givesmore flexibilityon engine control using various fuel compositions.

The purpose of this study is to investigate the capabilitiesof commercial engine simulation packages and in-houseMatlab code for control optimisation problems.Theflexibilityof the different packages is investigated, and the accuracy ofICE models is compared.

The paper is organised as follows. The main parametersneeded for engine control are described in Section 2. Thesubmodels behind the simulation packages are described inSection 3. Comparison of different packages under particularconditions and software simulation results are presented inSection 4. Finally, conclusions are drawn in Section 5.

2. Variables for Engine Control

During the compression stroke in the engine, the mixturerises to high temperature and maximum pressure. ECUcontrols the pressure and temperature thought the amount ofair intake, the amount of fuel injection, the amount of exhaustgas recirculation, the injection timing, and the ignitiontiming if applied. There are some researches dedicated tomodifications of cylinder geometries by the aid ofmechanicalparts to vary the compression ratio when needed [4, 5]. Thissection explains the current advances in this area.

2.1. Operating Parameters. Performance of ICEs dependson several geometric and thermodynamic parameters. Thevolume of the combustion chamber is calculated from

Modelling and Simulation in Engineering 3

the dimension of the cylinder. The cylinder is assumed tobe a fixed bore size and a stroke length. When the pistonis moving, volume is calculated according to crank-slidermodel which depends on the connecting rod length as afunction of crank angle.The piston movement is also a factorto estimate heat transfer, blow-by losses, and energy lossesdue to friction.The combustion performance is treated usingthermodynamic calculations with heat release submodels.

2.2. Avoid Autoignition and Engine Knocking. Autoignitionhappens when the mixture reaches autoignition temperatureand ignites in the different part of combustion chamberwhile spark plug ignites, causing more than one frameenvelope generate in the combustion chamber. Resultantthe energy needed does not locate in the designed positionpush the piston. Different fuel compositions have differentautoignition temperatures, for example, methane is 540∘Cand for diesel fuel is 260∘C. Engine control is to be designedto the best performance while avoiding autoignition. Engineknocking happens when the engine fires a spark at the wrongtiming with the incorrect amount of air to fuel ratio. It affectsthe engine performance and can cause damages to the engine.The latest technology of knock sensor can estimate whichcylinder may knock and thus retard the ignition angle for theparticular cylinder. This can be achieved through the use ofknock control. The knock control system consists of knocksensor, signal processing unit, and the knock controller,which is integrated in ECU.

2.3. Intake Manifold Length and Volume. Geometry of intakemanifold affects the flow of air/fuel mixture. The intakemanifold length can be changed in some engines usingadjustable mechanical parts [6, 7]. It affects the mixtureturbulence inside the combustion chamber and hence affectspressure and the burning velocity duration.

2.4. Start of Combustion. In term of engine control, sparkplug can be controlled by ECU to decide whether themixtureneeds to be ignited by spark. It gives more flexiblity toan engine with spark ignition and compression ignitioncombined, according to the engine condition.

The maximum pressure developed during combustionshould occur at about 10 to 40 degrees beyond TDC depend-ing on engine speed and engine load. The burning gastakes a comparatively long time to burn; therefore, the sparkmust be timed to occur well before TDC. Correct timingof ignition is important to give the highest engine powerwithout knocking. If the timing of spark is too early, a rapidburning results engine knocks; conversely, a late spark causesslow burning and result in low engine power and poor fueleconomy. The optimal moment of ignition is a function ofseveral variables, such as the engine speed and the engineload.

2.5. Valve Timing and Valve Lift. The valves opening in thecylinder determine the amount of gas that may enter (intakevalve) or exit (exhaust valve) to or from the combustionchamber [8]. It also determines the flow rate, and hence

the pressure and temperature while combustion occurs. Theeffective compression ratio can be affected by the valvestiming. In old engines, the duration of valves opening isfixed due to geometries of mechanical parts, that is, camshaftand valve lobe. Research has shown that engines performdifferently at different rotation speeds and different loads [9–12]. Current variable valve timing systems (VVT) are limitedto designed valve profile due to the dimension of lobe. Thereare twomainmethodologies in camVVT, camphasingwhichuses gear to shift the valve opening angle, and cam changingwhich shifts to different valve dimension when needed. Thevalve profile curve leads to a one-off open and close action.Electronic camless valve is under research for more valvetiming flexibility for future use.

2.5.1. Reaction on Compression Ratio Using Different ValveTiming. The intake valve opening timing affects the effectivecompression ratio. The intake valve opening is designed tohave the highest compression ratio in specification. Withearlier or later intake valve closing, compression ratio issmaller than the specification. The amount of air or air/fuelmixture that enters into cylinder depends on the lift ofthe valve where maximal lift is reduced when the openingduration is reduced, and hence maximum volume is reduced.The compression ratio is an important parameter of enginetorque.

2.5.2. Valve Overlapping. Valve overlapping happens whenboth intake and exhaust valve open for the same period [9],Figure 2. Typically, the period before the piston reaches theTDC in exhaust stroke and after the pistonmoves beyond theTDC in intake stroke. The advantage for valve overlappingis continuous variable valve timing (VVT) which providesadaptive adjustment to improve engine torque delivery acrossdifferent revolution range [11, 13]. When the engine speedis high, more air is needed for higher energy combustion;therefore, duration of intake valve opening needs to belonger. A portion of exhaust gas will suck back to thecombustion chamber while the exhaust valve is opening,and hence increase the mixture pressure and temperature. Atpresent, several vehiclemanufacturers offermechanical VVT,by means of cam phasing VVT. In this approach, the gearattached to the valve moves the position to change the twomechanical lobes and hence switch to different profiles.Theseprofiles correspond to the amount of valve lifting.

Negative valve overlapping (NVO) consists of closingboth valves for a period of time [5–8]. Exhaust valve closesbefore piston reaches the TDC, and intake valve opens afterpiston moves beyond the TDC. The concept is similar to therole of the exhaust valve in exhaust gas recirculation systems.When the intake valve opens, the hot residual gas heats up thefresh mixture, and the high temperature mixture is enhancedfor compression ignition.The results, presented in [15], showthat the injection timing is an important parameter. The newfuel injection strategy by injecting a portion of fuel during thenegative valve closing interval and inject the rest in the intakestroke was also examined in [15].


Intake Compression

CombustionExhaust

(1) Intake valve

(2) Intake valve closes, compression starts

(4) Exhaust valve opens

(5) Exhaust valve closes

Valve overlapping(3) Piston at T.D.C., combustion happens

Figure 2: Valve overlapping in term of piston position.

2.5.3. Late Intake Valve Closing on Diesel Engine. Anothercase on other VVT approach is late intake valve closing(LIVC) which is found in diesel engine research. Thereare some researches investigating the performance on LIVCon diesel engines [16, 17]. The idea behind LIVC is toburn the mixture in a lower combustion temperature andhence to reduce nitrogen oxide (NOx). However, as a result,concentration of hydrocarbons (HC) and concentration ofcarbon oxide (CO) increase. Then, emissions are reduced atlow tomedium engine loads and also less soot and particulatematter.

2.6. Exhaust Gas Recirculation. Exhaust gas recirculation(EGR) is based on the idea of sending a portion of the exhaustgas back to the cylinder for reburning, Figure 3. It can providefurther control of the temperature and pressure of the freshmixture by mixing it with exhaust gas. EGR system connectsthe exhaust manifold before the throttle. Fresh air is mixedwith exhaust gas through the throttle and consequently thetemperature of the gas rises.The amount of EGR is controlledby ECU. Research shows that the temperature of air is heatedup by exhaust gas, which acts as a preheater before enteringthe combustion chamber at low speed. The effect is cancelledout at high speed and low load due to the decline in thetemperature cycle, and hence no improvement is made [8].Also, it affects significant improvement of performance suchas NOx emission, engine break power, and fuel consumption.In [18], significant improvements were achieved in terms ofperformance by using 5% of EGR at low.

2.7. Compression Ratio. The compression ratio typicallyranges from 8 to 12 for spark-ignition ICEs, and it is inthe range of 12–24 for compression-ignition ICEs. Whencompression ratio is high, the fuel and air particles are closelypacked together and obtain a higher explosive velocity duringcombustion, that is, better performance and better fuel

consumption. However, it is limited by two considerations:material strength due to maximum stress at certain pressureand engine knock due to the maximum temperature duringprocess exceeds autoignition temperature. There is someresearch on variable compression ratio systemwhich changesgeometries of combustion chambers [4, 5, 19–21].

2.8. Length of Intake Manifold. Airflow is one of the mainfactors to simulate combustion. The pipe geometries controlthe air flow and affect the pressure and temperature ofair/fuel mixture inside engine cylinders. Software treats theair flow as one-dimensional. The amount of air and the flowvelocity entering the cylinders depend on its shape (circularor rectangular). In order to simplify the computational mesh,the duct or pipe is divided into discredited length where userscan modify this parameter according to the duct length.

3. Submodels

A number of submodels are needed to provide a descriptionof all relevant processes (burned gas composition, massburning rate, heat transfer, blow-by, and others). Enginesubmodels are divided into various categories and includesubmodels for cylinder geometry and piston motion, com-bustion sub-model, friction and heat transfer submodels, andsome others. Descriptions of submodels are based on datapresented in user’s guides [2, 3] and the literature [22, 23].

3.1. Engine Cycle. Engine cycle simulation model follows thechanging thermodynamic and chemical state of the workingfluid through the engine’s intake, compression, combustion,expansion, and exhaust processes.The starting point for thesecycle simulations is the first law of thermodynamics for anopen system. The structure of this type of engine simulationis indicated in Figure 4. During each process, submodelsare used to describe geometric features of the cylinder and


Air cleaner Air flowsensor

Throttlevalve

Inlet valves Intakemanifold

manifold

Catalyticconverter

Pressure

Fuel injectorsFuel filter

Fuel

Fuel pumpFuel tank

Air

regulator

Exhaust gas

Exhaust gas

Exhaust

recirculation

Cylinders

Mixture

O2 sensor

Figure 3: Air and fuel flow in ICE.

Intake

Thermodynamic analysis Phenomenologicalprocess models

Cylinder andvalve geometry

Thermodynamicproperties

Flow rate

Head transferTransportproperties

Combustion rate

Emissionsmechanism

of cylinder processes

Compression

Combustion

Expansion

Exhaust

Figure 4: Structure of thermodynamic based simulation of ICEoperating cycle.

valves/ports, the thermodynamic properties of unburned andburned gases, themass and energy transfers across the systemboundary, and the combustion process.

During intake and compression, the working fluidcomposition is frozen. The composition and thermody-namic properties are determined using the submodels. Massflows across open valves are usually calculated using one-dimensional compressible flow equations for flow througha restriction or filling and emptying models [22, 23]. Themore accurate unsteady gas dynamic intake and exhaust flowmodels are sometimes used to calculate themass flow into theengine cylinder in complete engine cycle simulations when

the variation in engine flow rate with speed is important (thedisadvantage ismuch increased computing time). Heat trans-fer during intake and compression is calculated using oneof the Nusselt-Reynolds number correlations for turbulentconvective heat transfer [22, 23]. The transport properties,viscosity, and thermal conductivity used in these correlationsare obtained from semiempirical relations.

Many approaches to predict the burning or chemicalenergy release rate have been used successfully to meetdifferent simulation objectives [2, 3, 22, 23]. The simplestapproach has been to use a one-zone model where a singlethermodynamic system represents the entire combustionchamber contents, and the energy release rate is defined byempirically based functions specified as part of the simulationinput. Quasi-geometric models of turbulent premixed flamesare used with a two-zone analysis of the combustion chambercontents (an unburned and a burned gas region) in moresophisticated simulations of ICEs [22].

The expansion process is either treated as a continuationof the combustion process or, once combustion is over, usesthe form of themass, fuel, and energy conservation equationswhich hold during compression. The exhaust process usesconservation equations for a one-zone open-systemmodel ofthe cylinder contents.

3.2. Cylinder Geometry. Geometric sub-model is used todefine the topology of the combustion chamber as well as tospecify geometric surfaces and volume zones used for heattransfer and combustion submodels. The cylinder volumedepends on the engine geometry (bore, stroke, connectingrod length, and compression ratio) and is a function of crankangle and the rate of change of cylinder volume.


The model used specifies the cylinder geometry as apiston, head, and cylindrical liner [22]. Both the piston andhead surface area are assumed to be equivalent to the areaderived from the bore diameter with multipliers of this areaallowed in the Woschni heat transfer model.

3.3. Piston Motion. Piston motion sub-model is used todetermine the instantaneous position of the piston in thecylinder and the volume of the combustion chamber as afunction of crank angle. The piston position is fed to othersubmodels.

The crank-slider model is used to characterize thearrangement of mechanical parts designed to convert trans-lational motion into rotational motion. The piston motionis defined by including geometric inputs for the cylindricalcombustion chamber, crankshaft, and connecting rod. Theposition of piston is derived from basic trigonometry equa-tions [22].

3.4. Air and Fuel Properties. Air consists of 21% oxygenand 79% of nitrogen (for each mole of O

2there are 3.76

moles of N2). Selected physical properties of oxygen and

nitrogen are given in the textbooks and tables [23]. Extensionof the analysis to different air mixtures encountered inpractice is straightforward. The most frequent differencesare the presence of water vapour, carbon dioxide, and argonin the atmospheric air. The amount of water vapour inthe atmospheric air normally depends on temperature anddegree of saturation.

In RW and MESC calculations, the fuel is a user-definedcomposition of C, H, O, and N atoms (the most commonhydrocarbon fuels have no O or N unless mixed with ethersor alcohols). In LESoft calculations, the fuel types are limitedto those composed of C, H, and O atoms [2].

Combustion products are a mixture of some gases.The gas properties are based on the appropriate govern-ing relations, for example, perfect gas equations for air-only thermochemistry or ideal/real gas equations for thethermochemistry of hydrocarbon/air mixture for generalC/H/O/N fuel types. The thermodynamic properties of theindividual gases are calculated as functions of temperaturewith these properties being averaged as molar functionsto give the overall properties of the mixture [22]. The gasproperty sub-model is based on polynomial curve fits tothermodynamic data for each species [22]. In a real gasmodel, the compressibility factor is calculated using empiricalcorrelation [3]. For a gasmixture, the critical temperature andpressure and acentric factor are calculated according to themixing rule. The real-gas enthalpy and internal energy arecalculated from their ideal-gas counterparts, along with thespecific heats at constant pressure and volume.

Dissociation effect on effective heat release is developedthrough the use of maldistribution factor [2]. The mald-istribution factor is incorporated to allow for a reductionin effective caloric value of the fuel due to poor chargemixing and dissociation. A factor of 0 implies almost perfectmixing and a high effective caloric value for the fuel. Atypical gasoline engine would have a maldistribution factor

of between 1.0 and 3.0. Values less than 1.0 imply bettercombustion and may be appropriate for gas fuelled engines.

The coefficients of chemical reactions are determinedfrom the equilibrium calculations [23]. Fuel evaporation andcondensation are not taken into account in the calculations(these submodels are available in RW).

3.5. Heat Release Model. The amount of heat released duringcombustion transfer to energy pushes down the piston in anappropriate time. Burn rate is defined as the rate at whichthe fuel mass in the cylinder is consumed in the combustionprocess to become products of combustion. Regardless of thefinal state of the air/fuel mixture (emissions products, freeradicals), the initial fresh air and fuel no longer exist in theirnatural states when combustion has completed.

The heat release rate is calculated using empirical heatrelease function derived from theWiebe equation.TheWiebemodel is widely used to describe the rate of fuel mass burnedin thermodynamic calculations, and this model is available inall packages. TheWiebe function is [22]

𝑥𝑏= 1 − 𝑎 exp [(

𝜃 − 𝜃𝑠

𝜃𝑑

)

𝑛

] , (1)

where 𝑎 is the range cover 10%–90% of mixture burned and𝑛 is the efficiency of combustion.The combustion duration isdefined as the number of crank degrees between 10% and 90%of mass fraction burnt. The combustion phasing is defined asthe number of crank degrees after TDC firing at which 50%of the fuel has been burnt.

Wiebe model allows the independent input of functionshape parameters, combustion duration, and combustionphasing which are functions of the type of fuel being used[22]. It is known to represent quite well the experimen-tally observed trends of premixed combustion. By adjust-ing the individual Wiebe function component shapes andtheir relative proportions through the fuel mass fractions,a complicated overall burn rate curve is defined. Varyingthe 50% burn point simply shifts the entire curve forwardor backward. Varying the 10%–90% duration extends thetotal combustion duration, making the profile extend longeror compress shorter. Varying the Wiebe exponent shifts thecurve to burn mass earlier or later.

The two-partWiebe function takes into account the massfraction burned in the premixed combustion period and themass fraction burned during the diffusion combustion period[2, 3].

As an alternative to Wiebe combustion models, the usermay choose to enter a complete profile of combustion [3].This way allows the user to enter a cylinder pressure profileobtained from testing, internally derive a heat release profileusing the applied heat transfer model and thermodynamicproperties, and output a fuel mass burn profile for combus-tion calculations.

A more advanced sub-model, turbulent flame combus-tion model, calculates the rate of fuel burned using theturbulent flame speed and of instantaneous flame area [22].It responds to changes in cylinder flows and in combustion


chamber geometry. The model also provides spatial resolu-tion as to the instantaneous flame location needed for heattransfer calculations.

A combustion model follows a constant combustion airto fuel ratio under the assumption that the air/fuel mixtureis fully premixed. In certain engine designs (e.g., SparkInjection Direct Ignition engines), the fuel and air are notfully mixed at the start of combustion. The spark-ignitionstratified sub-model is used in [3] to impose a combustionequivalence ratio that is different from the overall equivalenceratio of the cylinder, for the purpose of simulating a stratifiedchargemixture.Thismodel allows combustion stoichiometryto vary during combustion and is used to improve accuracyof results for systems that add or remove air or fuel duringcombustion, since the system will respond more accuratelyto changes in cylinder contents. The spark-ignition stratifiedsub-model creates an adaptive mixture which begins at auser-entered equivalence ratio and adjusts as burn progressesto consume the available air.

3.6. Conduction. Conduction sub-model is used to calculatecylinder surface temperature. Accurate surface temperatureimproves boundary conditions for the in-cylinder heat trans-fer sub-model and is used to assist in engine componentdesign. One conduction sub-model is available in MESC(it is referred to as “simple”), two conduction submodelsare available in LESoft [2] (they are referred to as “simple”and “conduction”), and three submodels are implemented inRW [3] (they are referred to as “simple,” “conduction” and“swing”).

The “simple” engine conduction sub-model uses a prede-fined thermal network which represents the cylinder liner,cylinder head, piston, and intake and exhaust valves. Thismodel enables the prediction of the surface temperature ofthe combustion chamber and heat rejection of the engine tothe coolant. This model is offered by all packages.

Other conduction submodels, offered by LESoft andRW, use a more detailed predefined thermal network than“simple” model.

In the “conduction” sub-model used in LESoft and RW,the cylinder wall temperature is specified or calculatedvia a simple one-dimensional heat transfer calculations forcylinder head, piston and linear walls. The cylinder walls areassumed to have a wall thickness that is directly proportionalto the bore. The heat transfer coefficients of coolant andcylinder head are constant value, and conductivity ofmaterialis specified by choosing material used. The cylinder headtemperature is calculated as the area average of the walltemperature and the valve temperature. Valve head tem-peratures are calculated for both inlet and exhaust valvesas a function of fuel type and air to fuel ratio. The pistontemperature is assumed that it is equal to the area averagedcylinder head temperature. In the RW, this sub-model hasmore spatial resolution and also allows substituting portionsof the predefined mesh with existing finite element modelsbuilt using other commercial software [3].

The “swing” sub-model from the RW accounts for cyclictemperature transients on combustion chamber surfaces and

other engine surfaces exposed to gases, which are producedby the peaked nature of the gas-wall heat fluxes [3]. Thisis especially relevant for in-cylinder surfaces with low con-ductivity where temperature fluctuation is quite high (100–200K).

The calculations presented are based on “simple” con-duction model, and cylinder wall temperature is fixed in allcalculations.

3.7. Heat Transfer Model. Two non-dimensional parameters,the Nusselt number and the Reynolds number, are used toestimate the heat transfer correlated with fluid conditions.Those two parameters vary in time, therefore heat transferto and from the cylinder gases are calculated at every crankangle increment. These calculations require knowledge ofwall area, wall temperatures, and surface heat transfer coeffi-cient. The cylinder surface areas are calculated using surfacearea to bore area ratio which are a function of combustionsystem. The linear area is calculated at each increment bythe piston displacement from TDC. Gas velocity dependson piston mean speed, amount of swirl, the amount ofcombustion and the gas turbulence [22].

Various heat transfer submodels are offered by eachpackage for the calculation of the convective heat transfercoefficient in the engine cylinder [2, 3]. The heat transfermodels proposed by Annand, Woschni and Eichleberg areavailable in LESoft and MESC. Heat transfer models offeredby the RW include Woschni, Annand and model referred toas “flow” model [3].

Woschni, Annand, and Eichleberg models are derivedfrom a basic Nusselt number correlation for flow in pipes[22, 23]. Each model employs coefficients that have beendeveloped to best reproduce the heat transfer results obtainedby experiment. The Woschni heat transfer sub-model is themost commonly used heat transfer model. It views the chargeas having a uniform heat flow coefficient and velocity onall surfaces of the cylinder, and calculates the amount ofheat transferred to and from the charge based on theseassumptions.The Annand and Eichleberg heat transfer mod-els are typically used as a mean of comparison for the resultsfrom the Woschni model. They are rarely used in practicalapplications [23]. Different from the Woschni model, wherethe increase in gas velocity in the cylinder during combustionis accounted for, the Annand model assumes a constant gasvelocity equal to the mean piston speed.

TheWoschni heat transfer correlation is [22]

Nu = 0.0035Re0.8. (2)

The heat transfer coefficient is

ℎ = 128ℎ𝑐𝑏−0.2

𝑝0.8

𝑇−0.53

(𝐶1𝑢𝑝𝑚+ 𝐶2𝑇𝑟

𝑉

𝑉𝑟

𝑝 − 𝑝𝑚

𝑝𝑟

)

0.8

, (3)

where 𝑏 is the bore (m), V is the cylinder volume (m3), pis the cylinder pressure (kPa), T is the cylinder temperature(K), 𝑢

𝑝𝑚is the mean piston velocity (m/s), and 𝑝

𝑚is the

motored cylinder pressure (kPa). Subscript 𝑟 correspondsto the reference pressure 𝑝

𝑟(kPa), reference temperature


𝑇𝑟(K) and reference volume𝑉

𝑟(m3). User specific multiplier

constant ℎ𝑐is 1. The constant 𝐶

1is 2.28 for combustion and

is changed to 6.18 for scavenging.The constant𝐶2defines the

combustion term. Its value is 3.24 × 10−3 in closed cycle, and0 before combustion and scavenging.

The “flow” model, implemented in RW, predicts convec-tive heat transfer from gases to combustion chamber walls,and its spatial and cyclic temporal variations are driven by theinstantaneous mean and turbulent gas motions [3]. Predic-tion of heat transfer coefficient is linked to the fluid dynamicsof combustion gases, determined by periodic piston motion,valve flows, injection process, combustion, motion of pistonand chamber geometry. The multi-zone flow model solvesdifferential equations for swirl and turbulence and an alge-braic equation for squish in a number of combustion chamberflow zones. The heat transfer coefficient for each surface isobtained from the Colburn analogy of heat and momentumtransfer and is directly linked to the effective velocities. Oneof the main benefits of this approach is that it captures thedependence of the heat transfer coefficient on the magnitudeof instantaneous in-cylinder flow velocities. Another benefitis the improved ability of the model to represent the spatialnonuniformity of the heat transfer coefficients.

Thermal radiation fromgases to surrounding combustionchamber surfaces is a significant component of heat transferfor diesel engines, due to soot formation during diesel com-bustion [22]. Instantaneous andmean levels of heat radiationare functions of the volume and distribution of burninggas, amount of soot present in the burning gas, combustionchamber geometry, and surface emissivity and temperature.The submodels employed in the radiation model calculatethe soot loading in the burned zone, volume of the radiatingburned zone and radiation temperature.

3.8. Friction. The Chen-Flynn friction model is used tocalculate the work done by friction in the engine andaccompanying systems [2, 3]. Friction model takes intoaccount engine speed,mean piston speed,maximumcylinderpressure and compression ratio. The Chen-Flynn correlationhas a constant term (for accessory friction), a term whichvaries with peak cylinder pressure, a term linearly dependenton mean piston velocity (for hydrodynamic friction), anda term quadratic with mean piston velocity (for windagelosses).

3.9. Blow-By. Blow-by occurs when a portion of gas escapesthrough the gap between piston and cylinder wall into thecrankcase [22]. The consequence is the pressure inside thecylinder decrease by a portion.This is also takes into accountwhen sub-model calculates the enthalpy. The blow-by modelcalculates howmuch gas is removed from the cylinder duringthe cycle and blows by the cylinder’s rings, optionally into thecrankcase.

InMESC calculations, the fraction of blow-by is specifiedwith initial value of 0.8. In RW calculations, blow-by sub-model assumes the net gas flow past the first ring in zeroand enables the thermodynamic model for the first crevice[3]. The input parameters for the model are cold piston

liner clearance and distance of the top ring from pistontop. During compression, the air/fuel mixture is packed intothe crevice volume (absorption process) allowing unburnedhydrocarbons to escape the combustion process occurringin the main chamber. The absorbed unburned hydrocarbonsreturn to the cylinder during expansion as a source ofhydrocarbons emissions.

3.10. Emissions. Emissions submodels are used to calcu-late the in-cylinder formation and destruction of commonengine emissions species [22, 23]. Emissions submodels forthe engine include CO, HC, and NOx. The CO emissionssub-model predicts CO production during combustion andexhaust in an engine cylinder element. The HC emissionssub-model predicts HC production during combustion andexhaust in an engine cylinder element as well as in theflow network. The NOx emissions sub-model predicts NOxproduction during combustion and exhaust in an enginecylinder element.

The emissions submodels are available in all packages butthey are not used in the simulations presented.

3.11. Valve Motion. Valve motion is specified by means of afew simple parameters such as diameter, maximum lift, andduration of opening and closing ramps [22, 23]. The lift valvesub-model defines a lift profile using polynomial lift curves.The valve lift duration is specified by the number of crankdegrees between valve opening and valve closing. The valve-flowmodel assumes a steady, a diabatic, and reversible flow ofan ideal fluid through a duct.Thevalve flow in an engine is notan ideal flow, and the discharge coefficient is introduced toaccount for real gas effects. The mass flow through the valvesdepends on the effective open area, pressure ratio, upstreampressure, specific heat, upstream temperature and ratio ofspecific heats.

3.12. Scavenging. The gas exchange process in the four-strokeengine is driven primarily by the piston motion, and a fullymixed scavenging model is most commonly used [22]. Withthe perfectmixingmodel any charge gas entering the cylinderis instantaneously and homogeneously mixed with the gascurrently in the cylinder.

3.13. Network Model. Flow elements are used to constructthe flow network, which is the primary part of any model.The collection of flow elements available are organizedinto groups, representing element types. These elements aresimilar in various packages [2, 3].

Ambient element is used as flow sources and sinks in amodel and attach to the rest of the flow network via a singleduct. This is most-commonly used element to represent aninfinite reservoir at standard atmospheric conditions.

Cylinder elements is typically used to model the cylinderof an engine, however additional elements are included tomodel piston compressors and crankcases. Cylinder elementis zero-dimensional and simply represents a volume changingwith time. Cylinder connects to multiple flow elements via


valves and usually require numerous submodels in order tocorrectly represent the desired system.

Duct element represents a portion of a flow networkwhich is long (in respect to cross-sectional area) creating one-dimensional flow. It is automatically discretized into a seriesof computational cells based on a discretization length.Aductmust be connected at each end to one ambient or cylinderelement.

A fuel/air injector element injects into a cylinder enoughfuel tomeet a specified fuel/air ratio based on the total air flowthrough engine from the previous engine cycle.The achievedfuel/air ratio is more or less than the target ratio based onsome of the air or fuel exiting through the valves.

3.14. Capabilities. Cylinder and plenum are zero-dimen-sional elements in that they have properties ofmass, pressure,temperature and volume but not length. The conditionswithin these elements are calculated at each crank angle bysolving the mass equation and the energy equation derivedfrom the first law of thermodynamics.

The mass equation accounts for changes in in-cylindermass due to flow through valves, and due to fuel injection. Aseparate accounting ismade for fluxes of air, fuel andproductsof combustion. The energy equation equates the change ofinternal energy of in-cylinder gases to the sum of enthalpyfluxes in and out of the chamber, heat transfer, and pistonwork.

There are two options: single zone and two zone. Inthe one-zone model, implemented in LESoft [2], the wholecylinder is treated as one region, while in the two-zonemodel,implemented in RW and MESC, the cylinder is divided totwo regions: an unburned zone and a burned zone whichshare a common pressure [3]. The two-zone model is usedto capture in more detail the processes taking place duringthe combustion period. All combustion models may be usedin either single or two-zone model, but emissions modelsgenerally require the two-zone model [22].

Table 1 shows the capabilities of different engine simu-lating software. The submodels used in simulations are onlypresented in the table.

3.15. Development of EngineModels. The solver reads amodelfrom a main input file which contains all data required for asimulation (themodel network, initial conditions for the flowfield and all control data for the run). The solver produces anoutput file when run. The output file contains information,important in understanding the input, run-time processingand output of a model (time step output, engine summary,fuel burn progress summary, engine geometry, operatingconditions, engine cylinder heat transfer, performance). Ageneral flowdiagram for the development of amodel is shownin Figure 5.

The basic engine data include engine geometry (bore,stroke, connector rod length, compression ratio), engineinertia (mass and inertia of various components), cylinderand valve event phasing. Initial conditions such as exhausttemperatures, intake temperatures, and wall temperaturesneed to be input as reasonable values. The motion of the

Table 1: Capability of engine simulation software.

Submodel/controlledparameter LESoft RW MESC

Geometry + + +Compressionratio + + +

Air/fuel properties C/H/O C/H/O/N C/H/O/NFuel composition + + +Air/fuel ratio + + +Injection timing − − −

Heat release Wiebe/Others Wiebe/Others WiebeIgnition timing + + +

Conduction Simple Simple Simple

Heat transferWoschni Woschni WoschniAnnandEichelberg

Friction + + +Blow-by Simple Simple PercentageEmission + + +Valve motion Simple Simple Simple

Valve lift + + +Valve timing + + +

Scavenging Simple Simple Simple

piston is calculated based on the data specified for bore,stroke, and connecting rod length. The clearance volume iscalculated based on the compression ratio.

Once the basic inputs are defined, the advanced inputsneed to be defined. These inputs are port flow coefficients,valve lift per crankshaft rotation, combustion and heat trans-fer modeling (types of models to be used for representingthe combustion and heat transfer processes and the surfaceareas and temperatures of various components within thecylinder, phasing of cylinder firing with respect to TDC).Some intuition is used to determine the parameters due tothe large difference in engine operating speed and differencein engine type.

To define the operation of the cylinder head, the valve liftper incremental camshaft rotation is defined. The results ofvalve lift versus incremental camshaft rotation are related tocrankshaft rotation based on published valve opening pointswith reference to crankshaft position. Flow versus valve liftis defined. The input is in the form of flow coefficients ordischarge coefficients taken from the empirical data [22].

Once these basic and advanced inputs have been satisfied,the simulation format is specified. Convergence detectionallows the package to move on to the next case once a user-specified tolerance for convergence (tolerance is about 1%)towards a solution is attained on the active case. Simulationduration sets the number of cycles for the engine to runbefore settling on a solution. If auto-convergence is enabled,the package will move on to the next case once convergenceis reached regardless of the number of cycles specified. Ifconvergence is not reached by the number of specified cycles,


Basic input

BoreStroke

Connecting rod lengthCompression ratio

Simulation parameters

Initial conditions

Advanced input

Combustion Submodels Heat transfer

Test scenario

Varying parameter

Simulation

Validation/analysis

FrictionConduction

Valve liftDischarge coefficients

Intake geometry Internal geometry Exhaust geometry

· · ·

· · ·

· · · · · ·

· · ·

· · ·· · ·

· · ·

Figure 5: Development of engine model.

the package will output a warning to let the user know thatconvergence was not reached within desired tolerances andwill continue on to the next case.

4. Results and Discussion

Three enginemodels are designed with various packages, andengine operating parameters and submodels are chosen toperform the tests.

4.1. Scenarios of Tests. The tests with three packages, RW,LESoft, and MESC, are performed for fixed engine geometryand different engine operating parameters. Test 1 is focusedon comparison of the results computedwith various packagesand reference results presented in [23]. Ignition timing, air tofuel ratio and compression ratio are varied in theTest 2, Test 3,and Test 4, respectively. Profiles of pressure and temperatureare plotted for various tests, and discrepancies of the resultscomputed with various packages are discussed.

The engine model, used in the tests, represents a simpleone-cylinder four-stroke engine. The engine geometry and

engine operating parameters, which are fixed in the tests, arespecified in Table 2. The engine cylinder has a bore of 0.1m,stroke of 0.08m and connecting rod length of 0.15m. Theengine speed is fixed at 2000 rpm. The intake temperatureand pressure are 300K and 1 bar. The exhaust temperatureand pressure are 700K and 1 bar. Initial temperature andpressure of the cylinder are 350K and 1 bar. The fuel isgasoline (C

8H18) with an equivalence ratio of 0.8. TheWiebe

combustion model is used in all tests with parameters whichfit many experimental data [22, 23] (a = 5 and 𝑛 = 3). Theburn duration is fixed at 60 degrees although crank anglecorresponding to the start of combustion may be varied incalculations. Heat transfer in the cylinder is described usingWoschni correlation (𝐶

1= 2.28 for combustion and 𝐶

1= 6.18

for scavenging). The averaged cylinder wall temperature isfixed at 420K. Exhaust valve opens and closes at 170 and 390degrees, and intake valve opens and closes at 330 and 550degrees (valve overlapping).

The engine operating parameters, which are varied inthe tests performed, are presented in Table 3. All engineoperating parameters are fixed for the Test 1. The burning


Table 2:Geometry of engine andfixed engine operating parameters.

Parameter ValueBore, m 0.10Stroke, m 0.08Connecting rod length, m 0.16Engine speed, rpm 2000Intake pressure, bar 1Intake temperature, K 350Exhaust pressure, bar 1Exhaust temperature, K 700Fuel C8H18

Equivalence ratio 0.8Combustion model Wiebe modelCombustion duration, deg 60Heat transfer model Woschni modelWall temperature, K 420Valve timing, deg EVO 170, EVC 390, IVO 330, IVC 550

Table 3: Varied engine operating parameters.

Test Varied parameter Interval of valuesTest 1 No Fixed values

Test 2 Ignition timing,deg From −60 to 10 with step 10

Test 3 Air to fuel ratio From 10 to 18 with step 2Test 4 Compression ratio From 6 to 14 with step 2

initiation time is varied in Test 2, air to fuel ratio is variedin Test 3, and compression ratio is varied in Test 4.

Detailed verification and validation of submodelsdescribing chemical composition of combustion products,heat transfer, friction and combustion processes, andgas dynamics of intake and exhaust systems have beenperformed. The results obtained have been compared withthe reference data presented in [22, 23], and these resultsare not presented in the paper. Main attention of the studypresented is focused on predicting cylinder pressure andtemperature distributions with various packages and use ofthe engine simulation packages in development of controlsystems.

4.2. Engine Models. Three packages are used to model thecycle of one-cylinder four-stroke engine, and three enginemodels are designed with various packages. The enginemodel itself consists of the flow network designed using flowelements, submodels applied to the flow elements, and acontrol network, built using control elements.The piping andmanifolds of the intake and exhaust systems are alsomodelledusing the flow elements. The flow network and intake andexhaust networks are then linked together through engineelements and submodels.

4.2.1. Ricardo Wave. The engine model, designed with RW,is shown in Figure 6. The engine model consists of cylinder

Intake

Injector1

Duct3Duct2

CYL1

1

Exhaust

Figure 6: Engine model designed with RW.

(“cyl1” element) and intake and exhaust networks each ofwhich is designed using ambient and duct elements. Theintake network consists of “Intake” ambient element con-taining a fresh air at the temperature of 350K and pressureof 1 bar, and “duct2” duct element connecting manifold andcylinder. The exhaust network consists of “Exhaust” ambientelement containing exhaust gases at the temperature of 700Kand pressure of 1 bar, and “duct3” duct element connectingcylinder andmanifold.The fuel injector (“injector1” element)is installed in the middle of the connecting duct (“duct2”element). Fuel is mixed with fresh air and premixed air,and fuel mixture enters the cylinder. The cross-sectionalarea of intake pipe is 35mm (“duct2” element), and cross-sectional area of exhaust pipe is 28mm (“duct3” element).The maximum valve lift is 9mm, and the default lift profileis used. The friction and blow-by parameters are by default.Other engine operating parameters are specified in Table 2.

4.2.2. Lotus Engine Simulation. The engine model, designedwith LESoft, is shown in Figure 7. The engine model consistsof the cylinder (“CYL1” element), and intake and exhaustnetworks each of which is designed using three elementsrepresenting connector, port, and valve. The inlet connector(“INL1” element) contains fresh air at the temperature of350K and pressure of 1 bar. Fresh air goes through an intakeport (“PORT1” element) with the diameter of 35mm. Theintake port is connected with the intake valve (“PVAL1”element). The exhaust gases leave the cylinder through theexhaust valve (“PVAL2” element) and exhaust port (“PORT2”element) with the diameter of 28mm. The exhaust port isattached to the exit pipe (“EXT1” element), and the boundaryconditions for exhaust gases are fixed at the temperature of700K and pressure of 1 bar. The maximum valve lift for bothintake and exhaust valves are 9mm, and the lift profile is“fast lift” profile. The friction and blow-by parameters are bydefault. Other engine operating parameters are specified inTable 2.

4.2.3. Matlab Engine Simulation Code (MESC). The enginemodel, implemented with in-house package, MESC, isdesigned using flow and control elements specified in the textfile (there is no graphics user interface, as this is in-housepackage). The friction and blow-by parameters are the sameas in RW and LESoft engine models. Other engine operatingparameters are specified in Table 2.


Fuel 1

INL1 PORT1 PORT2 EXT1PVAL1 PVAL2CYL1

Figure 7: Engine model designed with LESoft.

4.3. Comparison with Reference Results. The reference test isintended to match input parameters and settings for threepackages as similar as possible. The engine models, designedwith various packages, have been validated by comparing theresults computed with reference results presented in [23].Thegeometric parameters and operating parameters of engine areshown inTable 2, and someother input parameters, which arespecific for this test, are presented in Table 4. The engine is abasic one-cylinder engine. The ignition timing is 35 degreesprior to TDC in the combustion stroke, compression ratio is10, and air to fuel ratio is 14.7.

The accuracy of the curve coefficients that are used todescribe the thermodynamic properties of the fuel, air, andcombustion products is established through internal consis-tency check and comparison to well-established propertiesat the reference conditions (T = 298.15 K and p = 1 bar).Based on these results, it is concluded that the fundamentalthermodynamic properties of the fuel, air, and combustionproduct species are in agreement with published results in[23] to within the scatter of those results (typically less than1%).

The pressure and temperature distributions in the enginecylinder, computedwith various packages, are comparedwiththe tabulated outputs from [23]. Comparisons between theresults computed and those from [23] confirm that threepackages produce reasonable data.

The pressure profiles in the compression and combustionstrokes are presented in Figure 8. The pressure distributionsare started with crank angle of −180 degrees when the pistonis at BDC. The pressure increases when the piston moves upto compress the air and fuel mixture. Ignition starts at −35degrees before TDC when the combustion model starts toestimate the heat generated during combustion.The pressurerises due to the heat added to the compressed gas. Theenergy generated in the combustion process pushes pistondown, volume increases and pressure drops. The results,predicted with RW and LESoft, are very close. The pressureprofile, predicted with MESC, has a slightly delay in themaximum pressure point (this shift is about 2 degrees), andthe maximum pressure level is about 6% lower compared toRW and LESoft results. Comparing the results obtained withthe reference results of [23], calculations give delay in themaximum pressure point (this shift is about 10 degrees for

Table 4: Input parameters for reference test.

Parameter ValueIgnition timing, deg −35Compression ratio 10Air to fuel ratio 14.7

0 90 1800

1

2

3

4

5

6

7

−180 −90

𝜃 (deg)

RWLESoftMESCReference results

p (M

Pa)

Figure 8: Comparison of pressure profiles with reference data of[23].

RW and LESoft calculations, and it is about 8 degrees forMESC calculations), although the maximum pressure level,predicted with various packages, is close to reference data(discrepancy is about 3% for all packages).

The results, presented in Figure 9, show the distributionsof temperatures of burned and unburned mixture duringcompression and combustion strokes. Ignition starts at −35degrees before TDC, when the combustion model starts towork in the burned zone. Calculations in the unburned zonecontinue without addition of heat due to combustion. Igni-tion duration is 60 degrees, and work of combustion modelis finished at 25 degrees after the TDC. The temperatureprofiles corresponding to the burned and unburnedmixturesare overlapped during the combustion interval (it is fixed at60 degrees).The calculations after combustion depend on thetemperature level reached in the burned zone. The referenceresults from [23] are plotted with the temperature step of 10degrees. The reference results show a slightly lower level oftemperature in the unburned zone and slightly higher levelof temperature in the burned zone compared to the resultscomputed with RW and MESC. Calculations with LESoft arebased on one-zone model allowing prediction of averagedtemperature in the engine cylinder. Discrepancy of the resultscomputed with various packages may be explained by the


1000

2000

3000

0 90 1800−180 −90

𝜃 (deg)

RWMESCReference results

T(K

)

Figure 9: Comparison of unburned and burned temperatureprofiles with reference data of [23].

different implementation of combustion models (there aresome parameters and settings which are not controlled bythe user) and small discrepancy in physical and chemicalproperties of fuel and combustion products.

The results, presented in Figure 10, show distributionsof temperature averaged over burned and unburned zones(averaged temperature is not presented in [23] but may becalculated using results computed in burned and unburnedzones). LESoft calculations are based on one-zone model,whereas results predicted with RW and MESC calculationsare based on two-zone models. The temperature profilesout of combustion interval, which is fixed at 60 degrees,are very close for all packages. As for the profiles of tem-perature, presented in Figure 10, main discrepancy of theprofiles of averaged temperature is related to the resultscorresponding to the combustion interval. LESoft and RWresults are similar, although there is a slight delay in the peakof temperature occurring in RW calculations compared toLESoft calculations. MESC calculations have a visible delayin the occurrence of peak of averaged temperature and themaximum level of temperature which are about 3 degrees and5%, respectively.

The main factor affecting the pressure and tempera-ture distributions in the engine cylinder is calculation ofheat generated during combustion (combustion sub-model).Although the heat release equation is matched in variouspackages considered, other factors may affect the resultscomputed. In MESC and reference calculations from [23],the heat release escapes through the piston and cylinderwall gap using a blow-by fraction of 0.8, whereas RW andLESoft use default settings. Other factor affecting pressureand temperature profiles is heat transfer sub-model, which

1000

2000

3000

0 90 1800−180 −90

𝜃 (deg)

T(K

)

RWLESoftMESCReference results

Figure 10: Comparison of averaged temperature profiles withreference data of [23].

Table 5: Input parameters for varying ignition timing test.

Parameter ValueIgnition timing,deg −60, −50, −10, 0

Compression ratio 10Air to fuel ratio 14.7Valve timing, deg EVO 170, EVC 330, IVO 390, IVC 550

is not flexible to modify in RW. Inlet and exhaust pipedimensions also affect the flows of air and combustionproducts in the cylinder, but these factors are not consideredin the calculations presented.

The indicated mean effective pressure from the calcula-tions with MESC package is 0.9528MPa, whereas a value of0.9510MPa was obtained in [23] (a difference is about 0.2%).Discrepancies in the conservation of mass and energy areabout 0.05% for all calculations.

4.4. Varying Ignition Timing. Ignition timing is one of themost important parameters in engine control. In this test, dif-ferent ignition timings are examined in terms of distributionsof pressure and temperature in the combustion chamber.Thegeometric parameters and operating parameters of engine areshown inTable 2, and someother input parameters, which arespecific for this test, are presented in Table 5.

The pressure distributions, predictedwith three packages,are presented in Figure 11. The results obtained are verysimilar in early ignition timing. For the late ignition timing,the discrepancy of the results computed is more significant in


RWLESoftMESC

p (M

Pa)

2

4

6

8

00

90 180−180 −90

𝜃 (deg)

(a)

RWLESoftMESC

p (M

Pa)

2

4

6

8

00 90 180−180 −90

𝜃 (deg)

(b)

RWLESoftMESC

p (M

Pa)

0.9

1.8

2.7

3.6

0

4.5

0 90 180−180 −90

𝜃 (deg)

(c)

p (M

Pa)

0.9

1.8

2.7

3.6

0

4.5

0 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

(d)

Figure 11: Profiles of pressure for different ignition timing of −60 (a), −50 (b), −10 (c), and 0 (d) degrees.

terms of occurrence of pressure peak. LESoft gives the earliestself-ignition which takes place just before crank angle of −10degrees. The pressure peaks, predicted with RW and MESCcalculations, are in a good agreement.

There is a small difference in the maximum value ofpressure, predicted with various packages. The location ofpressure peak, predicted with three packages, is shown in

Table 6. RW calculations give the earliest location of pressurepeak, and LESoft calculations give the latest location ofpressure peak in terms of crank angle.

The profiles of averaged temperature are shown inFigure 12. The profiles of temperature are very similar forthree packages. There are, however, small differences in themaximum level of temperature and small differences in


800

1600

2400

3200

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(a)

800

1600

2400

3200

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(b)

800

1600

2400

3200

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(c)

800

1600

2400

3200

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(d)

Figure 12: Profiles of temperature for different ignition timing of −60 (a), −50 (b), −10 (c), and 0 (d) degrees.

the corresponding crank angle. Those differences are causedby different assumptions on energy transfer in intake mani-fold and intake and exhaust valves.

Themain discrepancy of the distributions of temperature,predicted with various packages, is related to the location ofpeak of temperature. These results are shown in Table 7. Thetemperature profiles before crank angle corresponding to the

ignition are very similar. The location of peak of tempera-ture, predicted with LESoft and MESC, takes place earlierthan peak of temperature, predicted with RW. The sharptemperature decreasing taking place after the maximumof temperature during the expansion stroke is also similarbetween LESoft and MESC calculations. RW gives moreslowly temperature decreasing during the expansion stroke.


2

4

6

8

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

p (M

Pa)

(a)

2

4

6

8

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

p (M

Pa)

(b)

2

4

6

8

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

p (M

Pa)

(c)

2

4

6

8

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

p (M

Pa)

(d)

Figure 13: Profiles of pressure for air to fuel ratio of 0.8 (a), 0.9 (b), 1.0 (c), and 1.1 (d).

This behaviour may be explained by the combustion sub-model implemented in various packages. In the combustionmodel, implemented in RW, the timing is determined by themoment of 50% burned, rather than the start of ignition.Thecurve of mass fraction burned in LESoft andMESC packagesmay not be the same as in RW simulation. There is a morerapid increase in temperature when ignition started, until

the maximum of temperature reached. This result leads to ahigher temperature when combustion finished even thoughthe duration of burn is the same.

The results presented show that if the start of heat releasebegins too late, it occurs in an expending volume, resulting inlower combustion pressure and lower network. If the start ofheat release begins too early during the compression stroke,

Modelling and Simulation in Engineering 17T

(K)

1000

2000

3000

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

(a)

1000

2000

3000

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(b)

1000

2000

3000

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(c)

1000

2000

3000

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(d)

Figure 14: Profiles of temperature for air to fuel ratios of 0.8 (a), 0.9 (b), 1.0 (c), and 1.1 (d).

the negative compression work increases, since the piston isdoing work against the expanding combustion gases. Delayof the initiation of combustion leads to the pressure peakoccurring later during the expansion stroke.The pressure riseat 𝜃𝑠=−10∘ is almost double that at 𝜃

𝑠= 10∘. As the start of heat

release is changed from 𝜃𝑠= −10∘ to 𝜃

𝑠= 10∘, the work input

and the mean effective pressure decrease.

4.5. Varying Air to Fuel Ratio. The profiles of pressure, com-puted with various packages, are presented in Figure 13 fordifferent air to fuel ratios. All packages predict the maximumpressure level when the air to fuel ratio is stoichiometric.For leaner mixture, RW results are closer to those predictedwith LESoft, whereas RW results for richer mixture are closerto the data computed with MESC. The comparison of the


2

4

6

8

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

p (M

Pa)

(a)

0

2

4

6

8

0 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

p (M

Pa)

(b)

0

2

4

6

8

10

0 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

p (M

Pa)

(c)

0

2

4

6

8

10

0 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

p (M

Pa)

(d)

Figure 15: Profiles of pressure for compression ratios of 6 (a), 8 (b), 12 (c), and 14 (d).

pressure profiles show that the shape of profiles is very similarfor various packages, but the maximum pressure level isslightly different. The location of pressure peak takes place atthe same crank angle for various packages.

The impact of air to fuel ratio on temperature profilesis similar to pressure as this is presented in Figure 14. Themaximum temperature occurs when the air to fuel ratio isstoichiometric. RW results are closer to LESoft results forlearner mixture and closer to those predicted with MESC forricher mixture. In leaner mixture, RW predicts the location

of temperature peak closer to the ignition point comparedto MESC results. This may be because RW considers theposition of the injection in a pipe, which affects the flowof mixture, and hence takes into account this effect inpressure and temperature distributions computed in theengine cylinder.

4.6. Varying Compression Ratio. Compression ratio isdefined as the ratio of the maximum cylinder volume tothe minimum volume. The compression ratio determines


1000

2000

3000

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(a)

1000

2000

3000

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(b)

1000

2000

3000

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(c)

1000

2000

3000

00 90 180−180 −90

𝜃 (deg)

RWLESoftMESC

T(K

)

(d)

Figure 16: Profiles of temperature for compression ratios of 6 (a), 8 (b), 12 (c), and 14 (d).

the peak pressure and the peak temperature in the cycle.The efficiency of the engine and formation of pollutants arestrong functions of the compression ratio. The compressionratio is limited by the material strength and engine knock.Engine heads and blocks have a design of a maximum stress,which should not be exceeded, limiting the compressionratio. Compression ratio is difficult to control due to fixed

geometry of cylinder. In practice, compression ratio can bevaried by changing the piston head surface shape.

The results, computed with three packages, are presentedin Figure 15. The pressure increases as compression ratioincreases, and the data obtained are in agreement with theory[22, 23]. The position of maximum level of pressure slightlydepends on the compression ratio.


Table 6: Location of pressure peak predicted with various packages.

Start of ignition,deg RW MESC LESoft

−60 −2 −2 −4−50 2 0 −6−40 2 6 2−30 8 10 8−20 14 14 16−10 22 20 240 32 30 34

Table 7: Location of temperature peak predicted with variouspackages.

Start of ignition,deg

RWburned

RWaveraged

MESCburned LESoft

60 −8 −8 −16 −16−50 −2 0 −12 −6−40 −6 2 −8 2−30 0 12 −4 10−20 12 24 2 20−10 22 34 10 300 28 46 20 40

In LESoft and MESC calculations, the temperatureslightly increases as the compression ratio increases as thisis presented in Figure 16. In RW calculations, the highesttemperature is observed at the compression ratio of 10.

5. Conclusions and Recommendations

Two commercial engine simulation packages, Ricardo Waveand Lotus Engine Simulation, and in-house package, imple-mented with Matlab, have been considered, and their capa-bilities to the simulation of compression, expansion, andcombustion processes in one-cylinder engine have beencompared and discussed. The engine geometry is fixed in thetests performed. The main submodels affecting the resultscomputed are Woschni heat transfer model and Wiebe heatrelease model.

One-cylinder engine model has been designed with thementioned packages, and similar engine operating parame-ters have been specified (where it was possible). The results,computed with three packages, have been compared withreference data. The results obtained for reference test showthat RW, LESoft, and MESC predict similar profiles ofpressure and temperature which are in reasonable agreementwith reference results. In the combustion period, there is aslight shift of location of pressure peak with respect to thelocation of pressure peak predicted with MESC compared totwo commercial packages. Both commercial packages takeinto account position of fuel injectors, inlet pipes, and valvemodels which affect air flow and heat transfer in the cylinder.

Engine operating conditions, such as various ignitiontimings, various air to fuel ratios, and various compression

ratios have been considered in the tests, and the resultsobtained have been discussed. The results, computed withvarious packages, are in a good agreement. Small discrepan-cies may be explained by the various implementations of dif-ferent submodels (heat transfer sub-model and combustionsub-model) and different physical and chemical properties offuel and combustion products.

By considering the engine control in real driving con-dition, the in-house engine simulation package is the onlyone that allows running simulations with changing speed andother operating parameters throughout a simulation periodthat is required. The code can be used for real-time controlpurpose using exhaust gas composition of 10 spices.

Future work will focus on comparison of the resultscomputed with real engine bed tests, implementation ofthe control algorithms of valve timing, ignition timing, andinjection timing to control the performance of engine. ECUis capable to update the needed parameters while engines runwith alternated fuel compositions.

List of Notations

C: CarbonCFD: Computation Fluid DynamicsCO: Carbon oxideECU: Engine control unitEGR: Exhaust gas recirculationH: HydrogenHC: HydrocarbonICE: Internal Combustion EngineLESoft: Lotus Engine SimulationLIVC: Late intake valve closingMESC: Matlab Engine Simulation CodeN: NitrogenNOx: Nitrogen oxideO: OxygenRW: Ricardo WaveSI: Spark ignitionTDC: Top Dead CentreVVT: Variable valve timing.

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Review Article Comparison of Engine Simulation Software ...downloads.hindawi.com/journals/mse/2013/401643.pdf · e in-house package, Matlab Engine Simulation Code (MESC), has been