AN ANALYTICAL MODEL OF SPARK IGNITION ENGINEFOR PERFORMANCE PREDICTIONMR.SITTHICHOK SITTHIRACHAA THESIS SUBMITTED IN PARTIAL FULFILLMENT OF THE REQUIREMENTSFOR THE MASTER OF SCIENCE IN AUTOMOTIVE ENGINEERINGSIRINDHORN INTERNATIONAL THAI-GERMAN GRADUATE SCHOOL OF ENGINEERING(TGGS)GRADUATE COLLEGEKING MONGKUT’S INSTITUTE OF TECHNOLOGY NORTH BANGKOKACADEMIC YEAR 2006COPYRIGHT OF KING MONGKUT’S INSTITUTE OF TECHNOLOGY NORTH BANGKOK
Name : Mr. Sitthichok SitthirachaThesis Title : An Analytical Model of Spark Ignition Engine forPerformance PredictionMajor Field : Automotive EngineeringKing Mongkut’s Institute of Technology North BangkokThesis Advisors : Assistant Professor Dr.Suthum PatumsawadAssistant Professor Dr.Saiprasit KoetniyomAcademic Year : 2006AbstractThe objective of this thesis is to develop a mathematical model of spark ignitionengine based on cylinder-by-cylinder engine model which combines both physicalformulae, e.g. engine geometries, and empirical formulae, e.g. burning duration. Theengine performance, torque and power, can be calculated by integrating the pressureinside cylinder within one engine cycle. The model is verified by data from 8 enginemodels. It can capture torque and power characteristics very well. The overall errorsare in between -6% to 4%.The model is used for simulating in order to predict the burning duration of thealternative fuels. Furthermore, the model indicates the effects on η v when using thesealternative fuels too.(Total 60 pages)Keywords : Spark ignition engine, Gasoline engine, Engine modeling, Cylinder-bycylinderengine model, Engine simulation______________________________________________________________Advisorii
่่่่่่่่ํ้่ชื่อ : นายสิทธิโชค สิทธิราชาชื่อวิทยานิพนธ์ : แบบจําลองวิเคราะห์สําหรับการทํานายสมรรถนะของเครื่องยนต์จุดระเบิดด้วยประกายไฟสาขาวิชา: วิศวกรรมยานยนต์สถาบันเทคโนโลยีพระจอมเกล้าพระนครเหนือที่ปรึกษาวิทยานิพนธ์ : ผู้ชวยศาสตราจารย์ ดร.สุธรรม ปทุมสวัสดิผู้ชวยศาสตราจารย์ ดร.สายประสิทธิ เกดนิยม ิปีการศึกษา : 2549บทคัดย่อวัตถุประสงค์ของวิทยานิพนธ์นีคื ้ อ เพื่อพัฒนาแบบจําลองทางคณิตศาสตร์ของเครื่องยนต์จุดระเบิดด้วยประกายไฟบนพืนฐานของ ้ แบบจําลองเครื่องยนต์แบบสูบตอสูบ ซึ ่งรวมเอาทัง ้ ทฤษฎีทางกายภาพ เป็นต้นวา เรขาคณิตของเครื่องยนต์ และทฤษฎีที่ได้จากการสังเกต เป็นต้นวาระยะเวลาการเผาไหม้ สมรรถนะของเครื่องยนต์อันได้แก แรงบิดและกาลัง สามารถคํานวณได้โดยการอินทีเกรตความดันในกระบอกสูบภายใต้หนึ ่งรอบการหมุนของเครื่องยนต์ แบบจําลองได้ทดสอบเปรียบเทียบคากบเครื่องยนต์ ่ ั8 รุน พบวา สามารถทํานายลักษณะกราฟสมรรถนะของเครื่องยนต์ได้เป็นอยางดี ่ โดยมีคาผิดพลาดโดยรวมอยูระหวาง่ ่ ่ -6% ถึง 4%แบบจําลองนี ้ได้จําลองสถานการณ์เพื่อที่จะทํานายระยะเวลาในการเผาไหม้ของเชือเพลิงทางเลือกตางๆ ่ ยิงกวานัน ่ ่ ้ แบบจําลองยังสามารถที่จะอธิบายผลกระทบของการใช้เชือเพลิง ้ทางเลือกที่มีตอประสิทธิภาพเชิงปริมาตรได้อีกด้วย(วิทยานิพนธ์มีจํานวนทังสิน ้ ้ 60 หน้า)คําสําคัญ : เครื่องยนต์จุดระเบิดด้วยประกายไฟ, เครื่องยนต์เบนซิน, แบบจําลองเครื่องยนต์,แบบจําลองเครื่องยนต์แบบสูบตอสูบ ่ , การจําลองเครื่องยนต์________________________________________________อาจารย์ที่ปรึกษาวิทยานิพนธ์iii
ACKNOWLEDGEMENTSI would like to express my sincere gratitude to Professor Dr.Gyeung-Ho Choiof Power Train Laboratory, Keimyung University, Republic of Korea and AssistantProfessor Dr.Noppavan Chananpanich of King Mongkut’s Institute of TechnologyNorth Bangkok who are initiators of this story. My colleague and I had a chance tohave internship in the Power Train Laboratory under their allowance. This laboratorybrought me to the engine technology field. Until Associate Professor Dr.SuwatKuntanapreeda and Dr.Boonchai Watjatrakul went to visit us in South Korea. Theyadvised me to undertake this topic for my thesis. This thesis has been developed sincethat point. After I came back to Thailand, this thesis is strengthened by many preciouscomments and suggestions from Assistant Professor Dr.Suthum Patumsawad andAssistant Professor Dr.Saiprasit Koetniyom. Thanks to all persons who arementioned here. And also the staffs of the Power Train Laboratory who took care ofus along 4 months of our visit in South Korea.Sitthichok Sitthirachaiv
TABLE OF CONTENTSPageAbstract (in English)iiAbstract (in Thai)iiiAcknowledgementivList of TablesviiiList of FiguresixList of Abbreviations and SymbolsxiiChapter 1 Introduction 11.1 Background 11.2 Objectives 11.3 Approach 11.4 Scope 11.5 Assumptions 21.6 Impacts or Benefits from Research 21.7 Thesis Outline 2Chapter 2 Literature Review 32.1 Spark Ignition Engine 32.2 Modeling 52.2.1 Mean Value vs. Cylinder-by-cylinder Models 52.2.2 Limit of Physical Properties 5Chapter 3 The Model 63.1 Model Overview 63.2 Crank Slider Model 83.3 Cylinder Pressure Model 83.4 Wiebe Function 93.4.1 Burning Duration 93.5 Heat Input 103.6 Air/Fuel Ratio 103.7 Heat Transfer 113.7.1 Heat Transfer Coefficient Correlations 11v
TABLE OF CONTENTS (CONTINUED)Page3.8 Volumetric Efficiency 123.8.1 Flow through Valves 153.8.2 Valve Lift 163.8.3 Discharge Coefficient 173.8.4 Frictional Losses 173.8.5 Charge Heating 193.9 Residual Gas 193.10 Friction 213.11 Torque & Power 213.12 Minimum Spark Advance for Best Torque 213.13 Simulation conditions 223.14 Alternative Fuels 223.14.1 Ethanol-blended Gasoline or Gasohol 233.14.2 Ethanol 243.14.3 Compressed Natural Gas 243.14.4 Liquefied Petroleum Gas 25Chapter 4 Model Validation and Sensitivity Analysis 264.1 Performance Validation 264.2 Sensitivity Analysis 314.2.1 Burning Duration 314.2.2 Discharge Coefficient 324.2.3 Frictional Losses 334.2.4 Charge Heating 344.2.5 Exhaust Gas Temperature 344.3 Combustion Duration of Alternative Fuels 354.3.1 Ethanol-blended Gasoline or Gasohol 354.3.2 Ethanol 374.3.3 Compressed Natural Gas 384.3.4 Liquefied Petroleum Gas 39vi
TABLE OF CONTENTS (CONTINUED)PageChapter 5 Conclusions and Suggestions 415.1 Conclusions 415.2 Suggestions 41References 43Appendix A Engine Specifications 45Appendix B Matlab/Simulink Block Diagrams 50Biography 60vii
LIST OF TABLESTablePage3-1 Air/fuel ratios of many substances 113-2 Parameters for simulation 223-3 Summaries of major properties of gasoline, ethanol, E10 and E20 233-4 Summaries of major properties of methane, ethane, and natural gas 243-5 Summaries of major properties of propane, butane, and LPG 25viii
LIST OF ABBREVIATIONS AND SYMBOLS (CONTINUED)D ivD vfhHVIVCIVOklL iv,maxL vInlet Valve DiameterValve DiameterFraction of Heat AddedConvection Heat Transfer CoefficientHeating Value of FuelInlet Valve Close Angle After BDCInlet Valve Open angle Before TDCSpecific Heat RatioConnecting Rod LengthMaximum Inlet Valve LiftValve Lift Functionm air,stoich Theoretical Amount of Air Requirementm& Mass Flow Ratem aNp fp mep mip Tp 0PP eQQ inQ lossRsT gT exhT wAir MassEngine SpeedFriction Mean Effective PressureBrake Mean Effective PressureIndicated Mean Effective PressurePressure at RestrictionUpstream Stagnation PressurePressure Inside CylinderEffective PowerHeat AdditionOverall Heat InputHeat TransferGas ConstantStrokeTemperature of Cylinder GasExhaust Gas TemperatureCylinder Wall Temperaturexiii
LIST OF ABBREVIATIONS AND SYMBOLS (CONTINUED)T 0VV dΔθεη vθθ 0ρ aStagnation TemperatureCylinder VolumeDisplacement VolumeDuration of Heat AdditionCompression RatioVolumetric EfficiencyCrank AngleAngle of Start of Heat AdditionAir Densityxiv
CHAPTER 1INTRODUCTION1.1 BackgroundFour-stroke spark ignition engine was developed by Otto in 1876. This engineprovided power output 3 HP. Since that point, the engine developing has been donecontinuously over 100 years. Even now, some engines can provide power output morethan 1,000 HP.However, developments of spark ignition engine along 100 years has beenconducting very slow due to lots of parameter, such as physical geometries (bore,stroke, crank radius, compression ratio), ignition advanced, valve timing, combustioncharacteristic etc., influencing the performances. Many studies on effects of eachparameter were done by experiments. But this approach spent lots of expenses andtime such as building test engine, setting up laboratory, etc.Another approach is simulation method that allows engine designer to changeand test many different parameters without building real parts or even real engines.The engine model can be used in various ways including designing engine controlsystem or ECU, designing transmission, etc. Although the computational model isunable to obtain the exact characteristics because there are no perfect models andthere are many complex phenomena taking place in the engine, it simply estimates thetrend of those characteristics and effects with very low cost and less timeconsumption that is very helpful for speeding up the engine development processbefore making real one.1.2 Objectives1.2.1 Develop a physically based, cylinder-by-cylinder spark ignition enginemodel for predicting torque and power characteristics in order to study effects ofparameters such as engine geometries, fuel properties, valve timing layout, etc.1.2.2 Predict the burning duration of alternative fuels such as LiquefiedPetroleum Gas (LPG), Compressed Natural Gas (CNG) and gasohol which are neededinformation for simulation.1.3 ApproachThe engine model is developed under MatLab/Simulink which describes thetorque and power characteristics of the engine. The parameters, engine geometriesand valve timing layout, from existed gasoline engines are used. The model isevaluated and compared with test data in order to determine the accuracy.1.4 Scope1.4.1 Use Matlab/Simulink for developing the engine model.1.4.2 Gasoline is used for developing the model.1.4.3 Outputs of the model are torque and power characteristics under full loador wide open throttle (WOT) condition. For the alternative fuels, output is burningdurations.
21.5 Assumptions1.5.1 The model will be derived from the spark ignition engine without turbo orsuper charger and Exhaust Gas Recirculation (EGR) system.1.5.2 The engine uses multi-port injection system for adding fuels.1.5.3 Fuel evaporates to be gas phase completely.1.5.4 Fuel and air mix together perfectly.1.5.5 Fuel-air mixture is assumed to be an ideal gas all the time.1.5.6 There is no any effect of combustion chamber design.1.5.7 Combustion chamber wall temperature is set at 400 Kelvin.1.5.8 There is no effect of throttle body or there is no pressure drop at throttlebody due to WOT condition.1.5.9 There are no effect of exhaust stroke and exhaust system.1.6 Impacts or Benefits from ResearchThe model which is developed under this thesis can be used for many purposes.It can be used as an engineering tool in the development of the spark ignition engines.The simulation can reduce the time consumption and costly tests needed in laboratory.And also the model can be used for teaching a tool in the internal combustion enginecourse.1.7 Thesis OutlineThe details of this thesis are shown as followingChapter 1 IntroductionChapter 2 This chapter explains of fundamental of spark ignition engineand engine processes, review previous researches and theory ofengine modelingChapter 3 Focusing on the performance prediction, this chapter describesthe model and theory of each component in details andexplanations of the alternative fuelsChapter 4 Model validation by comparing to performance data fromvehicle manufacturer, analyze sensitivity when changingparameters, and find out combustion duration of the alternativefuelsChapter 5 Conclusions and suggestions
CHAPTER 2LITERATURE REVIEW2.1 Spark Ignition EngineThe spark ignition engines are used in different applications, such as cars, boatsand small power generators. Depending on the field of application, the spark ignitionengine has certain structure and components which may differ from field to field. Abasic spark ignition engine used within the automotive industry has the followingstructure and components as shown in Fig.2-1.FIGURE 2-1 Schematic of gasoline engine The basic operation of a four stroke engine involves intake, compression,expansion (or power), and exhaust strokes as shown in Fig.2-2. The piston starts at thetop, the intake valve opens, and the piston moves down to let the engine take in acylinder-full of air and gasoline. This is the intake stroke. Only the tiniest drop ofgasoline needs to be mixed into the air for this to work. Then the piston moves backup to compress this fuel/air mixture. Compression makes the explosion morepowerful. When the piston reaches almost the top of its stroke, the spark plug emits aspark to ignite mixture. The power generates from combustion and driving the pistondown. Once the piston hits the bottom of its stroke, the exhaust valve opens and theexhaust leaves the cylinder to go out the exhaust pipe.
4FIGURE 2-2 Basic four stroke cycleWith models for each of these processes, a simulation of complete engine cyclecan be built up and be analyzed to provide information on engine performances.These ideal models that describe characteristic of each process are proposed. Howeverthe calculation needs information from each state as shown in Fig.2-3a whichsometime could not obtain from real condition as shown in Fig.2-3b.(a)(b)FIGURE 2-3 Pressure-volume diagram of Otto cycle: (a) ideal; (b) realOverall engine work can be determined by integrating the area under thepressure-volume diagram. So many previous works concerned mainly prediction thepressure inside the combustion chamber [2, 3, 4]. But the pressure and volume areinfluenced by engine geometries during variation of crank angle. So the pressure anddisplacement volume are needed to convert as functions of crank angle. Kuo  andKirkpatrick  proposed the method that can calculate the pressure and volume at anycrank angle. The combustion process can be described by simple correlation, Wiebefunction .The results from Kuo  and Zeng et al  indicated that heat transfer frominside the cylinder to engine cooling water had much influences on the pressure insidethe cylinder. So the heat transfer function is needed to take into account in the model.Many researches reported that the mass of mixture that flows into the cylinderduring intake stroke is a very importance parameter [1, 2, 3],  because it affectsamount of fuel which mixes with the air. This mass can be determined by combiningthe ideal gas law and volumetric efficiency. However it is very difficult to evaluatebecause they are affected by many factors, such as manifold geometries and valvetiming . So Eriksson et al  and Kuo  assumed that the pressure inside
5manifold and inside the cylinder are the same, and neglect effect of volumetricefficiency. But Kuo  used corrective equation from real experiment to compensatethe errors. While Zeng et al  took the effect of volumetric efficiency into account.However the data were obtained from the real experiment and stored in a 3-dimensiontable by relation between engine speed and intake manifold pressure.It can say that combining of those methods that are mentioned above can predictthe engine performances precisely only if some testing data are known, mainly thevolumetric efficiency. So this thesis tries to use the one-dimension flow model topredict the volumetric efficiency in order to reduce the testing data dependence.2.2 ModelingThere are numerous ways of describing reality through a model [7, 8]. Some aremore complex than others and the different approaches may differ in both structureand accuracy. Choosing the model depends on the particular situation and especiallythe field of application. The model classifications are summarized follow.2.2.1 Mean Value vs. Cylinder-by-cylinder ApproachesWhen considering the cylinder, two main approaches can be found. The mostcommon use is mean value engine method. The mean value method defines numberof cylinders as one which occupies whole displacement volume. The fluctuating flowthrough the inlet port is modeled by average value over a cycle. The dynamics ofspeed, engine torque, pressure build-up in the inlet and exhaust manifolds are theaspects of most interest in this approach.Another method to the mean method is the cylinder-by-cylinder engineapproach. Unlike the mean method, it describes each cylinder individually andgenerates for example a torque signal with each individual combustion pulse present.Normally the mean model is sufficient enough for use in processes such as controlsystem design. But in engine performance development aspect, the cylinder-bycylindermethod is better because this method derives from engine geometries, whichis very useful for improving and optimizing the engine in the future.2.2.2 Limit of Physical PropertiesThere are a number of approaches available when deciding on the basis of amodel. Physical equation theoretically describing the system is the most commonmethod since it creates a general model working for many operating areas. Itsdrawback is that reality might be difficult to describe correctly in theory.Another common approach bases on the model entirely on measurements. Themeasured data is stored as a table of two or more dimensions in a so called black boxdepending on input signals. This approach often provides an accurate result since it isbased directly on empirical formulation. However it is only defined for a limitedregion. A combination of both approaches is commonly used. The main basis of themodel rests on physical equations. And empirical equations are used to model certaincomplexities.
CHAPTER 3THE MODELIn this chapter the model is described in detail. The engine to be modeled is agasoline engine which has no EGR and turbo system. The chosen model bases on thepressure inside the cylinder prediction. There are two main approaches to consider;mean value and cylinder-by-cylinder models. Since the objective of this thesis intendsto develop the model which can describe effects of each parameter on the engineperformance, the cylinder-by-cylinder approach is used in order to achieve this goal.Another consideration for model selection is limited by physical properties.Since there are no perfect equations which can describe phenomena in the engine,both physical and empirical formulae are used in the model.3.1 Model OverviewStartRPM Stroke Bore Conn. Rod Comp.Ratio Spark AngleIVCIVOkConn. RodA/F RatioHeating ValueLiv Div Tw No. Cylinderθ = -360Find displacementvolumeFind compressionvolumeADNoθ = 180YesEndFind exposed combustionchamber surface area, Eq.3-2BFind exhaust gastemperature, Eq.3-21Wiebe Function,Eq.3-4Find total heatinput, Eq.3-7Find pressure insidecylinder by integratingFind residual massCHeat release,Eq.3-5Find pressureincrement, Eq.3-3Find Temperatureof gas in cylinderFIGURE 3-1 Model overviewFind cylindervolume, Eq.3-1Find meanpiston speedFind heat transfer coefficient,Eq.3-10Find total heattransfer, Eq.3-8
7AFind valve liftfunction, Eq.3-16Find frictional lossfactor, Eq.3-19Find charge heatingfactor, Eq.3-20Find pressure that belongto flow-in massFind dischargecoefficient, Eq.3-18Find total flow-in massby integratingFind flow-in mass,Eq.3-15Cp T =p 0p0Yesp T=p Tp0Find valve curtainarea, Eq.3-14⎡ 2 ⎤⎢⎣k+ 1⎥⎦k /( k −1)⎛ p⎜⎝ p⎞⎟⎡≥2T⎟ ⎢0 ⎠ ⎣k+ 1Critic⎤⎥⎦k /( k −1)NoBFind indicated meanpressure, Eq.3-23Find friction loss,Eq.3-22Find effectivepressure, Eq.3-24Find effective power,Eq.3-25P eθ = θ + 1DFIGURE 3-2 (CONTINUED)
8Fig.3-1 and Fig.3-2 shows the overview of the model. Engine geometries, suchas bore, stroke, compression ratio, etc., are calculated to obtain physical informationsuch as displacement volume, area and volume variation as function of crank angle,etc. That information will be used for cylinder pressure prediction with another line ofinformation about heat input. Heat energy needs data from amount of flow in massand burn characteristic which is described by Wiebe function. Predicted pressure willbe used to determine temperature inside cylinder and also heat transfer from cylinderto wall chamber. Rate of heat loss will be fed back to the pressure prediction function.Resulted pressure will be converted to indicated mean effective pressure subtracted bymean friction, then work and power will be known finally. The Matlab/Simulinkblock diagrams of the model are shown in Appendix B. The details of each moduleare described in following section.3.2 Crank Slider ModelThe volume of the piston cylinder can be determined as a function of crankangle from the compression ratio (ε), the stroke (s), bore (b) and connecting rodlength (a). The geometric parameters of the piston cylinder can be described by thecrank slider model which is represented in Fig.3-3.FIGURE 3-3 Piston cylinder and geometriesThe equations of volume and area that relate to crank angle are described asfollowing equation :1VdVdl⎛ l ⎞ 2 2V ( θ ) = + [ + 1−cosθ− ( ⎜ ⎟ − sin θ)]Eq.3-1ε-12 a⎝ a ⎠π 2 s l⎛ l ⎞ 2 2A( θ ) = b + πb [ + 1−cosθ+ ( ⎜ ⎟ − sin θ)] Eq.3-22 2 a⎝ a ⎠3.3 Cylinder Pressure ModelThis model is derived from the first law of thermodynamics. The pressure isderived as a function of crank angle also .221dP=dθk −1⎛∂Q⎜ − QV ⎝ dθloss⎞ P dV⎟ − k⎠ V dθEq.3-3
11TABLE 3-1 Air/fuel ratios of many substances3.7 Heat TransferGenerally there are three modes of heat transfer in the combustion engine.Conduction - Conduction in solid matter is caused by molecular movement. Thedriving physical characteristic is the thermal conductivity. The heat flow in thecombustion chamber walls occurs through heat conduction.Convection - Convection is a term for heat transfer occurring in a moving fluid.The exchange of heat between the coolant or combustion gas and the combustionchamber wall occurs through convection. The velocity and the degree of turbulence ofthe moving fluid determine the degree of heat transfer via the convection.Radiation - Heat transfer through occurs in the form of electromagnetic waves.The energy of the radiation is proportional to T 4 . Radiation is only relevant in thecombustion chamber and only short period of time during and after combustion whenhigh gas temperatures are present.In spark ignition engines, the primary heat transfer mechanism from thecylinder gases to the wall is convection, with only 5% from radiation. Using aNewtonian model, the heat loss to the wall is given by:Q loss = hA(T g -T w )Eq.3-8∂QWhen determining the heat release term ( ) the heat loss to the walls has todθbe taken into account. Substitute Eq.3-5 and Eq.3-8 in Eq.3-3, the pressure over crankangle now becomes:k −1⎡ df hA ⎤= QinTgTw− kθ V ⎢ − ( − )⎣ dθ6N⎥⎦dPdPVdVdθEq.3-93.7.1 Heat Transfer Coefficient Correlations (h)The heat transfer coefficient is needed in order to calculate heat loss from thecylinder in Eq.3-8. For design purposes, simplified analyzes are often performedusing empirical heat transfer correlations such as Annand, Woschni and Hohenberg.These give at most estimates of the surface-averaged heat transfer coefficient, whichare defined in terms of the bulk gas temperature and used with this to calculatesurface-averaged or total heat flux. Annand assumed a constant characteristic gasvelocity equal to the mean piston speed, while Woschni assumed that the average gasvelocity should be proportional to the mean piston speed. Hohenberg examinedWoschni’s formula and made changes to give better predictions.
12Kleeman et al  compared these empirical correlations with ComputationalFluid Dynamic (CFD) prediction by using Standard Wall Function (SWF) andModified Wall Function (MWF) methods. Both SWF and MWF are boundary layermodels which have been employed to calculate wall shear and heat transfer in CFD.SWF derives the near-wall velocity and temperature profiles from a Couette flowanalysis, assuming steady one dimensional flow and ignore the variation within thelayer of the fluid thermophysical properties (e.g. density, viscosity, thermalconductivity). The use of SWF in CFD calculation of engine heat transfer generallyleads to underprediction of the wall heat flux. So MWF is developed further for usingin engine simulation by taking the fluid thermophysical properties into account. Theresults are shown in Fig.3-6.FIGURE 3-6 Comparison of surface-averaged heat flux variations predicted by CFDcalculations and empirical correlations The results from  showed that the MWF method gave the most accurateresults. So using these simple correlations will introduce some errors. However MWFcan only be obtained from CFD because the CFD program can calculate theinstantaneous local heat fluxes which are not uniform throughout the combustionchamber, while this thesis assumes the charge are burned completely throughout thecylinder at the same time. According to the results in Fig.3-6, the Hohenbergcorrelation is selected because it is closest to MWF and also there is no empiricalform for MWF. The Hohenberg correlation is following equation .−0.060.8 −0.40.8h = 130 V P ( + 1.4)Eq.3-10T gC m3.8 Volumetric Efficiency (η v )The intake system – the air filter, carburetor, and throttle plate (in a sparkignition engine), intake manifold, intake port, intake valve – restricts the amount of airwhich an engine of given displacement can induct. The parameter used to measure theeffectiveness of an engine’s induction process is the volumetric efficiency (η v ).
13Volumetric efficiency is only used with four-stroke engines which have distinctinduction process. It is defined as the volume of air which is drawn into the intakesystem divided by the volume which is displaced by the piston, Eq.3-11. Typicalmaximum values of volumetric efficiency for naturally aspirated engines are in therange 80 to 90%. The volumetric efficiency for diesels is somewhat higher than thespark ignition engines as shown in Fig.3-7.maηv= Eq.3-11ρaVdFIGURE 3-7 Volumetric efficiency versus mean piston speed for a four-cylinderindirect-injection diesel and a six-cylinder spark-ignition engine According to Fig.3-7, there are no such a model can predict the trend of thevolumetric efficiency exactly because it is affected by the following fuel, enginedesign and engine operating variables:1. Fuel type, air/fuel ratio, fraction of fuel vaporized in the intake system, andfuel heat of vaporization2. Mixture temperature as influenced by heat transfer3. Ratio of exhaust to inlet manifold pressures4. Compression ratio5. Engine speed6. Intake and exhaust manifold and port design7. Intake and exhaust valve geometry, size, lift and timingsThe manifold and valve geometry design has great effects on the volumetricefficiency since the designs of each engine model have never been the same. So thisthesis tries to develop the model in order to predict η v caused by the different valvedesign.
14FIGURE 3-8 Effect on volumetric efficiency of different phenomena which affectthe air flow rate as a function of speed. Solid line is final volumetricefficiency versus speed curve The shape of volumetric efficiency can be explained by Fig.3-8. It is affected bymany different phenomena. Non-speed-dependent effects (such as fuel vaporpressure) drop the volumetric efficiency below 100% (curve A). Charge heating in themanifold and cylinder drops curve A to curve B. It has greater effect at low enginespeeds due to longer gas residence time. Frictional flow losses increase as the squareof engine speed, and drop curve B to curve C. At higher engine speeds, the flow intothe engine during at least part of intake process becomes choked. Once this occurs,further increases in speed do not increase the flow rate significantly so volumetricefficiency decreases sharply (curve C to D). The induction ram effect at higher enginespeeds which occurs from inertia of mixture raises curve D to curve E. Late inletvalve closing, which allows advantage to be taken of increased charging at higherspeeds, results in a decrease in the volumetric efficiency at low engine speeds due tobackflow (curve C and D to F). Finally, intake and/or exhaust tuning can increase thevolumetric efficiency (often by substantial amount) over part of the engine speedrange, curve F to G. The terms which appear in Fig.3-8 are described as following.Frictional losses – During the intake stroke, the pressure in the cylinder is lessthan atmospheric pressure due to friction in each part of the intake system. This totalpressure drop is the sum of the pressure loss in each component of the intake system:air filter, carburetor and throttle, manifold, inlet port, and inlet valve.Ram effect – The pressure in the inlet manifold varies during each cylinder’sintake process due to the piston velocity variation, valve open area variation, and theunsteady gas-flow effects that result from geometric variations. At high enginespeeds, the inertia of the gas in the intake system as the intake valve is closingincreases the pressure in the port and continue charging process when the piston slowdown around BDC and starts the compression stroke. This effect becomesprogressively greater as engine speed is increased.
15Back flow – Because the inlet valve closes after the start of the compressionstroke, a reverse flow of fresh charge from the cylinder back into the intake can occuras the cylinder pressure rises due to piston motion toward TDC. This reverse flow islargest at the lowest engine speeds. This phenomenon cannot be avoided due tochoosing the inlet valve closing time for taking advantage of the ram effect at highspeeds.Tuning – The time-varying inlet flow to the cylinder causes expansion waves tobe propagated back into the inlet manifold. These expansion waves can be reflected atthe open end of the manifold (at the plenum) causing positive pressure waves to bepropagated toward the cylinder. If the timing of these waves is appropriately arranged,the positive pressure wave will cause raising the pressure at the inlet valve above thenominal at the end of intake process. This will increase the inducted air mass and bedescribed as tuned. This phenomenon can occur in exhaust system also.Charge Heating - According to the conduction heat transfer, the heat inside thecylinder transfers to inlet manifold via connecting ports and intake valves. This heat isabsorbed by air/fuel mixture directly and makes the mixture expand its volume due toincreasing of temperature. Finally, the volumetric efficiency decreases automatically.Generally, the residence times, length of inlet manifold and manifold geometries arethe main factors which influence how much heat can be transferred to the mixture.Choking - Choked flow of a fluid is caused by the Venturi effect. When aflowing fluid at a certain pressure and temperature flows through a restriction (such asthe hole in an orifice plate or a valve in a pipe) into a lower pressure environment,under the conservation of mass the fluid velocity must increase for initially subsonicupstream conditions as it flows through the smaller cross-sectional area of therestriction. At the same time, the Venturi effect causes the pressure to decrease.Choked flow is a limiting condition which occurs when the mass flow rate will notincrease with a further decrease in the downstream pressure environment.3.8.1 Flow through ValvesThe valve is usually the most important flow restriction in the intake and theexhaust system of four-stroke cycle engines. In this thesis considers only the intakevalve in order to determine η v . The mass flow rate through a poppet valve is usuallydescribed by the equation for compressible flow through a flow restriction, Eq.3-12.This equation is derived from a one-dimensional isentropic flow analysis, and real gasflow effects are included by means of an experimentally determined dischargecoefficient (C D ).0.51/ k( k −1)/ kC0 ⎪⎧2 ⎡⎤⎪⎫DARp ⎛ p ⎞⎛ ⎞T k pT10.5 ⎨ ⎢ − ⎥ ⎬(0)⎜⎟01⎜⎟RT p k − ⎢ p0⎥⎪ ⎭m& =Eq.3-12⎝ ⎠ ⎪⎩ ⎣ ⎝ ⎠ ⎦When the flow is choked, the pressure ratio (following value so called critical pressure ratio.p Tp 0) will not lower than the⎛ p⎜⎝ p⎞⎟⎡=2T⎟ ⎢0 ⎠ ⎣k+ 1Critic⎤⎥⎦k /( k −1)Eq.3-13
16For the mass flow of the mixture into the cylinder through the intake valve, p 0 isambient pressure, and p T is the cylinder pressure , . T 0 is ambient temperature.For A R , the most convenient reference area in practice is the so called valve curtainarea since it varies linearly with valve lift and is simple to determine [5, 6].A R = π D vLvEq.3-14Eq.3-12 should be converted into a function of crank angle also by dividingwith 6N same as Eq.3-9.dmd=θ0.51/ k( k −1)/ kC0 ⎪⎧2 ⎡⎤⎪⎫DARp ⎛ p ⎞⎛ ⎞T k pT10.5 ⎨ ⎢ − ⎥ ⎬6 (0)⎜⎟01⎜⎟N RT p k − ⎢ p0⎥⎪ ⎭⎝⎠⎪⎩⎣⎝⎠⎦Eq.3-15Eq.3-9 doesn’t consider about mass flow into cylinder. It determines pressuredifferent based on pressure at previous crank angle. But the pressure inside cylinder isinfluenced not only by volume variation, incoming mass also. So Eq.3-15 isintegrated to obtain amount of mass over a crank angle degree and then finding thepressure of mixture inside cylinder by using ideal gas equation of state, PV = mRT .The pressure is added with integrated pressure from Eq.3-9 to obtain p T in Eq.3-18.104.22.168 Valve LiftThe fundamental the valve lift design is to satisfy an engine breathingrequirement at the design speeds. However, it is philosophies and secrecies of eachcarmaker. One of the valve lift design uses polynomial function, for example,Hermann, McCartan and Blair (HMB) technique  uses up to 11th orderpolynomial functions. The alternative approach is the G. P. Blair (GPB) method which considers jerk characteristic of valve motion.Which method is employed depends on how smooth of the lift and/oracceleration diagrams are otherwise the forces and impacts on the cam followermechanism will be considered. In other words, a good mathematical smoothingtechnique within the valve lift design process is absolutely essential, may be degreeby degree level. So this thesis assumes to use cosine function for valve lift instead inorder to reduce complexity as following equation:L, max(1 + cosϕ)L ( θ ) = ivvEq.3-162π ( IVO − IVC + 2θ+ 540)ϕ =Eq.3-17IVO + IVC + 180
173.8.3 Discharge Coefficient (C D )Fig.3-9 shows the results of steady state flow tests on a typical inlet valveconfiguration with a sharp-cornered valve seat. The discharge coefficient based onvalve curtain area is a discontinuous function of the valve-lift/diameter ratio. Thethree segments shown correspond to different flow regimes as indicated. At very lowlifts, the flow remains attached to the valve head and seat, giving high values for thedischarge coefficient. At intermediate lifts, the flow separates from the valve head atthe inner edge of the valve seat as shown. An abrupt decrease in discharge coefficientoccurs at this point. The discharge coefficient then increases with increasing lift sincethe size of separated region remains approximately constant while the minimum flowarea is increasing. At high lifts, the flow separates from the inner edge of the valveseat as well. Typical maximum values of valve-lift/diameter ratio are 0.25.Although the flow through valve is dynamic behavior, it has been shown thatover the normal engine speed range, steady flow discharge coefficient results can beused to predict dynamic performance with reasonable precision. The averagedequation which obtained from the results in Fig.3-9 is shown as Eq.3-18.AverageFIGURE 3-9 Discharge coefficient of typical inlet poppet valve Lv4 Lv3 Lv2 LvCD= 190.47( ) −143.13() + 31.248( ) − 2.5999( ) + 0.6913 Eq.3-18DDDDiviv3.8.4 Frictional LossesTakizawa et al  made an experiment to measure pressure losses due tofriction across the air cleaner, carburetor, throttle and manifold plenum of a standardfour-cylinder automotive engine intake system. The results are shown in Fig.3-10.However, the parameters which are used to calculate the pressure losses stillunknown, mainly frictional factor for each component. So a frictional loss factor (C f )is introduced to adjust p 0 in Eq.3-15. The products of C f and p 0 are assumed as shownin Fig.3-11. The basis of C f is assumed to be a function of engine speed linearly inorder to reduce complexity as following equation.iviv
18NC f= −0 .019738( ) + 0. 986923 for 1000 ≤ N ≤ 6000 Eq.3-191000FIGURE 3-10 Pressure losses in the intake system of a four-stroke cycle sparkignitionengine determined under steady flow conditions 100Remaining Pressure, kPa989694929088860 1000 2000 3000 4000 5000 6000 7000Engine Speed, RPMFIGURE 3-11 Assumption of remaining pressure after subtracted by pressure losses
193.8.5 Charge HeatingAs mentioned in Sec.3.8, the residence times, length of inlet manifold andmanifold geometries are the main factors which influence how much heat can betransferred to the mixture. However, the length and geometries of manifold cannot bedetermined easily and the mixture velocity is dynamic behavior. So a charge heatingfactor (C heat ) is introduced to apply with T 0 in Eq.3-15. The assumption is defined thattemperatures inside manifold are in between cylinder wall temperature and ambienttemperature. The products of C heat and T 0 are assumed as shown in Fig.3-12, 373 K at1000 rpm and 308 K at 6000 rpm. The basis of C heat is assumed to be a function ofengine speed linearly in order to reduce complexity as following.NC heat= −0 .043624( ) + 1.2953 for 1000 ≤ N ≤ 6000 Eq.3-201000Temperature Inside Manifold, K3803603403203002802600 1000 2000 3000 4000 5000 6000 7000Engine Speed, RPMFIGURE 3-12 Assumption of temperature inside manifold3.9 Residual GasThe residual gas affects volumetric efficiency and engine performance directly,and efficiency and emissions through its effect on working fluid thermodynamicproperties . The residual gas is primarily a function of inlet and exhaust pressure,speed, compression ratio, valve timing, and exhaust system dynamics. However, theequation which can describe the magnitude of residual gas is still unknown. So thisthesis assumes that the amount of residual gas can be determined by using the idealgas equation of state, PV = mRT at TDC which is a function of exhaust pressure,compressed volume, exhaust gas molecular weight, and exhaust gas temperature. Theexhaust pressure is between 1 and 1.5 atm . So it is selected at 1.5 atm. The exhaustgas molecular weight is 30.4 g/mol.Caton and Heywood  measured cylinder pressure, calculated cylinder gastemperature and exhaust mass flow rate, and measured gas temperature at the exhaustport exit for a single-cylinder spark ignition engine at 1000 rpm. The results of are shown in Fig.3-13. According to Fig.3-13, the exhaust gas temperature at 1000rpm can be assumed starting at 900 K. The exhaust temperatures along other enginespeeds are calculated at point 4 in Fig.2-3a under ideal process. The increments oftemperature at point 4 on each engine speed are added to 900 K as shown in Fig.3-14.And the exhaust gas temperature equation is obtained from those results as following.
20N 3 N 2 NT exh= 3.3955( ) − 51.9( ) + 279.49( ) + 676.21 for 1000 ≤ N ≤ 60001000 1000 1000Eq.3-21FIGURE 3-13 Measured cylinder pressure p c , calculated cylinder-gas temperatureT c , exhaust mass flow rate m&e, and measured gas temperature atexhaust port exit T p , for single-cylinder spark ignition engine at speed= 1000 rpm 1400Exhaust Gas Temperature, K1200100080060040020000 1 2 3 4 5 6 7Engine Speed, RPMFIGURE 3-14 Assumption of temperature inside manifold
3.10 FrictionFriction losses influence the indicated power and useful output, the brakepower. Barnes-Moss  tested several four-stroke cycle four cylinder SI enginesbetween 845 and 2000 cm 3 displacement at wide-open throttle. The total friction workper cycle (and thus the friction mean effective pressure) for a given engine geometrywill vary with the engine speed as following equation.N 2 Np f= 0.05( ) + 0.15( ) + 0.97 for 1000 ≤ N ≤ 6000 Eq.3-221000 10003.11 Torque & PowerTo determine the overall performance, indicated mean effective pressure is used. The indicated mean effective pressure represents the work per combustion cyclenormalized by the displacement volume also called specific work. This valueindicates amount of maximum available power that can generate from single cylinder.21∫pmi= PdVEq.3-23VdThe engine has to overcome the friction loss. So the available output becomes:pme= p − pEq.3-24mifEffective power can be determined by following equation.Pe= 0 . 5N⋅ p ⋅VEq.3-25medFinally, torque can be determined by following relation.πNTP e= Eq.3-26303.12 Minimum Spark Advance for Best Torque (MBT)The charge of the air/fuel is burned by a flame-front beginning at the spark plug.The flame starts a kernel with a rather slow rate of expansion, but once a smallpercentage of the charge is ignited, the combustion process accelerates at a faster rate.Due to the very slow initial reaction rates, ignition must occur before TDC. This is the"advance" in ignition and is measured in degrees of crankshaft rotation. The bestadvance depends on design and operating conditions. This value is called “MinimumSpark Advance” which will produce the maximum torque at a given operatingcondition of speed and load of a given engine combination. In most cases the sparkadvance curve can be advanced several degrees before torque begins to drop off. If"knock" occurs, advance can be determined. The advance is referred to as "knocklimits". The fuel octane, camshaft profile and/or the compression ratio will need to beaddressed before maximum output can be achieved.
22(a)(b)FIGURE 3-15 (a) Cylinder pressure versus crank angle for over-advanced sparktiming (50°), MBT timing (30°), and retarded timing (10°). (b) Effectof spark advance on brake torque at constant speed and air/fuel ratio,at wide open throttle 3.13 Simulation conditionsThe conditions for simulation are summarized in Table 3-2.TABLE 3-2 Parameters for simulationParametersValueAmbient Pressure1 atmAmbient Temperature298 KMean cylinder wall temperature 400 K [11, 17]Specific heat ratio (k) 1.3 [3, 6]Air molecular weight28.97 g/molAir density 1.2 kg/m 3Gasoline molecular weight114 g/molGasoline heating value44,000 kJ/kgGasoline air/fuel ratio 14.6:1In real engine test (SAE, DIN standard), the performance of engine is measuredwith all accessories and standard intake and exhaust systems. All parameters whichaffect to the volumetric efficiency due to using those systems are taken into accountas well. Although the temperature inside manifold is varied from 308 to 373 Kelvindue to the heat transfer from cylinder (Sec.3.8.4) which affects to density of air, theair density is still kept constantly at 1.2 because the volumetric efficiency is measuredcomparing to the ambient condition.3.14 Alternative FuelsBy researches of many scientists, quantity of petroleum in the world has a limitand it cannot be renewed. 70% of petroleum is consumed in transportation and it isincreasing every year. With this rate, the petroleum will deplete from the world oneday in the near future. So many countries are searching the other sources of energy toreplace the petroleum in long run. Furthermore, improvement of air quality is anotherobjective too.
23Many road-tests on the alternative fuels are done to prove these potentials.However, there are a few on analytical studies. So this thesis would like to study thealternative fuels in analytical aspect. This section is going to mention about thealternative fuels, such as LPG, CNG and alcohol-blended gasoline in details and howthe model is applied to study these fuels.The model needs lots of parameter which are distinguished as shown in Fig.3-2,engine geometries, valve geometries, fuel properties, and engine operating condition,in order to predict the performance. However, the alternative fuel properties in detailsare still unknown, mainly the combustion duration. So the studies are set to predict thecombustion range of these alternative fuels by comparing between the outputperformances and simulated values which the data are obtained from previousresearches. The model is used to explain the volumetric efficiency in order todetermine effects of using those alternative fuels too.The previous researches are selected under specific condition. They have tostudy the using both gasoline and one of alternative fuels on the same engine in orderto compare the information. The gasoline engine model which is developed under thisthesis uses parameters of the engine in order to determine the spark advanced angle.With additional information of specific spark advanced setting on each alternativefuels, the burning duration can be determined.3.14.1 Ethanol-blended Gasoline or GasoholEthanol (ethyl alcohol) and methanol (methyl alcohol) are two types of alcoholfuels. The use of pure alcohols in internal combustion engines is only possible if theengine is designed or modified for that purpose. However, in their anhydrous or pureforms, they can be mixed with gasoline in various ratios for use in unmodifiedautomobile engines. Typically, only ethanol is used widely in this manner,particularly since methanol is toxic.E10, also frequently called gasohol, is a fuel mixture of 10% ethanol and 90%gasoline by volume that can be used in the internal combustion engines of mostmodern automobiles. However, not enough scientific tests have been done todetermine if E10 is harmful to older cars' fuel systems. It has been introducednationwide in Denmark and Thailand, and will replace high octane pure gasoline inThailand by 2007.E20 contains 20% ethanol and 80% gasoline. This fuel is not yet widely used inthe world. Since February 2006, this is the standard ethanol-gasoline mixture sold inBrazil, where concerns with the alcohol supply resulted in a drop in the ethanolpercentage, previously at 25%. Flexible-fuel cars are set up to run with gasoline insuch concentration range and few will work properly with lower concentrations ofethanol. The properties of E10 and E20 are compared in Table 3-3.TABLE 3-3 Summaries of major properties of gasoline, ethanol, E10 and E20 Properties Gasoline Ethanol E10 E20Heating value (kJ/kg) 44,000 27,000 41,900 40,000Stoichiometric air/fuel ratio 14.6 9 14 13.5
24Al-Farayedhi et al  investigated the effect of using unleaded gasoline–ethanol blends on typical SI engine performance in order to replace leaded gasoline.He founded that using of ethanol-blended fuel increase spark timing by average -1degree for E10 and -6 degree for E20 when comparing to gasoline. And also theengine brake thermal efficiency is improved when compared to the leaded fuel.3.14.2 EthanolEthanol fuel is an alternative to gasoline. Anhydrous ethanol or ethanol withoutwater can be blended with gasoline in any concentration up to pure ethanol (E100) toreduce the consumption of petroleum fuels, as well as to reduce air pollution. InBrazil, ethanol-powered and flexible-fuel vehicles are manufactured to be capable ofoperation by burning hydrated ethanol. In addition, flexible-fuel vehicles can run onany mixture of hydrated ethanol and gasoline, as long as there is at least 20% ofethanol. A few flexible-fuel systems, like the Hi-Flex, used by Renault and Fiat, canalso run with pure gasoline.Renault Clio is one of many vehicle models which are sold in Brazil. It isequipped by Hi-Flex technology which allows using pure ethanol as fuel with fullefficiency due to the automatic adjustment on engine system. The power curve of Clioincreases only 1 horsepower throughout engine speed range when using ethanol. Andalso, information from  indicates that the proper timing for an ethanol engine isfive to eight degrees advanced from the optimum gasoline setting.3.14.3 Compressed Natural Gas (CNG)CNG is a substitute for gasoline or diesel fuel. It is considered to be anenvironmentally "clean" alternative to those fuels. It comprises mainly methane (CH 4 )and ethane (C 2 H 6 ). It is stored and distributed in hard containers, usually cylinders,and keeped under high pressure. In response to high fuel prices and environmentalconcerns, compressed natural gas is starting to be used in light-duty passengervehicles and pickup trucks, medium-duty delivery trucks, and in transit and schoolbuses. The properties of natural gas are shown in Table 3-4TABLE 3-4 Summaries of major properties of methane, ethane, and natural gas Properties Methane(CH 4 )Ethane(C 2 H 6 )Natural Gas(CH 3.76 ) Heating value (kJ/kg) 50,100 47,400 49,500Stoichiometric air/fuel ratio 17.16 16 16.9Mello et al  evaluated the maximum horsepower in many vehicles whichwere converted to use both natural gas and gasoline, so called “bi-fuel vehicle”. Theyfounded that there is substantial drop in horsepower 13-17% when using the naturalgas with electronically lambda control. However, emissions are reduced dramatically.The drop of power came from using of gas mixer which restricts air flow throughmanifold. And also the natural gas is gaseous fuel which occupies a larger volumethan liquid fuel. It causes reduction of η v . Not only that, spark advanced angle isneeded to increase about 21-25 degree in order to obtain the maximum output due tolow burning rate.
253.14.4 Liquefied Petroleum Gas (LPG)LPG comprises primarily propane (C 3 H 8 ) and mixed by butane (C 4 H 10 ). LPG ismainly used in household, industries and transportation respectively. LPG can beproduced from crude oil and natural gas. However, the composition between propaneand butane depends on the sources. The characteristics of LPG are summarized inTable 3-5.TABLE 3-5 Summaries of major properties of propane, butane, and LPG Properties Propane(C 3 H 8 )Butane(C 4 H 10 )LPG(97.6%C 3 H 8 ) Heating value (kJ/kg) 46,000 45,400 45,900Stoichiometric air/fuel ratio 15.6 15.4 15.6Caton et al  developed a dedicated LPG fueled engine from production car.They founded that torque drops about 10-12% when using LPG without any enhancedmodification. The reason of performance reduction came from using the gaseous fuelwhich decreases η v . For spark timing, LPG needs an advanced spark timing usuallyaround +10 degree due to slower burning rate .
CHAPTER 4MODEL VALIDATION AND SENSITIVITY ANALYSISThis chapter presents and describes results acquired through simulations madewith the model. The results are compared to data which appear in vehicle technicalmanual in order to prove the accuracy.4.1 Performance ValidationThis section shows the results of simulation comparing to test data from 8engines. All engine geometries are obtained from Mercedes-Benz model year 1969which are summarized in Appendix A. The model is simulated between 1,000 to6,000 rpm of engine speed range.12030010025080200Power, kW60150Torque, Nm40Power100Power(Sim)20TorqueTorque(Sim)50001000 2000 3000 4000 5000 6000Engine Speed, RPMFIGURE 4-1 Simulation results of Mercedes-Benz 250SE
2712035010030080250Power, kW60200150Torque, Nm40PowerPower(Sim)10020TorqueTorque(Sim)50001000 2000 3000 4000 5000 6000Engine Speed, RPMFIGURE 4-2 Simulation results of Mercedes-Benz 250SL120400100350300Power, kW806040Power250200150Torque, NmPower(Sim)10020TorqueTorque(Sim)50001000 2000 3000 4000 5000 6000Engine Speed, RPMFIGURE 4-3 Simulation results of Mercedes-Benz 250E/8
28140400120350100300Power, kW806040PowerPower(Sim)250200150100Torque, Nm20TorqueTorque(Sim)50001000 2000 3000 4000 5000 6000Engine Speed, RPMFIGURE 4-4 Simulation results of Mercedes-Benz 280SE/8150400120320Power, kW9060240160Torque, NmPower30Power(Sim)Torque80Torque(Sim)001000 2000 3000 4000 5000 6000Engine Speed, RPMFIGURE 4-5 Simulation results of Mercedes-Benz 280SL/8
29150400120320Power, kW9060240160Torque, Nm30PowerPower(Sim)TorqueTorque(Sim)80001000 2000 3000 4000 5000 6000Engine Speed, RPMFIGURE 4-6 Simulation results of Mercedes-Benz 300SEL/8150400120320Power, kW9060240160Torque, Nm30PowerPower(Sim)TorqueTorque(Sim)80001000 2000 3000 4000 5000 6000Engine Speed, RPMFIGURE 4-7 Simulation results of Mercedes-Benz 300SEL
30200800160600Power, kW12080400Torque, Nm40PowerPower(Sim)TorqueTorque(Sim)20001000 2000 3000 4000 5000 6000Engine Speed, RPMFIGURE 4-8 Simulation results of Mercedes-Benz 600According to the results from Fig.4-1 to Fig.4-8, torque and powercharacteristics at low and high speed are almost the same to reference data for allcases. There are some cases greater at high speed. But all graphs are lower at midrange. When considering relative errors which are summarized from simulationresults of 8 engines in Fig.4-9, overall errors are in between -6% to 4%. But at lowand high engine speed, the error values are in positive side. While the error values innegative side occur in mid engine speed.0108642%0-2-4-6-8-100 1000 2000 3000 4000 5000 6000 7000Engine Speed (RPM)FIGURE 4-9 Summarized relative errors from simulation results of 8 engines
31According to the results, the model can interpret the output engine performancevery well. The errors might come from inadequate information which has to beassumed such as pressure losses in the intake system, temperature inside manifold,exhaust gas temperature, etc. And also some effects which are not included in themodel such as tuning, ram effect, exhaust stroke, and exhaust valve. However, thenext section is going to determine the effects of each parameter in order to find whichparameter has to be studied seriously in the future.4.2 Sensitivity AnalysisThis section is going to determine the sensitivity of the model affected by eachparameter.4.2.1 Burning DurationThis section is going to determine the effects of changing the gasoline burningduration from Eq.3-6 by increasing and decreasing 10 degree of burning durationwhile keeping the optimum gasoline spark advanced setting.1201.510010.5Power, kW806040%0-0.5-1+10 deg-10 deg20Gasoline+10 deg-10 deg-1.5-201000 2000 3000 4000 5000 6000Engine Speed, RPM-2.51000 2000 3000 4000 5000 6000Engine Speed, RPM(a)(b)FIGURE 4-10 Comparison between gasoline and +/-10 degree modified combustionrange without spark advanced adjusting (a) output power; (b) relativepower gainResults from Fig.4-10a indicate that power curves are almost the same whenincreasing and decreasing 10 degree of combustion range. When considering therelative power gain in Fig.4-10b, 10 degree increasing reduces power gradually from1.4% to 2.0%, while 10 degree decreasing raises power gradually up to 0.9%maximum except at low speed. These results roughly indicate that shorter combustiontime brings better performance to engine. The wave-liked curves cause frominappropriate spark advanced timing. So adjusting the spark timing is the next thing tostudy.
3221.510.5%0-0.5-1+10 deg-10 deg+10 deg optimum-10 deg optimum-1.5-2-2.51000 2000 3000 4000 5000 6000Engine Speed, RPMFIGURE 4-11 Relative errors of before and after adjusting the spark timingAccording to Fig.4-11, adjusting the optimum spark timing increases power inboth cases. For 10 degree increasing, power shifts up average 0.5% due to more 5degree advanced adjusting. And 10 degree decreasing case, power shifts up average0.7% due to less 5 degree advanced adjusting. It can conclude that changing ofcombustion duration affects to the output performance. But it does not affect muchand also it has no any effect to η v . Only one thing affected is the spark advancedangle.4.2.2 Discharge coefficientThis section is going to determine the effects of changing the dischargecoefficient from Eq.3-18 by increasing and decreasing 10%.14010120861004Power, kW8060%20-24020Normal+10% Cd-10% Cd-4-6-8+10% Cd-10% Cd01000 2000 3000 4000 5000 6000Engine Speed, RPM-101000 2000 3000 4000 5000 6000Engine Speed, RPM(a)(b)FIGURE 4-12 Comparison between normal C D and +/-10% changing(a) output power; (b) relative power gain
33According to Fig.4-12a, output powers do not look different much at low andmid engine speed. But at high engine speed, less C D value causes lower power thannormal, while greater C D value causes greater power than normal. When consideringrelative power gain in Fig.4-12b, less C D value causes more power a bit at low andmid engine speed, while greater C D value causes less power. For high engine speed,the results are explicitly the same as in Fig.4-12a.When considering Eq.3-15, less C D value causes less p T due to less mass flowinto the cylinder during intake stroke. With this reason, total mass flow in is increaseda bit at the end of intake process. But at high speed, lesser magnitude of p T does notaffect to mass flow rate any more due to choking which limits the mass flow rate. Soless C D value limits mass flow into the cylinder itself. The greater C D value givesopposite results to the explanations above. However, changing of C D alters the outputperformance very much only at high engine speed about +/-10%.4.2.3 Frictional LossesThis section is going to determine the effect of non-applying frictional lossfactor (C f ) in Eq.3-15. p 0 is kept at 1 atm constantly. Changing of p 0 affects η v mainly.So η v graph is considered in order to explain the effect.160140120110100Power, kW1008060nv, %90804020With CfWithout Cf7060Without CfWith Cf01000 2000 3000 4000 5000 6000Engine Speed, RPM500 1000 2000 3000 4000 5000 6000 7000Engine speed, rpm(a)(b)FIGURE 4-13 Comparison between using and non-using C f (a) output power; (b) η vThe using of ambient pressure gives the maximum possible power andmaximum η v to the engine. It is clearly that non-applying of the frictional lossescauses more power according to Fig.4-13a. When considering Fig.4-13b, η v curve isincreased also. The difference of η v increases gradually throughout the engine speedrange due to the assumption which defines more pressure drop for more speed.However, η v increases up to 100.8% which might cause from accumulative sum oferror. The lower values of η v are still under interferences of charge heating, back floweffect at low speed, and choking effect at high speed. So C f applying in Eq.3-15 isreasonable.
344.2.4 Charge HeatingThis section is going to determine the effect of non-applying charge heatingfactor (C heat ) in Eq.3-15. T 0 is kept at 298 K constantly. Changing of T 0 affects η vmainly. So η v graph is considered in order to explain the effect.14011012010010090Power, kW8060nv, %80407020With CheatWithout Cheat60Without CheatWith Cheat01000 2000 3000 4000 5000 6000Engine Speed, RPM500 1000 2000 3000 4000 5000 6000 7000Engine speed, rpm(a)(b)FIGURE 4-14 Comparison between using and non-using C heat(a) output power; (b) η vThe using of ambient temperature gives the maximum possible power andmaximum η v to the engine. It is clearly that non-applying of the charge heating causesmore power according to Fig.4-14a. When considering Fig.4-14b, η v curve isincreased also. The difference of η v reduces gradually throughout the engine speedrange due to the assumption which defines less heat transfer for more speed.However, η v increases up to 101.6% which might cause from accumulative sum oferror. The lower values of η v are still under interferences of frictional losses, backflow effect at low speed, and choking effect at high speed. So C heat applying in Eq.3-15 is reasonable.4.2.5 Exhaust Gas TemperatureThe exhaust gas temperature, Eq.3-21, is used to determine the residual masswhich is affected to η v as mentioned in Sec.3.9. This section is going to determine theeffects of changing the exhaust gas temperature (T exh ) by increasing and decreasingtemperature 50 K.
351400.11200.05100Power, kW8060%0-0.054020Normal+50 K Texh-50 K Texh-0.1-0.15+50 K Texh-50 K Texh01000 2000 3000 4000 5000 6000Engine Speed, RPM-0.21000 2000 3000 4000 5000 6000Engine Speed, RPM(a)(b)FIGURE 4-15 Comparison between normal and +/-50 K exhaust gas temperature(a) output power; (b) relative power gainAccording to Fig.4-15a, output powers are almost the same in both cases. Whenconsidering Fig.4-15b, Increasing of T exh raises power up to 0.85%, while reduction ofT exh reduces power up to 0.14%. Increasing of T exh reduces the residual massaccording to ideal gas equation of state, PV = mRT , and increases η v . Whilereduction of T exh gives opposite results. These results agree with explanations in .4.3 Combustion Duration of Alternative fuelsThis section is going to interpret the results of simulation in order to determinethe combustion duration of the alternative fuels.4.3.1 Ethanol-blended Gasoline or Gasohol908580nv, %757065GasolineE10E20600 1000 2000 3000 4000 5000Engine speed, rpmFIGURE 4-16 Effect of ethanol-blended fuel on volumetric efficiency
36120Combustion Duration, deg10080604020GasolineE10E2000 1000 2000 3000 4000 5000FIGURE 4-17 Combustion duration of gasoline, E10, and E20Using of ethanol-blended gasoline introduces increasing of η v by average 2.7%for E10 and 4.3% for E20 as results in Fig.4-16. The results agree with .Increasing of η v can compensate reduction of heating value and cause toque andpower gain. However, conditions of air/fuel ratio in  are rich mixture in all caseswhich increase η v greater than stoich condition due to the excess fuel. So η v might belower than simulation results around 0.5-1% at stoich condition.According to information of spark timing, it’s clear that combustion durationincreases in parallel as percentage of ethanol increase throughout engine speed rangeas shown in Fig.4-17. The equations of combustion duration are shown below.For E10N 2 NΔθ = −1.6429() + 20.479( ) + 41.271 for 1000 ≤ N ≤ 6000 Eq.4-11000 1000For E20N 2 NΔθ = −1.8571() + 21.443( ) + 50.343 for 1000 ≤ N ≤ 6000 Eq.4-21000 10001103100Power, kW9080706050403020GasolineE10E20%2.521.510.5E10E20101000 2000 3000 4000 5000 6000Engine Speed, RPM01000 2000 3000 4000 5000 6000Engine Speed, RPM(a)(b)FIGURE 4-18 Comparison between gasoline, E10, and E20 when using withunmodified engine (a) output power; (b) relative power gain
37In Thailand, the gasohol is introduced through nation for a while. However, theeffects of using gasohol are still questioned. So the scenario is set to determine effectof using E10 and E20 with unmodified spark ignition engine. The results are shown inFig.4-18. According to Fig.4-18a, the output powers seem to be almost the same. Butwhen considering in details, using gasohol brings more power both E10 and E20 asshown in Fig.4-18b. The maximum power gains are about 2% for E10 and 2.5% forE20 at low engine speed and gradually reduce through high engine speed. Accordingto the results, E10 and E20 can replace gasoline immediately in aspect ofperformance. Nevertheless, the maximum efficiency cannot be achieved. However,real experiments are needed to confirm the simulation results because the condition ofscenario is set to run under stoich condition which the ECU might not handle muchdifferent air/fuel ratio of E20. This effect occurs also in transient driving. Manydrivers may feel that their vehicles have less acceleration when using E10. It causesfrom ECU of feedback fuel control system getting “hesitation”. So if the ECU is notdesigned to be smart enough, this problem might appear even steady state running.4.3.2 Ethanol1009080nv, %706050GasolineE100400 1000 2000 3000 4000 5000 6000 7000Engine speed, rpmFIGURE 4-19 Effect of ethanol as a fuel on volumetric efficiency120Combustion Duration, deg10080604020GasolineE10E20E10000 1000 2000 3000 4000 5000FIGURE 4-20 Combustion duration of gasoline, E10, E20, and E100
38Using of ethanol as a fuel increases η v by average 7.2% as results in Fig.4-19.Although increasing of η v can compensate reduction of heating value, the power andtorque do not take advantages much due to very low heating value and lowcombustion rate when comparing to gasoline. According to the results in Fig.4-20, thecombustion duration of ethanol does not much difference to E20. It might concludethat the combustion lengths of E20 to E100 are almost the same. Only stoichiometricair/fuel ratio is different which can be controlled by the Flex-fuel technology. Theequation of combustion duration is shown below.For E100N 2 NΔθ = −1.6843() + 21.178( ) + 52.787 for 1000 ≤ N ≤ 6000 Eq.4-31000 10004.3.3 Compressed Natural Gas (CNG)908580nv, %757065GasolineCNG602500 3500 4500 5500Engine speed, rpmFIGURE 4-21 Effect of CNG as a fuel on volumetric efficiency160Combustion Duration, deg14012010080604020GasolineCNG00 1000 2000 3000 4000 5000FIGURE 4-22 Combustion duration of gasoline and CNG
39According to Fig.4-21, it’s clear that using of gaseous fuel and gas mixerintroduce η v reduction as mentioned in . It decreases about average 6.9%. Fig.4-22 shows the combustion duration of CNG comparing to gasoline. The combustionrange of CNG is almost twice time of the gasoline. So when considering to use CNGin both bi-fuel and dedicated system, spark timing adjusting is recommended. For bifuelsystem, electronic control box is needed to control the spark timing during usingof CNG. The equation of CNG combustion duration is shown below.For CNGN 2 NΔθ = −2.4286() + 27.976( ) + 72.405 for 1000 ≤ N ≤ 6000 Eq.4-41000 10004.3.4 Liquefied Petroleum Gas (LPG)908580nv, %757065GasolineLPG600 1000 2000 3000 4000 5000Engine speed, rpmFIGURE 4-23 Effect of ethanol as a fuel on volumetric efficiency140Combustion Duration, deg12010080604020GasolineLPG00 1000 2000 3000 4000 5000FIGURE 4-24 Combustion duration of gasoline and LPG
40The results agree as mentioned in  when using gaseous fuel. η v decreasesabout 3% average as shown in Fig.4-23. Caton et al  used sequential LPG gasinjection system for delivering fuel which sacrifices some performance due to the η vreduction. Combustion duration of LPG is greater than gasoline in Fig.4-24 and stillless than CNG. Nevertheless, using of LPG needs electronic control box in order toadjust the spark timing as well. The equation of LPG combustion duration is shownbelow.For LPGN 2 NΔθ = −1.8095() + 21.762( ) + 58.571 for 1000 ≤ N ≤ 6000 Eq.4-51000 1000
CHAPTER 5CONCLUSIONS AND SUGGESTIONS5.1 ConclusionsAn analytical model of spark ignition engine has been constructed based oncylinder-by-cylinder engine model which combines both physical formulae, e.g.engine geometries, and empirical formulae, e.g. burning duration. The engineperformance, torque and power, can be calculated by integrating the pressure insidecylinder within one engine cycle.In engine modeling, the model needs design parameters from real engine. It isthe same as 3-D engine model which needs the perfect 3-D geometry of combustionchamber, valves, and ports in order to achieve the accuracy. The parameters which areusually provided by commercial brochures are not enough for engine modeling.The model is verified by data from 8 engine models. It can capture torque andpower characteristics very well. The overall errors are in between -6% to 4%. But theerror values at low engine speed occur in positive side, while most of error valuesoccur in negative side at mid and high engine speed. These errors might cause fromcalculation itself, many assumptions and unknown phenomena which are notconsidered yet such as the tuning, ram effect, exhaust stroke, and exhaust valve, etc.There are significant changes to the performance curves when changing theparameters which affect η v , for example, pressure and temperature. So η v affectsdominantly to the performance shapes because it represents how much the fuel isdrawn into the cylinder. But the prediction of η v is still very difficult. It is combinedresult from a lot of parameters and many phenomena. If those phenomena can berealized, the accuracy can be achieved.The simulations in order to predict the burning duration of the alternative fuelsexpress many interesting information. All alternative fuels in this thesis have greatercombustion durations when compare to gasoline. Furthermore, the model indicates theeffects on η v when using these alternative fuels. Using of ethanol and ethanol blendedfuel can increase η v due to high latent heat value of ethanol itself. While using LPGand CNG reduce η v due to gaseous fuels.5.2 SuggestionsEq.3-3, which is used for predicting the pressure inside cylinder, is derived fromclose cycle. Kirkpatrick  made assumptions that compression and power strokes areclose system due to all valves closing and neglecting effects of incoming mixturewhich is drawn into cylinder. Only Q in is considered as input for Eq.3-3. But Q in canbe determined only if mixture mass is known. Therefore, 1-D mass flow model, Eq.3-12, is applied by comparing between ambient pressure and pressure inside cylinder.Then pressure is determined to by amount of flow-in mass and fed back to pressurecomparing loop. However, the ambient pressure and also the ambient temperature areinfluenced by frictional losses and charge heating respectively. So both topics shouldbe studied more in order to make better assumptions. For residual mass, theassumption should include other parameters as described in .
For part load condition, the throttle body model is needed to integrate into theengine model. However, the real geometries of the throttle body, e.g. close angle,maximum angle, inside throttle body diameter, are needed also in order to calculatethe mass flow through the throttle itself.All equations which describe the combustion range of the alternative fuelsshould be verified by experiment in order to prove model prediction accuracy.However, the results from experiments may deviate from the predictions caused byproportion of fuel mixture and also testing conditions.In order to predict the output performance of engines which use the alternativefuels, the model is needed to be revised due to the different technique of fueldispensing system, e.g. gas mixer, gas injector.42
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17. Klein, M. and Eriksson, L. “Models, methods and performance when estimatingthe compression ratio based on the cylinder pressure.” Proceedings ofCCSSE. (2002).18. Orbital Engine Company. A literature review based assessment on the impacts ofa 10% and 20% ethanol gasoline fuel blend on non-automotive engines.Available from:http://www.deh.gov.au/atmosphere/fuelquality/publications/review-nonautomotive/pubs/review.pdf[2006, November].19. Al-Farayedhi, A. A., Al-Dawood, A. M. and Gandhidasan, P. “Experimentalinvestigation of SI engine performance using oxygenated fuel.” Journal ofEngineering for Gas Turbines and Power. Vol. 126 (2004) : 178-191.20. Converting gasoline engines to run on alcohol. Available from:http://running_on_alcohol.tripod.com/id26.html [2006, December].21. Mello, P., et al. “Evaluation of the maximum horsepower of vehicles convertedfor use with natural gas fuel.” Fuel. Vol. 85, Issues 14-15 (2006) : 2180-2186.22. Caton, J. A., McDermott, M. and Chona, R. “Development of a dedicated LPGfueledspark-ignition engine and vehicle for the 1996 propane vehiclechallenge.” SAE Transactions - Journal of Fuels and Lubricants. Section 4,Vol. 106 (1998) : 792–805.23. RPi Engineering Ltd. V8 Rover LPG information. Available from:http://www.v8dualfuel.com/pdf/LPG_Info_2003.pdf [2006, December]44
APPENDIX AEngine Specifications
46This appendix summarizes all engine specifications which are used forsimulation in this thesisA.1 Mercedes-Benz 250SE (Model Year 1969)No. of Cylinders : 6No. of Valves per Cylinder : 2Displacement (cc.) : 2,500Bore × Stroke (mm.) : 82 × 78.8Compression Ratio : 9.3 : 1Intake Valve Open before TDC (degree) : 11°Intake Valve Close after BDC (degree) : 53°Maximum Valve Lift (mm.) : 8Inlet Valve Diameter (mm.) : 41.2FIGURE A-1Performance curve of Mercedes-Benz 250SEA.2 Mercedes-Benz 250SL (Model Year 1969)No. of Cylinders : 6No. of Valves per Cylinder : 2Displacement (cc.) : 2,500Bore × Stroke (mm.) : 82 × 78.8Compression Ratio : 9.5 : 1Intake Valve Open before TDC (degree) : 11°Intake Valve Close after BDC (degree) : 53°Maximum Valve Lift (mm.) : 8Inlet Valve Diameter (mm.) : 41.2FIGURE A-2Performance curve of Mercedes-Benz 250SL
47A.3 Mercedes-Benz 250E/8 (Model Year 1969)No. of Cylinders : 6No. of Valves per Cylinder : 2Displacement (cc.) : 2,500Bore × Stroke (mm.) : 82 × 78.8Compression Ratio : 9.5 : 1Intake Valve Open before TDC (degree) : 16°Intake Valve Close after BDC (degree) : 46°Maximum Valve Lift (mm.) : 8.5Inlet Valve Diameter (mm.) : 41.2FIGURE A-3 Performance curve of Mercedes-Benz 250E/8A.4 Mercedes-Benz 280SE/8 (Model Year 1969)No. of Cylinders : 6No. of Valves per Cylinder : 2Displacement (cc.) : 2,800Bore × Stroke (mm.) : 86.5 × 78.8Compression Ratio : 9.5 : 1Intake Valve Open before TDC (degree) : 11°Intake Valve Close after BDC (degree) : 47°Maximum Valve Lift (mm.) : 8.5Inlet Valve Diameter (mm.) : 41.2FIGURE A-4Performance curve of Mercedes-Benz 280SE/8
48A.5 Mercedes-Benz 280SL/8 (Model Year 1969)No. of Cylinders : 6No. of Valves per Cylinder : 2Displacement (cc.) : 2,800Bore × Stroke (mm.) : 86.5 × 78.8Compression Ratio : 9.5 : 1Intake Valve Open before TDC (degree) : 12°Intake Valve Close after BDC (degree) : 56°Maximum Valve Lift (mm.) : 9Inlet Valve Diameter (mm.) : 41.2FIGURE A-5Performance curve of Mercedes-Benz 280SL/8A.6 Mercedes-Benz 300SEL/8 (Model Year 1969)No. of Cylinders : 6No. of Valves per Cylinder : 2Displacement (cc.) : 2,800Bore × Stroke (mm.) : 86.5 × 78.8Compression Ratio : 9.5 : 1Intake Valve Open before TDC (degree) : 12°Intake Valve Close after BDC (degree) : 56°Maximum Valve Lift (mm.) : 9Inlet Valve Diameter (mm.) : 41.2FIGURE A-6Performance curve of Mercedes-Benz 300SEL/8
49A.7 Mercedes-Benz 300SEL (Model Year 1969)No. of Cylinders : 6No. of Valves per Cylinder : 2Displacement (cc.) : 3,000Bore × Stroke (mm.) : 85 × 88Compression Ratio : 8.8 : 1Intake Valve Open before TDC (degree) : 18°Intake Valve Close after BDC (degree) : 58°Maximum Valve Lift (mm.) : 7Inlet Valve Diameter (mm.) : 49FIGURE A-7Performance curve of Mercedes-Benz 300SELA.8 Mercedes-Benz 600 (Model Year 1969)No. of Cylinders : 8No. of Valves per Cylinder : 2Displacement (cc.) : 6,300Bore × Stroke (mm.) : 103 × 95Compression Ratio : 9 : 1Intake Valve Open before TDC (degree) : 2.5°Intake Valve Close after BDC (degree) : 52.5°Maximum Valve Lift (mm.) : 7Inlet Valve Diameter (mm.) : 48.95FIGURE A-8 Performance curve of Mercedes-Benz 600
APPENDIX BMatlab/Simulink Block Diagrams
51FIGURE B-1 Main model
52FIGURE B-2 Cm block detailsFIGURE B-3 Engine geometry block detailsFIGURE B-4 Engine geometry/Vd block detailsFIGURE B-5 Engine geometry/A(CA) block details
53FIGURE B-6 Engine geometry/Crank geometry block detailsFIGURE B-7 Engine geometry/V(CA),Vc block detailsFIGURE B-8 Wiebe fn block detailsFIGURE B-9 Burn duration block details
54FIGURE B-10 P block
55FIGURE B-11 P/Pratio block detailsFIGURE B-12 P/Lv fn block detailsFIGURE B-13 P/Cd block details
56FIGURE B-14 P/mdot block detailsFIGURE B-15 P/Cheat factor block detailsFIGURE B-16 P/Cf factor block details
57FIGURE B-17 Mw block detailsFIGURE B-18 Residual mass block detailsFIGURE B-19 T block detailsFIGURE B-20 Heat tran block details
58FIGURE B-21 h block detailsFIGURE B-22 Work&Power block detailsFIGURE B-23 Work&Power/Work block details
59FIGURE B-24 Work&Power/FMEP block detailsFIGURE B-25 Work&Power/Effective power block details
60BIOGRAPHYName : Mr.Sitthichok SitthirachaThesis Title : An Analytical Model of Spark Ignition Engine forPerformance PredictionMajor Field : Automotive EngineeringBiographyMr.Sitthichok Sitthiracha has studied in King Mongkut's Institute ofTechnology North Bangkok (KMITNB) since vocational school. It was first time forhim studying automotive vehicle. After that he studied Mechanical Engineering inKMITNB. As undergraduate, he joined the cooperative education program whichgave chance to him for working in automotive industry about one year. Along the oneyear program, he worked in many positions such as manufacturing & processengineer, product engineer, and quality engineer. Then he got a bachelor degree inMechanical Engineering in 2003. After one year of graduate, he decided to studyMaster degree in Automotive Engineering in KMITNB also. He was selected to havean internship in South Korea about LPG retrofit on heavy duty truck engines andengine testing for 4 months. He earned many experiences about engine technologythrough practicalities which become the inspiration for this thesis.