MR.SITTHICHOK SITTHIRACHA
A THESIS SUBMITTED IN PARTIAL FULFILLMENT OF THE REQUIREMENTS FOR THE MASTER OF SCIENCE IN AUTOMOTIVE ENGINEERING
SIRINDHORN INTERNATIONAL THAI-GERMAN GRADUATE SCHOOL OF ENGINEERING (TGGS)
GRADUATE COLLEGE
KING MONGKUT’S INSTITUTE OF TECHNOLOGY NORTH BANGKOK ACADEMIC YEAR 2006
Name : Mr. Sitthichok Sitthiracha
Thesis Title : An Analytical Model of Spark Ignition Engine for Performance Prediction
Major Field : Automotive Engineering
King Mongkut’s Institute of Technology North Bangkok Thesis Advisors : Assistant Professor Dr.Suthum Patumsawad
Assistant Professor Dr.Saiprasit Koetniyom Academic Year : 2006
Abstract
The objective of this thesis is to develop a mathematical model of spark ignition engine based on cylinder-by-cylinder engine model which combines both physical formulae, e.g. engine geometries, and empirical formulae, e.g. burning duration. The engine performance, torque and power, can be calculated by integrating the pressure inside cylinder within one engine cycle. The model is verified by data from 8 engine models. It can capture torque and power characteristics very well. The overall errors are in between -6% to 4%.
The model is used for simulating in order to predict the burning duration of the alternative fuels. Furthermore, the model indicates the effects on ηv when using these
alternative fuels too.
(Total 60 pages)
Keywords : Spark ignition engine, Gasoline engine, Engine modeling, Cylinder-by- cylinder engine model, Engine simulation
______________________________________________________________Advisor ii
ชื่อวิทยานิพนธ์ : แบบจําลองวิเคราะห์สําหรับการทํานายสมรรถนะของเครื่องยนต์
จุดระเบิดด้วยประกายไฟ
สาขาวิชา
: วิศวกรรมยานยนต์
สถาบันเทคโนโลยีพระจอมเกล้าพระนครเหนือ
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ปีการศึกษา : 2549
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แบบจําลองได้
ทดสอบเปรียบเทียบคากบเครื่องยนต์
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8 รุน พบวา สามารถทํานายลักษณะกราฟสมรรถนะของ
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เครื่องยนต์ได้เป็นอยางดี โดยมีคาผิดพลาดโดยรวมอยูระหวาง
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แบบจําลองนี้ได้จําลองสถานการณ์เพื่อที่จะทํานายระยะเวลาในการเผาไหม้ของเชือเพลิง
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60หน้า
)คําสําคัญ : เครื่องยนต์จุดระเบิดด้วยประกายไฟ, เครื่องยนต์เบนซิน, แบบจําลองเครื่องยนต์,
แบบจําลองเครื่องยนต์แบบสูบตอสูบ
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, การจําลองเครื่องยนต์
________________________________________________อาจารย์ที่ปรึกษาวิทยานิพนธ์
iiiACKNOWLEDGEMENTS
I would like to express my sincere gratitude to Professor Dr.Gyeung-Ho Choi of Power Train Laboratory, Keimyung University, Republic of Korea and Assistant Professor Dr.Noppavan Chananpanich of King Mongkut’s Institute of Technology North Bangkok who are initiators of this story. My colleague and I had a chance to have internship in the Power Train Laboratory under their allowance. This laboratory brought me to the engine technology field. Until Associate Professor Dr.Suwat Kuntanapreeda and Dr.Boonchai Watjatrakul went to visit us in South Korea. They advised me to undertake this topic for my thesis. This thesis has been developed since that point. After I came back to Thailand, this thesis is strengthened by many precious comments and suggestions from Assistant Professor Dr.Suthum Patumsawad and Assistant Professor Dr.Saiprasit Koetniyom. Thanks to all persons who are mentioned here. And also the staffs of the Power Train Laboratory who took care of us along 4 months of our visit in South Korea.
Sitthichok Sitthiracha
Abstract (in English) ii
Abstract (in Thai) iii
Acknowledgement iv
List of Tables viii
List of Figures ix
List of Abbreviations and Symbols xii
Chapter 1 Introduction 1 1.1 Background 1 1.2 Objectives 1 1.3 Approach 1 1.4 Scope 1 1.5 Assumptions 2
1.6 Impacts or Benefits from Research 2
1.7 Thesis Outline 2
Chapter 2 Literature Review 3
2.1 Spark Ignition Engine 3
2.2 Modeling 5
2.2.1 Mean Value vs. Cylinder-by-cylinder Models 5
2.2.2 Limit of Physical Properties 5
Chapter 3 The Model 6
3.1 Model Overview 6
3.2 Crank Slider Model 8
3.3 Cylinder Pressure Model 8
3.4 Wiebe Function 9
3.4.1 Burning Duration 9
3.5 Heat Input 10 3.6 Air/Fuel Ratio 10 3.7 Heat Transfer 11
3.7.1 Heat Transfer Coefficient Correlations 11
TABLE OF CONTENTS (CONTINUED)
Page
3.8 Volumetric Efficiency 12
3.8.1 Flow through Valves 15
3.8.2 Valve Lift 16 3.8.3 Discharge Coefficient 17 3.8.4 Frictional Losses 17 3.8.5 Charge Heating 19 3.9 Residual Gas 19 3.10 Friction 21
3.11 Torque & Power 21
3.12 Minimum Spark Advance for Best Torque 21
3.13 Simulation conditions 22
3.14 Alternative Fuels 22
3.14.1 Ethanol-blended Gasoline or Gasohol 23
3.14.2 Ethanol 24
3.14.3 Compressed Natural Gas 24
3.14.4 Liquefied Petroleum Gas 25
Chapter 4 Model Validation and Sensitivity Analysis 26
4.1 Performance Validation 26 4.2 Sensitivity Analysis 31 4.2.1 Burning Duration 31 4.2.2 Discharge Coefficient 32 4.2.3 Frictional Losses 33 4.2.4 Charge Heating 34
4.2.5 Exhaust Gas Temperature 34
4.3 Combustion Duration of Alternative Fuels 35 4.3.1 Ethanol-blended Gasoline or Gasohol 35
4.3.2 Ethanol 37
4.3.3 Compressed Natural Gas 38
4.3.4 Liquefied Petroleum Gas 39
Chapter 5 Conclusions and Suggestions 41
5.1 Conclusions 41
5.2 Suggestions 41
References 43
Appendix A Engine Specifications 45
Appendix B Matlab/Simulink Block Diagrams 50
Biography 60
LIST OF TABLES
Table Page
3-1 Air/fuel ratios of many substances 11
3-2 Parameters for simulation 22
3-3 Summaries of major properties of gasoline, ethanol, E10 and E20 23 3-4 Summaries of major properties of methane, ethane, and natural gas 24 3-5 Summaries of major properties of propane, butane, and LPG 25
Figure Page
2-1 Schematic of gasoline engine 6
2-2 Basic four stroke cycle 3
2-3 Pressure-volume diagram of Otto cycle 3
3-1 Model overview 6
3-2 Model overview (continued) 7
3-3 Piston cylinder and geometries 8 3-4 Burned mass fraction characteristic 9 3-5 Effect of engine speed on flame-development angle 10 3-6 Comparison of surface-averaged heat flux variations predicted by CFD
calculations and empirical correlations 12 3-7 Volumetric efficiency versus mean piston speed for a four-cylinder
indirect-injection diesel and a six-cylinder spark-ignition engine 13 3-8 Effect on volumetric efficiency of different phenomena which affect
the air flow rate as a function of speed. Solid line is final volumetric
efficiency versus speed curve 14
3-9 Discharge coefficient of typical inlet poppet valve 17 3-10 Pressure losses in the intake system of a four-stroke cycle spark-ignition
engine determined under steady flow conditions 18 3-11 Assumption of remaining pressure after subtracted by pressure losses 18 3-12 Assumption of temperature inside manifold 19 3-13 Measured cylinder pressure pc, calculated cylinder-gas temperature Tc,
exhaust mass flow rate m& , and measured gas temperature at exhaust port e
exit Tp, for single-cylinder spark ignition engine at speed = 1000 rpm 20
3-14 Assumption of exhaust gas temperature 20 3-15 Cylinder pressure versus crank angle for over-advanced spark timing (50°),
MBT timing (30°), and retarded timing (10°) and Effect of spark advance on brake torque at constant speed and air/fuel ratio, at wide open throttle 22 4-1 Simulation results of Mercedes-Benz 250SE 26
LIST OF FIGURES (CONTINUED)
Figure Page
4-2 Simulation results of Mercedes-Benz 250E/8 27 4-3 Simulation results of Mercedes-Benz 250SL 27 4-4 Simulation results of Mercedes-Benz 280SE/8 28 4-5 Simulation results of Mercedes-Benz 280SL/8 28 4-6 Simulation results of Mercedes-Benz 300SEL/8 29 4-7 Simulation results of Mercedes-Benz 300SEL 29
4-8 Simulation results of Mercedes-Benz 600 30
4-9 Summarized relative errors from simulation results of 8 engines 30 4-10 Comparison between gasoline and +/-10 modified combustion range
without spark advanced adjusting 31
4-11 Relative errors of before and after adjusting the spark timing 32 4-12 Comparison between normal CD and +/-10% changing 32
4-13 Comparison between using and non-using Cf 33
4-14 Comparison between using and non-using Cheat 34
4-15 Comparison between normal and +/-50 K exhaust gas temperature 35 4-16 Effect of ethanol-blended fuel on volumetric efficiency 35 4-17 Combustion duration of gasoline, E10, and E20 29 4-18 Comparison between gasoline, E10, and E20 when using with unmodified
engine 36
4-19 Effect of ethanol as a fuel on volumetric efficiency 37 4-20 Combustion duration of gasoline, E10, E20, and E100 37 4-21 Effect of ethanol as a fuel on volumetric efficiency 38 4-22 Combustion duration of gasoline and CNG 38 4-23 Effect of ethanol as a fuel on volumetric efficiency 39
4-24 Combustion duration of gasoline and LPG 39
A-1 Performance curve of Mercedes-Benz 250SE 46 A-2 Performance curve of Mercedes-Benz 250SL 46 A-3 Performance curve of Mercedes-Benz 250E/8 47 A-4 Performance curve of Mercedes-Benz 280SE/8 47
Figure Page A-5 Performance curve of Mercedes-Benz 280SL/8 48 A-6 Performance curve of Mercedes-Benz 300SEL/8 48 A-7 Performance curve of Mercedes-Benz 300SEL 49
A-8 Performance curve of Mercedes-Benz 600 49
B-1 Main model 51
B-2 Cm block details 52
B-3 Engine geometry block details 52
B-4 Engine geometry/Vd block details 52
B-5 Engine geometry/A(CA) block details 52
B-6 Engine geometry/Crank geometry block details 53 B-7 Engine geometry/V(CA),Vc block details 53
B-8 Wiebe fn block details 53
B-9 Burn duration block details 53
B-10 P block details 54
B-11 P/Pratio block details 55
B-12 P/Lv fn block details 55
B-13 P/Cd block details 55
B-14 P/mdot block details 56
B-15 P/Cheat factor block details 56
B-16 P/Cf factor block details 56
B-17 Mw block details 57
B-18 Residual mass block details 57
B-19 T block details 57
B-20 Heat tran block details 57
B-21 h block details 58
B-22 Work&Power block details 58
B-23 Work&Power/Work block details 58
B-24 Work&Power/FMEP block details 59
B-25 Work&Power/Effective power block details 59 xi
LIST OF ABBREVIATIONS AND SYMBOLS
Abbreviations
BDC Bottom Dead Center
CFD Computational Fluid Dynamic CNG Compressed Natural Gas CO Carbon Monoxide
DIN Deutschland Industrial Norm ECU Engine Control Unit
EGR Exhaust Gas Recirculation HC Hydrocarbon
HHV High Heating Value HP Horsepower
LPG Liquefied Petroleum Gas MBT Maximum Brake Torque MWF Modified Wall Function rpm Round per Minute
SAE Society of Automotive Engineers SWF Standard Wall Function
TDC Top Dead Center WOT Wide Open Throttle
Symbols
a Crank Radius
A Exposed Combustion Chamber Surface Area AR Reference Area
b Cylinder Bore
CD Discharge Coefficient
Cf Frictional Loss Factor
Cheat Charge Heating Factor Cm Mean Piston Speed
Div Inlet Valve Diameter
Dv Valve Diameter f Fraction of Heat Added
h Convection Heat Transfer Coefficient
HV Heating Value of Fuel
IVC Inlet Valve Close Angle After BDC IVO Inlet Valve Open angle Before TDC
k Specific Heat Ratio
l Connecting Rod Length
Liv,max Maximum Inlet Valve Lift
Lv Valve Lift Function
mair,stoich Theoretical Amount of Air Requirement
m& Mass Flow Rate ma Air Mass
N Engine Speed
pf Friction Mean Effective Pressure
pme Brake Mean Effective Pressure
pmi Indicated Mean Effective Pressure
pT Pressure at Restriction
p0 Upstream Stagnation Pressure
P Pressure Inside Cylinder
Pe Effective Power Q Heat Addition Qin Overall Heat Input Qloss Heat Transfer
R Gas Constant
s Stroke
Tg Temperature of Cylinder Gas
Texh Exhaust Gas Temperature Tw Cylinder Wall Temperature
LIST OF ABBREVIATIONS AND SYMBOLS (CONTINUED)
T0 Stagnation Temperature
V Cylinder Volume
Vd Displacement Volume Δθ Duration of Heat Addition
ε Compression Ratio ηv Volumetric Efficiency
θ Crank Angle
θ0 Angle of Start of Heat Addition
ρa Air Density
1.1 Background
Four-stroke spark ignition engine was developed by Otto in 1876. This engine provided power output 3 HP. Since that point, the engine developing has been done continuously over 100 years. Even now, some engines can provide power output more than 1,000 HP.
However, developments of spark ignition engine along 100 years has been conducting very slow due to lots of parameter, such as physical geometries (bore, stroke, crank radius, compression ratio), ignition advanced, valve timing, combustion characteristic etc., influencing the performances. Many studies on effects of each parameter were done by experiments. But this approach spent lots of expenses and time such as building test engine, setting up laboratory, etc.
Another approach is simulation method that allows engine designer to change and test many different parameters without building real parts or even real engines. The engine model can be used in various ways including designing engine control system or ECU, designing transmission, etc. Although the computational model is unable to obtain the exact characteristics because there are no perfect models and there are many complex phenomena taking place in the engine, it simply estimates the trend of those characteristics and effects with very low cost and less time consumption that is very helpful for speeding up the engine development process before making real one.
1.2 Objectives
1.2.1 Develop a physically based, cylinder-by-cylinder spark ignition engine model for predicting torque and power characteristics in order to study effects of parameters such as engine geometries, fuel properties, valve timing layout, etc.
1.2.2 Predict the burning duration of alternative fuels such as Liquefied Petroleum Gas (LPG), Compressed Natural Gas (CNG) and gasohol which are needed information for simulation.
1.3 Approach
The engine model is developed under MatLab/Simulink which describes the torque and power characteristics of the engine. The parameters, engine geometries and valve timing layout, from existed gasoline engines are used. The model is evaluated and compared with test data in order to determine the accuracy.
1.4 Scope
1.4.1 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 load or wide open throttle (WOT) condition. For the alternative fuels, output is burning durations.
1.5 Assumptions
1.5.1 The model will be derived from the spark ignition engine without turbo or super 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 throttle body due to WOT condition.
1.5.9 There are no effect of exhaust stroke and exhaust system.
1.6 Impacts or Benefits from Research
The 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 engine course.
1.7 Thesis Outline
The details of this thesis are shown as following Chapter 1 Introduction
Chapter 2 This chapter explains of fundamental of spark ignition engine and engine processes, review previous researches and theory of engine modeling
Chapter 3 Focusing on the performance prediction, this chapter describes the model and theory of each component in details and explanations of the alternative fuels
Chapter 4 Model validation by comparing to performance data from vehicle manufacturer, analyze sensitivity when changing parameters, and find out combustion duration of the alternative fuels
Chapter 5 Conclusions and suggestions
2.1 Spark Ignition Engine
The spark ignition engines are used in different applications, such as cars, boats and small power generators. Depending on the field of application, the spark ignition engine has certain structure and components which may differ from field to field. A basic spark ignition engine used within the automotive industry has the following structure and components as shown in Fig.2-1.
FIGURE 2-1 Schematic of gasoline engine [1]
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 the top, the intake valve opens, and the piston moves down to let the engine take in a cylinder-full of air and gasoline. This is the intake stroke. Only the tiniest drop of gasoline needs to be mixed into the air for this to work. Then the piston moves back up to compress this fuel/air mixture. Compression makes the explosion more powerful. When the piston reaches almost the top of its stroke, the spark plug emits a spark to ignite mixture. The power generates from combustion and driving the piston down. Once the piston hits the bottom of its stroke, the exhaust valve opens and the exhaust leaves the cylinder to go out the exhaust pipe.
FIGURE 2-2 Basic four stroke cycle
With models for each of these processes, a simulation of complete engine cycle can be built up and be analyzed to provide information on engine performances. These ideal models that describe characteristic of each process are proposed. However the calculation needs information from each state as shown in Fig.2-3a which sometime 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) real
Overall engine work can be determined by integrating the area under the pressure-volume diagram. So many previous works concerned mainly prediction the pressure inside the combustion chamber [2, 3, 4]. But the pressure and volume are influenced by engine geometries during variation of crank angle. So the pressure and displacement volume are needed to convert as functions of crank angle. Kuo [3] and Kirkpatrick [5] proposed the method that can calculate the pressure and volume at any crank angle. The combustion process can be described by simple correlation, Wiebe function [6].
The results from Kuo [3] and Zeng et al [4] indicated that heat transfer from inside the cylinder to engine cooling water had much influences on the pressure inside the 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 cylinder during intake stroke is a very importance parameter [1, 2, 3], [6] because it affects amount of fuel which mixes with the air. This mass can be determined by combining the ideal gas law and volumetric efficiency. However it is very difficult to evaluate because they are affected by many factors, such as manifold geometries and valve timing [6]. So Eriksson et al [2] and Kuo [3] assumed that the pressure inside 4
manifold and inside the cylinder are the same, and neglect effect of volumetric efficiency. But Kuo [3] used corrective equation from real experiment to compensate the errors. While Zeng et al [1] took the effect of volumetric efficiency into account. However the data were obtained from the real experiment and stored in a 3-dimension table by relation between engine speed and intake manifold pressure.
It can say that combining of those methods that are mentioned above can predict the engine performances precisely only if some testing data are known, mainly the volumetric efficiency. So this thesis tries to use the one-dimension flow model to predict the volumetric efficiency in order to reduce the testing data dependence.
2.2 Modeling
There are numerous ways of describing reality through a model [7, 8]. Some are more complex than others and the different approaches may differ in both structure and accuracy. Choosing the model depends on the particular situation and especially the field of application. The model classifications are summarized follow.
2.2.1 Mean Value vs. Cylinder-by-cylinder Approaches
When considering the cylinder, two main approaches can be found. The most common use is mean value engine method. The mean value method defines number of cylinders as one which occupies whole displacement volume. The fluctuating flow through the inlet port is modeled by average value over a cycle. The dynamics of speed, engine torque, pressure build-up in the inlet and exhaust manifolds are the aspects of most interest in this approach.
Another method to the mean method is the cylinder-by-cylinder engine approach. Unlike the mean method, it describes each cylinder individually and generates for example a torque signal with each individual combustion pulse present. Normally the mean model is sufficient enough for use in processes such as control system design. But in engine performance development aspect, the cylinder-by-cylinder method is better because this method derives from engine geometries, which is very useful for improving and optimizing the engine in the future.
2.2.2 Limit of Physical Properties
There are a number of approaches available when deciding on the basis of a model. Physical equation theoretically describing the system is the most common method since it creates a general model working for many operating areas. Its drawback is that reality might be difficult to describe correctly in theory.
Another common approach bases on the model entirely on measurements. The measured data is stored as a table of two or more dimensions in a so called black box depending on input signals. This approach often provides an accurate result since it is based directly on empirical formulation. However it is only defined for a limited region. A combination of both approaches is commonly used. The main basis of the model rests on physical equations. And empirical equations are used to model certain complexities.
CHAPTER 3
THE MODEL
In this chapter the model is described in detail. The engine to be modeled is a gasoline engine which has no EGR and turbo system. The chosen model bases on the pressure inside the cylinder prediction. There are two main approaches to consider; mean value and cylinder-by-cylinder models. Since the objective of this thesis intends to develop the model which can describe effects of each parameter on the engine performance, 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 Overview
FIGURE 3-1 Model overview
Start
RPM Stroke Bore Conn. Rod Comp.Ratio Spark Angle Heating Value
A/F Ratio Conn. Rod
IVO
IVC k
Liv Div Tw No. Cylinder
Find mean piston speed Find displacement volume Find compression volume Find Temperature of gas in cylinder Find residual mass
θ = -360 θ = 180 No Yes A B C End D Find cylinder volume, Eq.3-1
Find exposed combustion chamber surface area, Eq.3-2
Find exhaust gas temperature, Eq.3-21
Find total heat input, Eq.3-7 Wiebe Function,
Eq.3-4 Heat release, Eq.3-5 Find heat transfer coefficient, Eq.3-10
Find pressure increment, Eq.3-3
Find total heat transfer, Eq.3-8 Find pressure inside
FIGURE 3-2 (CONTINUED) θ = θ + 1 A B ) 1 /( 0 1 2 − ⎥⎦ ⎤ ⎢⎣ ⎡ + ≥ ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ k k Critic T k p p 0 p pT = 0 p pT 0 p pT = ) 1 /( 1 2 − ⎥⎦ ⎤ ⎢⎣ ⎡ + k k k C No Yes Pe D Find valve lift
function, Eq.3-16
Find valve curtain area, Eq.3-14 Find discharge
coefficient, Eq.3-18 Find frictional loss
factor, Eq.3-19 Find charge heating factor, Eq.3-20
Find flow-in mass, Eq.3-15
Find total flow-in mass by integrating
Find pressure that belong to flow-in mass
Find friction loss, Eq.3-22
Find indicated mean pressure, Eq.3-23 Find effective pressure, Eq.3-24 Find effective power, Eq.3-25
Fig.3-1 and Fig.3-2 shows the overview of the model. Engine geometries, such as bore, stroke, compression ratio, etc., are calculated to obtain physical information such 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 of information about heat input. Heat energy needs data from amount of flow in mass and burn characteristic which is described by Wiebe function. Predicted pressure will be used to determine temperature inside cylinder and also heat transfer from cylinder to 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 by mean friction, then work and power will be known finally. The Matlab/Simulink block diagrams of the model are shown in Appendix B. The details of each module are described in following section.
3.2 Crank Slider Model
The volume of the piston cylinder can be determined as a function of crank angle from the compression ratio (ε), the stroke (s), bore (b) and connecting rod length (a). The geometric parameters of the piston cylinder can be described by the crank slider model which is represented in Fig.3-3.
FIGURE 3-3 Piston cylinder and geometries
The equations of volume and area that relate to crank angle are described as following equation [5]: ] ) sin ( cos 1 [ 2 1 ) ( 2 1 2 2 θ a l θ a l V ε-V V d d ⎟ − ⎠ ⎞ ⎜ ⎝ ⎛ − − + + = θ Eq.3-1 ] ) sin ( cos 1 [ 2 2 ) ( 2 1 2 2 2 θ a l θ a l s πb b π A ⎟ − ⎠ ⎞ ⎜ ⎝ ⎛ + − + + = θ Eq.3-2
3.3 Cylinder Pressure Model
This model is derived from the first law of thermodynamics. The pressure is derived as a function of crank angle also [5].
θ θ θ d dV V P k Q d Q V k d dP loss⎟− ⎠ ⎞ ⎜ ⎝ ⎛∂ − − = 1 Eq.3-3 8
θ
d dV
can be determined from Eq.3-1 by taking derivative with respect to the crank angle (θ). The Wiebe function for the burn fraction is used for
θ d Q ∂ . 3.4 Wiebe Function (f)
The mass fraction burned profiles as a function of crank angle in each individual cycle shown in Fig.3-4. It has a characteristic S-Shape. The rate at which fuel-air mixture burns increases from a low value intermediately following the spark discharge to a maximum about halfway through the burning process and then decreases to a close to zero as the combustion process ends.
FIGURE 3-4 Burned mass fraction characteristic
A functional form often used to represent the mass fraction burned versus crank angle curve is the Wiebe function:
⎥ ⎥ ⎦ ⎤ ⎢ ⎢ ⎣ ⎡ ⎟ ⎠ ⎞ ⎜ ⎝ ⎛ Δ − − − = 3 0 5 exp 1 ) ( θ θ θ θ f Eq.3-4
The heat release (∂ ) over the crank angle change (Δθ) is: Q
θ θ d df Q d Q in = ∂ Eq.3-5
Then take the derivative of the heat release function (f(θ)), with respect to crank
angle, being θ d df from Eq.3-4. 3.4.1 Burning Duration (Δθ)
Mixture burning rate is influenced by engine speed. It is well established that the duration of combustion in crank angle degrees only increases slowly with increasing engine speed. Fig.3-5 shows how intervals of burning characteristics are. The burning rate throughout the combustion process does not increase linearly as engine speed. Additionally, increasing in-cylinder gas velocities (e.g. with intake generated swirl) increases the burning rate. Increasing engine speed and introducing swirl both increase the levels of turbulence in the engine cylinder.
(a) (b)
FIGURE 3-5 Effect of engine speed on flame-development angle (a) 10% and 90% burn [6] and (b) 100% burn [9]
However, the swirl characteristic is not able to determine easily depending on manifold geometries and piston head design. Therefore, the results in Fig.3-5a, which obtained from experiment, are used to estimate the burning duration (Δθ) for gasoline.
951 . 39 ) 1000 ( 886 . 19 ) 1000 ( 6189 . 1 2 + + − = Δθ N N for 1000 ≤ N ≤ 6000 Eq.3-6 3.5 Heat Input
Overall heat input can be determined by using heating value of fuel and amount of mass which is drawn into cylinder.
stoich air IVC IVO in m d m HV Q , 1 ) ( + =
∫
θ θ Eq.3-7Values of final mass fraction burned usually are in the range 93% to 98% due to the fact that not all of the fuel is burned [6]. Kuo [3] used 95%. So this value is applied to Qin.
3.6 Air/Fuel Ratio
In internal combustion engines, the air-fuel ratio refers to the proportion of air and fuel present during combustion. The chemically optimal point at which this happens is the stoichiometric ratio. Examples are shown in Table 3-1.
For gasoline fuel, the stoichiometric air/fuel mixture is approximately 14.6 times the mass of air to fuel. This is the mixture that modern engine management systems employing fuel injection attempt to achieve in light load cruise situations. Any mixture less than 14.6 to 1 is considered to be a rich mixture, any more than 14.6 to 1 is a lean mixture. 10 Mass Burn ( × 100) %
TABLE 3-1 Air/fuel ratios of many substances
3.7 Heat Transfer
Generally there are three modes of heat transfer in the combustion engine. Conduction - Conduction in solid matter is caused by molecular movement. The driving physical characteristic is the thermal conductivity. The heat flow in the combustion 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 combustion chamber wall occurs through convection. The velocity and the degree of turbulence of the 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 T4. Radiation is only relevant in the combustion chamber and only short period of time during and after combustion when high gas temperatures are present.
In spark ignition engines, the primary heat transfer mechanism from the cylinder gases to the wall is convection, with only 5% from radiation. Using a Newtonian model, the heat loss to the wall is given by:
Qloss = hA(Tg-Tw) Eq.3-8
When determining the heat release term ( θ
d Q
∂ ) the heat loss to the walls has to be taken into account. Substitute Eq.3-5 and Eq.3-8 in Eq.3-3, the pressure over crank angle now becomes:
θ θ θ d dV V P k T T N hA d df Q V k d dP w g in ⎥⎦− ⎤ ⎢⎣ ⎡ − − − = ( ) 6 1 Eq.3-9 3.7.1 Heat Transfer Coefficient Correlations (h)
The heat transfer coefficient is needed in order to calculate heat loss from the cylinder in Eq.3-8. For design purposes, simplified analyzes are often performed using empirical heat transfer correlations such as Annand, Woschni and Hohenberg. These give at most estimates of the surface-averaged heat transfer coefficient, which are defined in terms of the bulk gas temperature and used with this to calculate surface-averaged or total heat flux. Annand assumed a constant characteristic gas velocity equal to the mean piston speed, while Woschni assumed that the average gas velocity should be proportional to the mean piston speed. Hohenberg examined Woschni’s formula and made changes to give better predictions.
Kleeman et al [10] compared these empirical correlations with Computational Fluid Dynamic (CFD) prediction by using Standard Wall Function (SWF) and Modified Wall Function (MWF) methods. Both SWF and MWF are boundary layer models 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 flow analysis, assuming steady one dimensional flow and ignore the variation within the layer of the fluid thermophysical properties (e.g. density, viscosity, thermal conductivity). The use of SWF in CFD calculation of engine heat transfer generally leads to underprediction of the wall heat flux. So MWF is developed further for using in engine simulation by taking the fluid thermophysical properties into account. The results are shown in Fig.3-6.
FIGURE 3-6 Comparison of surface-averaged heat flux variations predicted by CFD calculations and empirical correlations [10]
The results from [10] showed that the MWF method gave the most accurate results. So using these simple correlations will introduce some errors. However MWF can only be obtained from CFD because the CFD program can calculate the instantaneous local heat fluxes which are not uniform throughout the combustion chamber, while this thesis assumes the charge are burned completely throughout the cylinder at the same time. According to the results in Fig.3-6, the Hohenberg correlation is selected because it is closest to MWF and also there is no empirical form for MWF. The Hohenberg correlation is following equation [4].
8 . 0 4 . 0 8 . 0 06 . 0 ( 1.4) 130 + = − − m g C T P V h Eq.3-10 3.8 Volumetric Efficiency (ηv)
The intake system – the air filter, carburetor, and throttle plate (in a spark ignition engine), intake manifold, intake port, intake valve – restricts the amount of air which an engine of given displacement can induct. The parameter used to measure the effectiveness of an engine’s induction process is the volumetric efficiency (ηv).
Volumetric efficiency is only used with four-stroke engines which have distinct induction process. It is defined as the volume of air which is drawn into the intake system divided by the volume which is displaced by the piston, Eq.3-11. Typical maximum values of volumetric efficiency for naturally aspirated engines are in the range 80 to 90%. The volumetric efficiency for diesels is somewhat higher than the spark ignition engines as shown in Fig.3-7.
d a a v V m ρ η = Eq.3-11
FIGURE 3-7 Volumetric efficiency versus mean piston speed for a four-cylinder indirect-injection diesel and a six-cylinder spark-ignition engine [6] According to Fig.3-7, there are no such a model can predict the trend of the volumetric efficiency exactly because it is affected by the following fuel, engine design and engine operating variables:
1. Fuel type, air/fuel ratio, fraction of fuel vaporized in the intake system, and fuel heat of vaporization
2. Mixture temperature as influenced by heat transfer 3. Ratio of exhaust to inlet manifold pressures 4. Compression ratio
5. Engine speed
6. Intake and exhaust manifold and port design
7. Intake and exhaust valve geometry, size, lift and timings
The manifold and valve geometry design has great effects on the volumetric efficiency since the designs of each engine model have never been the same. So this thesis tries to develop the model in order to predict ηv caused by the different valve
FIGURE 3-8 Effect on volumetric efficiency of different phenomena which affect the air flow rate as a function of speed. Solid line is final volumetric efficiency versus speed curve [6]
The shape of volumetric efficiency can be explained by Fig.3-8. It is affected by many different phenomena. Non-speed-dependent effects (such as fuel vapor pressure) drop the volumetric efficiency below 100% (curve A). Charge heating in the manifold and cylinder drops curve A to curve B. It has greater effect at low engine speeds due to longer gas residence time. Frictional flow losses increase as the square of engine speed, and drop curve B to curve C. At higher engine speeds, the flow into the 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 volumetric efficiency decreases sharply (curve C to D). The induction ram effect at higher engine speeds which occurs from inertia of mixture raises curve D to curve E. Late inlet valve closing, which allows advantage to be taken of increased charging at higher speeds, results in a decrease in the volumetric efficiency at low engine speeds due to backflow (curve C and D to F). Finally, intake and/or exhaust tuning can increase the volumetric efficiency (often by substantial amount) over part of the engine speed range, 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 less than atmospheric pressure due to friction in each part of the intake system. This total pressure 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’s intake process due to the piston velocity variation, valve open area variation, and the unsteady gas-flow effects that result from geometric variations. At high engine speeds, the inertia of the gas in the intake system as the intake valve is closing increases the pressure in the port and continue charging process when the piston slow down around BDC and starts the compression stroke. This effect becomes progressively greater as engine speed is increased.
Back flow – Because the inlet valve closes after the start of the compression stroke, a reverse flow of fresh charge from the cylinder back into the intake can occur as the cylinder pressure rises due to piston motion toward TDC. This reverse flow is largest at the lowest engine speeds. This phenomenon cannot be avoided due to choosing the inlet valve closing time for taking advantage of the ram effect at high speeds.
Tuning – The time-varying inlet flow to the cylinder causes expansion waves to be propagated back into the inlet manifold. These expansion waves can be reflected at the open end of the manifold (at the plenum) causing positive pressure waves to be propagated 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 the nominal at the end of intake process. This will increase the inducted air mass and be described as tuned. This phenomenon can occur in exhaust system also.
Charge Heating - According to the conduction heat transfer, the heat inside the cylinder transfers to inlet manifold via connecting ports and intake valves. This heat is absorbed by air/fuel mixture directly and makes the mixture expand its volume due to increasing of temperature. Finally, the volumetric efficiency decreases automatically. Generally, the residence times, length of inlet manifold and manifold geometries are the 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 a flowing fluid at a certain pressure and temperature flows through a restriction (such as the 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 subsonic upstream conditions as it flows through the smaller cross-sectional area of the restriction. 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 not increase with a further decrease in the downstream pressure environment.
3.8.1 Flow through Valves
The valve is usually the most important flow restriction in the intake and the exhaust system of four-stroke cycle engines. In this thesis considers only the intake valve in order to determine ηv. The mass flow rate through a poppet valve is usually
described 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 gas flow effects are included by means of an experimentally determined discharge coefficient (CD). 5 . 0 / ) 1 ( 0 / 1 0 5 . 0 0 0 1 1 2 ) ( ⎪⎭ ⎪ ⎬ ⎫ ⎪⎩ ⎪ ⎨ ⎧ ⎥ ⎥ ⎦ ⎤ ⎢ ⎢ ⎣ ⎡ ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ − − ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ = − k k T k T R D p p k k p p RT p A C m& Eq.3-12
When the flow is choked, the pressure ratio (
0 p pT
) will not lower than the following value so called critical pressure ratio.
) 1 /( 0 1 2 − ⎥⎦ ⎤ ⎢⎣ ⎡ + = ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ k k Critic T k p p Eq.3-13
For the mass flow of the mixture into the cylinder through the intake valve, p0 is
ambient pressure, and pT is the cylinder pressure [6], [11]. T0 is ambient temperature. For AR, the most convenient reference area in practice is the so called valve curtain
area since it varies linearly with valve lift and is simple to determine [5, 6].
AR =πDvLv Eq.3-14
Eq.3-12 should be converted into a function of crank angle also by dividing with 6N same as Eq.3-9.
5 . 0 / ) 1 ( 0 / 1 0 5 . 0 0 0 1 1 2 ) ( 6 ⎪⎭ ⎪ ⎬ ⎫ ⎪⎩ ⎪ ⎨ ⎧ ⎥ ⎥ ⎦ ⎤ ⎢ ⎢ ⎣ ⎡ ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ − − ⎟⎟ ⎠ ⎞ ⎜⎜ ⎝ ⎛ = − k k T k T R D p p k k p p RT N p A C d dm θ Eq.3-15
Eq.3-9 doesn’t consider about mass flow into cylinder. It determines pressure different based on pressure at previous crank angle. But the pressure inside cylinder is influenced not only by volume variation, incoming mass also. So Eq.3-15 is integrated to obtain amount of mass over a crank angle degree and then finding the pressure 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 pT in Eq.3-15.
3.8.2 Valve Lift
The fundamental the valve lift design is to satisfy an engine breathing requirement at the design speeds. However, it is philosophies and secrecies of each carmaker. One of the valve lift design uses polynomial function, for example, Hermann, McCartan and Blair (HMB) technique [12] uses up to 11th order polynomial functions. The alternative approach is the G. P. Blair (GPB) method [12] which considers jerk characteristic of valve motion.
Which method is employed depends on how smooth of the lift and/or acceleration diagrams are otherwise the forces and impacts on the cam follower mechanism will be considered. In other words, a good mathematical smoothing technique within the valve lift design process is absolutely essential, may be degree by degree level. So this thesis assumes to use cosine function for valve lift instead in order to reduce complexity as following equation:
2 ) cos 1 ( ) (θ = iv,max + ϕ v L L Eq.3-16 180 ) 540 2 ( + + + + − = IVC IVO IVC IVO θ π ϕ Eq.3-17 16
3.8.3 Discharge Coefficient (CD)
Fig.3-9 shows the results of steady state flow tests on a typical inlet valve configuration with a sharp-cornered valve seat. The discharge coefficient based on valve curtain area is a discontinuous function of the valve-lift/diameter ratio. The three segments shown correspond to different flow regimes as indicated. At very low lifts, the flow remains attached to the valve head and seat, giving high values for the discharge coefficient. At intermediate lifts, the flow separates from the valve head at the inner edge of the valve seat as shown. An abrupt decrease in discharge coefficient occurs at this point. The discharge coefficient then increases with increasing lift since the size of separated region remains approximately constant while the minimum flow area is increasing. At high lifts, the flow separates from the inner edge of the valve seat 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 that over the normal engine speed range, steady flow discharge coefficient results can be used to predict dynamic performance with reasonable precision. The averaged equation which obtained from the results in Fig.3-9 is shown as Eq.3-18.
FIGURE 3-9 Discharge coefficient of typical inlet poppet valve [6]
6913 . 0 ) ( 5999 . 2 ) ( 248 . 31 ) ( 13 . 143 ) ( 47 . 190 4 − 3 + 2 − + = iv v iv v iv v iv v D D L D L D L D L C Eq.3-18 3.8.4 Frictional Losses
Takizawa et al [13] made an experiment to measure pressure losses due to friction across the air cleaner, carburetor, throttle and manifold plenum of a standard four-cylinder automotive engine intake system. The results are shown in Fig.3-10. However, the parameters which are used to calculate the pressure losses still unknown, mainly frictional factor for each component. So a frictional loss factor (Cf)
is introduced to adjust p0 in Eq.3-15. The products of Cf and p0 are assumed as shown
in Fig.3-11. The basis of Cf is assumed to be a function of engine speed linearly in
order to reduce complexity as following equation.
986923 . 0 ) 1000 ( 019738 . 0 + − = N Cf for 1000 ≤ N ≤ 6000 Eq.3-19
FIGURE 3-10 Pressure losses in the intake system of a four-stroke cycle spark- ignition engine determined under steady flow conditions [13]
86 88 90 92 94 96 98 100 0 1000 2000 3000 4000 5000 6000 7000 Engine Speed, RPM R em ai n ing P re ss ur e, k P a
FIGURE 3-11 Assumption of remaining pressure after subtracted by pressure losses 18
3.8.5 Charge Heating
As mentioned in Sec.3.8, the residence times, length of inlet manifold and manifold geometries are the main factors which influence how much heat can be transferred to the mixture. However, the length and geometries of manifold cannot be determined easily and the mixture velocity is dynamic behavior. So a charge heating factor (Cheat) is introduced to apply with T0 in Eq.3-15. The assumption is defined that
temperatures inside manifold are in between cylinder wall temperature and ambient temperature. The products of Cheat and T0 are assumed as shown in Fig.3-12, 373 K at
1000 rpm and 308 K at 6000 rpm. The basis of Cheat is assumed to be a function of
engine speed linearly in order to reduce complexity as following. 2953 . 1 ) 1000 ( 043624 . 0 + − = N
Cheat for 1000 ≤ N ≤ 6000 Eq.3-20
260 280 300 320 340 360 380 0 1000 2000 3000 4000 5000 6000 7000 Engine Speed, RPM T em p er atu re I n si d e M an ifo ld , K
FIGURE 3-12 Assumption of temperature inside manifold
3.9 Residual Gas
The residual gas affects volumetric efficiency and engine performance directly, and efficiency and emissions through its effect on working fluid thermodynamic properties [6]. The residual gas is primarily a function of inlet and exhaust pressure, speed, compression ratio, valve timing, and exhaust system dynamics. However, the equation which can describe the magnitude of residual gas is still unknown. So this thesis assumes that the amount of residual gas can be determined by using the ideal gas equation of state, PV =mRT at TDC which is a function of exhaust pressure, compressed volume, exhaust gas molecular weight, and exhaust gas temperature. The exhaust pressure is between 1 and 1.5 atm [3]. So it is selected at 1.5 atm. The exhaust gas molecular weight is 30.4 g/mol.
Caton and Heywood [14] measured cylinder pressure, calculated cylinder gas temperature and exhaust mass flow rate, and measured gas temperature at the exhaust port exit for a single-cylinder spark ignition engine at 1000 rpm. The results of [14] are shown in Fig.3-13. According to Fig.3-13, the exhaust gas temperature at 1000 rpm can be assumed starting at 900 K. The exhaust temperatures along other engine speeds are calculated at point 4 in Fig.2-3a under ideal process. The increments of temperature 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.
21 . 676 ) 1000 ( 49 . 279 ) 1000 ( 9 . 51 ) 1000 ( 3955 . 3 3 − 2 + + = N N N Texh for 1000 ≤ N ≤ 6000 Eq.3-21
FIGURE 3-13 Measured cylinder pressure pc, calculated cylinder-gas temperature Tc, exhaust mass flow rate m& , and measured gas temperature at e
exhaust port exit Tp, for single-cylinder spark ignition engine at speed
= 1000 rpm [14] 0 200 400 600 800 1000 1200 1400 0 1 2 3 4 5 6 7 Engine Speed, RPM E xh au st G as T em p er at u re , K
FIGURE 3-14 Assumption of temperature inside manifold
3.10 Friction
Friction losses influence the indicated power and useful output, the brake power. Barnes-Moss [15] tested several four-stroke cycle four cylinder SI engines between 845 and 2000 cm3 displacement at wide-open throttle. The total friction work per cycle (and thus the friction mean effective pressure) for a given engine geometry will vary with the engine speed as following equation.
97 . 0 ) 1000 ( 15 . 0 ) 1000 ( 05 . 0 2 + + = N N pf for 1000 ≤ N ≤ 6000 Eq.3-22
3.11 Torque & Power
To determine the overall performance, indicated mean effective pressure is used [16]. The indicated mean effective pressure represents the work per combustion cycle normalized by the displacement volume also called specific work. This value indicates amount of maximum available power that can generate from single cylinder.
d mi
V PdV
p =
∫
Eq.3-23The engine has to overcome the friction loss. So the available output becomes:
f mi
me p p
p = − Eq.3-24
Effective power can be determined by following equation.
d me
e N p V
P = 50. ⋅ ⋅ Eq.3-25 Finally, torque can be determined by following relation.
30
NT
Pe =π Eq.3-26
3.12 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 small percentage 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 best advance depends on design and operating conditions. This value is called “Minimum Spark Advance” which will produce the maximum torque at a given operating condition of speed and load of a given engine combination. In most cases the spark advance 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 "knock limits". The fuel octane, camshaft profile and/or the compression ratio will need to be addressed before maximum output can be achieved.
(a) (b)
FIGURE 3-15 (a) Cylinder pressure versus crank angle for over-advanced spark timing (50°), MBT timing (30°), and retarded timing (10°). (b) Effect of spark advance on brake torque at constant speed and air/fuel ratio, at wide open throttle [6]
3.13 Simulation conditions
The conditions for simulation are summarized in Table 3-2.
TABLE 3-2 Parameters for simulation
Parameters Value
Ambient Pressure 1 atm
Ambient Temperature 298 K
Mean cylinder wall temperature 400 K [11, 17] Specific heat ratio (k) 1.3 [3, 6]
Air molecular weight 28.97 g/mol
Air density 1.2 kg/m3
Gasoline molecular weight 114 g/mol Gasoline heating value 44,000 kJ/kg
Gasoline air/fuel ratio 14.6:1
In real engine test (SAE, DIN standard), the performance of engine is measured with all accessories and standard intake and exhaust systems. All parameters which affect to the volumetric efficiency due to using those systems are taken into account as well. Although the temperature inside manifold is varied from 308 to 373 Kelvin due to the heat transfer from cylinder (Sec.3.8.4) which affects to density of air, the air density is still kept constantly at 1.2 because the volumetric efficiency is measured comparing to the ambient condition.
3.14 Alternative Fuels
By researches of many scientists, quantity of petroleum in the world has a limit and it cannot be renewed. 70% of petroleum is consumed in transportation and it is increasing every year. With this rate, the petroleum will deplete from the world one day in the near future. So many countries are searching the other sources of energy to replace the petroleum in long run. Furthermore, improvement of air quality is another objective too.
Many 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 the alternative fuels in analytical aspect. This section is going to mention about the alternative fuels, such as LPG, CNG and alcohol-blended gasoline in details and how the 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 details are still unknown, mainly the combustion duration. So the studies are set to predict the combustion range of these alternative fuels by comparing between the output performances and simulated values which the data are obtained from previous researches. The model is used to explain the volumetric efficiency in order to determine effects of using those alternative fuels too.
The previous researches are selected under specific condition. They have to study the using both gasoline and one of alternative fuels on the same engine in order to compare the information. The gasoline engine model which is developed under this thesis uses parameters of the engine in order to determine the spark advanced angle. With additional information of specific spark advanced setting on each alternative fuels, the burning duration can be determined.
3.14.1 Ethanol-blended Gasoline or Gasohol
Ethanol (ethyl alcohol) and methanol (methyl alcohol) are two types of alcohol fuels. The use of pure alcohols in internal combustion engines is only possible if the engine is designed or modified for that purpose. However, in their anhydrous or pure forms, they can be mixed with gasoline in various ratios for use in unmodified automobile 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 most modern automobiles. However, not enough scientific tests have been done to determine if E10 is harmful to older cars' fuel systems. It has been introduced nationwide in Denmark and Thailand, and will replace high octane pure gasoline in Thailand by 2007.
E20 contains 20% ethanol and 80% gasoline. This fuel is not yet widely used in the world. Since February 2006, this is the standard ethanol-gasoline mixture sold in Brazil, where concerns with the alcohol supply resulted in a drop in the ethanol percentage, previously at 25%. Flexible-fuel cars are set up to run with gasoline in such concentration range and few will work properly with lower concentrations of ethanol. 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 [18]
Properties Gasoline Ethanol E10 E20
Heating value (kJ/kg) 44,000 27,000 41,900 40,000
Al-Farayedhi et al [19] 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 -1 degree for E10 and -6 degree for E20 when comparing to gasoline. And also the engine brake thermal efficiency is improved when compared to the leaded fuel.
3.14.2 Ethanol
Ethanol fuel is an alternative to gasoline. Anhydrous ethanol or ethanol without water can be blended with gasoline in any concentration up to pure ethanol (E100) to reduce the consumption of petroleum fuels, as well as to reduce air pollution. In Brazil, ethanol-powered and flexible-fuel vehicles are manufactured to be capable of operation by burning hydrated ethanol. In addition, flexible-fuel vehicles can run on any mixture of hydrated ethanol and gasoline, as long as there is at least 20% of ethanol. A few flexible-fuel systems, like the Hi-Flex, used by Renault and Fiat, can also run with pure gasoline.
Renault Clio is one of many vehicle models which are sold in Brazil. It is equipped by Hi-Flex technology which allows using pure ethanol as fuel with full efficiency due to the automatic adjustment on engine system. The power curve of Clio increases only 1 horsepower throughout engine speed range when using ethanol. And also, information from [20] indicates that the proper timing for an ethanol engine is five 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 an environmentally "clean" alternative to those fuels. It comprises mainly methane (CH4)
and ethane (C2H6). It is stored and distributed in hard containers, usually cylinders,
and keeped under high pressure. In response to high fuel prices and environmental concerns, compressed natural gas is starting to be used in light-duty passenger vehicles and pickup trucks, medium-duty delivery trucks, and in transit and school buses. The properties of natural gas are shown in Table 3-4
TABLE 3-4 Summaries of major properties of methane, ethane, and natural gas [21]
Properties Methane (CH4) Ethane (C2H6) Natural Gas (CH3.76) [20] Heating value (kJ/kg) 50,100 47,400 49,500
Stoichiometric air/fuel ratio 17.16 16 16.9
Mello et al [21] evaluated the maximum horsepower in many vehicles which were converted to use both natural gas and gasoline, so called “bi-fuel vehicle”. They founded that there is substantial drop in horsepower 13-17% when using the natural gas with electronically lambda control. However, emissions are reduced dramatically. The drop of power came from using of gas mixer which restricts air flow through manifold. And also the natural gas is gaseous fuel which occupies a larger volume than liquid fuel. It causes reduction of ηv. Not only that, spark advanced angle is
needed to increase about 21-25 degree in order to obtain the maximum output due to low burning rate.
3.14.4 Liquefied Petroleum Gas (LPG)
LPG comprises primarily propane (C3H8) and mixed by butane (C4H10). LPG is
mainly used in household, industries and transportation respectively. LPG can be produced from crude oil and natural gas. However, the composition between propane and butane depends on the sources. The characteristics of LPG are summarized in Table 3-5.
TABLE 3-5 Summaries of major properties of propane, butane, and LPG [22]
Properties Propane (C3H8) Butane (C4H10) LPG (97.6%C3H8) [21] Heating value (kJ/kg) 46,000 45,400 45,900
Stoichiometric air/fuel ratio 15.6 15.4 15.6
Caton et al [22] developed a dedicated LPG fueled engine from production car. They founded that torque drops about 10-12% when using LPG without any enhanced modification. The reason of performance reduction came from using the gaseous fuel which decreases ηv. For spark timing, LPG needs an advanced spark timing usually
CHAPTER 4
MODEL VALIDATION AND SENSITIVITY ANALYSIS
This chapter presents and describes results acquired through simulations made with the model. The results are compared to data which appear in vehicle technical manual in order to prove the accuracy.
4.1 Performance Validation
This section shows the results of simulation comparing to test data from 8 engines. All engine geometries are obtained from Mercedes-Benz model year 1969 which are summarized in Appendix A. The model is simulated between 1,000 to 6,000 rpm of engine speed range.
0 20 40 60 80 100 120 1000 2000 3000 4000 5000 6000 Engine Speed, RPM Po w er , k W 0 50 100 150 200 250 300 Tor que , N m Power Power(Sim) Torque Torque(Sim)
0 20 40 60 80 100 120 1000 2000 3000 4000 5000 6000 Engine Speed, RPM Po w er , k W 0 50 100 150 200 250 300 350 Tor que , N m Power Power(Sim) Torque Torque(Sim)
FIGURE 4-2 Simulation results of Mercedes-Benz 250SL
0 20 40 60 80 100 120 1000 2000 3000 4000 5000 6000 Engine Speed, RPM Po w er , k W 0 50 100 150 200 250 300 350 400 Tor que , N m Power Power(Sim) Torque Torque(Sim)
0 20 40 60 80 100 120 140 1000 2000 3000 4000 5000 6000 Engine Speed, RPM Po w er , k W 0 50 100 150 200 250 300 350 400 To rque , N m Power Power(Sim) Torque Torque(Sim)
FIGURE 4-4 Simulation results of Mercedes-Benz 280SE/8
0 30 60 90 120 150 1000 2000 3000 4000 5000 6000 Engine Speed, RPM Po w er , k W 0 80 160 240 320 400 Tor que , N m Power Power(Sim) Torque Torque(Sim)
FIGURE 4-5 Simulation results of Mercedes-Benz 280SL/8
0 30 60 90 120 150 1000 2000 3000 4000 5000 6000 Engine Speed, RPM Po w er , k W 0 80 160 240 320 400 Tor que , N m Power Power(Sim) Torque Torque(Sim)
FIGURE 4-6 Simulation results of Mercedes-Benz 300SEL/8
0 30 60 90 120 150 1000 2000 3000 4000 5000 6000 Engine Speed, RPM Po w er , k W 0 80 160 240 320 400 Tor que , N m Power Power(Sim) Torque Torque(Sim)
