engine seal fluent

DESIGN AND OPTIMIZATION OF THE SCAVENGING SYSTEM OF A

MULTI-CYLINDER TWO-STROKE SCOTCH-YOKE ENGINE

NG TEE NENG

UNIVERSITI TEKNOLOGI MALAYSIA

BAHAGIAN A – Pengesahan Kerjasama*

Adalah disahkan bahawa projek penyelidikan tesis ini telah dilaksanakan melalui

kerjasama antara _______________________ dengan _______________________

Disahkan oleh :

Tandatangan : Tarikh :

Nama :

Jawatan :

(Cop rasmi)

* Jika penyediaan tesis/projek melibatkan kerjasama.

BAHAGIAN B – Untuk Kegunaan Pejabat Sekolah Pengajian Siswazah

Tesis ini telah diperiksa dan diakui oleh :

Nama dan Alamat Pemeriksa Luar : Dr.Victor Selvaratnam Chelliah

Special Projects Unit, CPDD

Level 48, Tower One,

PETRONAS Twin Towers,

Kuala Lumpur City Centre,

50088 Kuala Lumpur.

Nama dan Alamat Pemeriksa Dalam : Prof.Madya Dr. Sanjayan A/L veelautham

Fakulti Kejuruteraan Mekanikal

UTM, Skudai.

Nama Penyelia Lain (jika ada) :

Disahkan oleh Penolong Pendaftar di Sekolah Pengajian Siswazah :

Tandatangan : Tarikh :

Nama : GANESAN A/L ANDIMUTHU

DESIGN AND OPTIMIZATION OF THE SCAVENGING SYSTEM OF A

MULTI-CYLINDER TWO-STROKE SCOTCH-YOKE ENGINE

NG TEE NENG

A thesis submitted in fulfilment of the

requirements for the award of the degree of

Master of Engineering (Mechanical)

Faculty of Mechanical Engineering

Universiti Teknologi Malaysia

APRIL 2006

iii

To my beloved mother and father

iv

ACKNOWLEDGEMENTS

First and Foremost, I would wish to express my profound gratitude to my

supervisor, Prof. Ir. Dr. Azhar Bin Abdul Aziz for his precious guidance and

encouragements during the whole course of this project. All his advices and ideas

have eventually contributed to the success of this project. I am grateful and honored

to have taken up this project as part of the ongoing research in UTM.

I would also like to wish my special thank to Prof. Madya Dr.Rosli Abu

Bakar, Hj. Sairaji Suhadi, En Zulkarnain Abdul Latif, and other research officers at

the Automotive Development Center (ADC) who have helped me throughout my

research work. I also wish to thank my research-mate, Mr. Fong Kok Weng for his

co-operation during this project.

Last but not least, my sincere thanks to Engineer, Mr. Tan as well as all the

collaborators from IA Engineering Sdn.Bhd, in Johor Bahru for their help in the

fabrication of the engine model rig.

v

ABSTRACT

A two-stroke engine complete with a scavenging system, operating with

external pumping mechanism is being developed. The engine comprises of two pair

of combustion chambers and a pair of piston pumps that are integrated onto the

Scotch-Yoke crank mechanism. The Schnurle type loop scavenging arrangement was

selected for the scavenging arrangement for the engine port design. The pump was to

be driven by the engine’s pistons linkages. The significant advantages of this

opposed piston-driven pump concept is the double action of air pumping at every

180° interval of crankshaft revolution. In addition, extensive work using

Computational Fluid Dynamic code simulation tools were applied throughout the

project to ensure that the scavenging port geometry is optimized. Also developed was

a scavenging test rig specifically to verify the simulation results. The unfired tracer

gas sampling method was developed for the scavenging measurement purposes. The

experimental testing was carried out successfully with the use of instrumentations

such as Dewetron High Speed Data Acquisition and crank encoder. Both the

simulation results and experimental results showed good scavenging characteristic,

where the scavenging efficiency is closed to the perfect mixing scavenging model.

The development of the scavenging system will allow for the reduction of the

pollutant emission, and the overcome short-circuiting problem of the two-stroke

engine.

vi

ABSTRAK

Rekabentuk enjin dua lejang telah dibangunakan dengan sistem hapus-sisa

lengkap, di mana ia beroperasi dengan pengepaman udara dari luar. Rekabentuk

enjin ini adalah berpandukan mekanisme Scotch-Yoke yang beroperasi dengan dua

pasang kebuk pembakaran dan sepasang pemampat piston. Susunan sistem hapus-

sisa jenis “Schnurle” telah dipilih untuk mereka pembukaan udara dalam enjin. Pam

ini dipandu secara terus oleh penyambungan omboh yang bersalingan. Kelebihan

mekanisme yang ketara ialah ia dapat menghasilkan dua kali kerja pengepaman pada

setiap 180° putaran aci engkol. Tambahan lagi, kajian yang mendalam pada bahagian

penukaran udara telah dijalankan dengan kod program computer Computational

Fluid Dynamic untuk memastikan rekabentuk pembukaan udara dalam keadaan yang

memuaskan. Sementara itu, rangka uji-kaji sistem hapus-sisa telah dibangunkan

untuk mengesahkan nilai keputusan daripada simulasi komputer. Teknik

pengambilan jenis gas dalam keadaan tanpa bakar, telah dijalankan dengan

kelengkapan alat pengukuran seperti Dewetron High Speed Data Acquisition dan

pengecod engkol. Kedua-dua keputusan daripada simulasi komputer dan ujikaji telah

menunjukkan kecekapan hapus-sisa dalam keadaan yang baik, di mana nilainya

adalah menghampiri model hapus-sisa yang sempurna. Pembangunan sistem hapus-

sisa ini telah menjamin keberkesanannya dalam mengurangkan hasil kotoran

daripadan pembakaran dalam enjin dan juga menyelesaikan masalah short-circuiting

rekabentuk enjin jenis dua lejang.

vii

TABLE OF CONTENT

CHAPTER CONTENT PAGE

ABSTRACT v

ABSTRAK vi

TABLE OF CONTENT vii

LIST OF FIGURES xi

LIST OF TABLES xv

LIST OF APPENDICES xix

1 INTRODUCTION

1.1 Preface 1

1.2 Objectives 2

1.3 Statement of Problem 2

1.4 Hypotheses 2

1.5 Scopes 3

1.6 Methodology 4

1.6.1 Literature Review 5

1.6.2 Design Concept 5

1.6.3 Calculations and Analysis 5

1.6.4 CFD Simulations 6

1.6.5 Development of a Scavenging system

Test Rig 6

1.7 The Gantt Chart 7

2 LITERATURE STUDY

2.1 Internal Combustion Engines 8

viii

2.2 Two Stroke Engine 10

2.3 The Scotch-Yoke Mechanism 12

2.3.1 The Differences between Scotch-Yoke

Engine and Conventional Engine 13

2.3.2 Advantages of Scotch-Yoke Engine 15

2.3.2.1 Size Reduction 15

2.3.2.2 Engine Balance 17

2.3.2.3 Noise, Vibration and Harshness

(NVH) 18

2.3.2.4 Emission 19

2.3.2.5 Efficiency 20

2.3.2.6 Cost 21

2.4 Scavenging Process 22

2.4.1 Cross-Scavenged 23

2.4.2 Loop-Scavenged 24

2.4.3 Uniflow-Scavenged 26

2.5 Scavenging Parametric 28

2.6 Scavenging Mathematical Models 31

2.7 Scavenging Measurement Methods 33

2.7.1 The Global Parameters Measurement

Method 34

2.7.2 The Running Engine Parameter

Measurement Method 35

2.7.3 The Computer Simulation Method 36

2.8 Supercharger 38

2.9 Future Challenges of Two-Stroke Gasoline engine 39

3 ENGINE DESIGN CONCEPT

3.1 Introduction 42

3.2 Scavenging System Design 43

3.2.1 Scavenging Arrangement 45

3.2.2 External Pump Design 46

ix

4 CALCULATIONS AND ANALYSIS

4.1 Engine Component Design 50

4.1.1 Crankcase and Cylinder Block 51

4.1.2 Cylinder Liner s 53

4.1.3 Cylinder Head 56

4.1.4 Chamber 56

4.1.5 The Intake and Exhaust Manifold 58

4.1.6 Reed Valves 59

4.1.7 Piston Pump Design 60

4.2 The Scotch-Yoke Crank mechanism 62

4.2.1 Sliders 62

4.2.2 C-plates 65

4.2.3 Piston Heads 66

4.2.4 Crankshaft 67

5 FLOW SIMULATION AND ANALYSIS

5.1 Introduction 71

5.2 Flow Pattern Static Condition Analysis 73

5.2.1 The Main Port Design 75

5.2.1.1 The Simulation Results 78

5.2.1.2 Conclusion of the Main Port

Design Simulation results 88

5.2.2 The Upsweep Design 88

5.2.2.1 The Simulation Results 90

5.2.2.2 Conclusion of the Upsweep Angle

Design Simulation results 100

5.3 Analysis of the Simulated Scavenging Process 100

5.3.1 The Simulation Results 107

5.3.1.1. Velocity distribution 107

5.3.1.2. Species Transport Mass Fraction

Distribution 112

5.3.2 Discussion on the Dynamic Simulation

Results 117

x

6 FABRICATION OF A SCAVENGING SYSTEM TEST RIG

6.1 Introduction 121

6.2 The Test Rig Components 122

6.3 The Test Rig Set-Up 129

6.3.1 The Scavenging Measurement Results 134

6.4 The Pressure Inside Cylinder 140

6.4.1 Results of Pressure-In-Cylinder Analysis 141

6.5 Scavenging Performance Analysis 147

7 CONCLUSIONS & RECOMMENDATIONS

FOR FURTHER WORK

7.1 Conclusions 149

7.2 Recommendation for Further Works 151

REFERENCES 152

APPENDICES 156

xi

LIST OF FIGURES

FIGURE NO. TITLE PAGE

1.0 Flow chart of project implementation 4

2.1 The different types of reciprocating engine 9

2.2 The two-stroke SI engine operating cycle with crankcase

compression 11

2.3 The gas exchanges process of the crankcase compression

Two Stroke Engine 11

2.4 The crank mechanism of a Scotch-Yoke engine 12

2.5 The application of the SYTech Engine 14

2.6 The comparison of the Scotch-Yoke engine with the

conventional horizontal opposed cylinder engine 17

2.7 The 2nd order noise level of the CMC 422 SYTech engine

at WOT acceleration Cabin Noise 19

2.8 The results of SYTech Fuel consumption and NOx

Emissions advantages compare to conventional engine 20

2.9 The comparison of mechanical losses of the CMC 422

Scotch-Yoke and the conventional boxer engine 21

2.10 The comparison percentage of total Engine costs between

the CMC 422 SYTech and conventional engine 22

2.11 Scavenging arrangements 24

2.12 Various port plan layout of Schnurle type loop scavenging 26

2.13 Physical representation of isothermal scavenge model 28

2.14 a) Perfect displacements scavenging;

b) Perfect mixing scavenging 31

xii

2.15 Benson-Brandham model of trapping characteristic 33

2.16 Schematic diagram of single-cycle scavenge rig with

cylinder block for externally scavenged three cylinders

engine in place 35

2.17 The PIV on the two-stroke engine 36

2.18 The 3D mesh with inlet and exhaust ducts 37

2.19 A supercharged and turbocharged fuel injected

two-stroke engines 38

2.20 The Piston pumps 39

2.21 The future development of a two-stroke engine 40

3.1 Design of a Two-stroke Horizontal Opposed

Scotch-Yoke Engine 44

3.2 The Schnurle loop scavenging 46

3.3 The Piston pump mechanism design 48

3.4 The scavenging process 49

4.1 The cylinders arrangement 52

4.2 The main cylinder block design 52

4.3 Cylinder liner design 53

4.4 Port openings timing 54

4.5 The height of the transfer ports and exhaust port 55

4.6 The cylinder head design 56

4.7 The detail of the hemi-spherical chamber design 57

4.8 The intake manifold design 58

4.9 The exhaust manifold design 59

4.10 The overview of the reed valve assembly 60

4.11 The two ways control by the reed valve design 60

4.12 The piston pump liner design 61

4.13 The volume A and B inside the piston pump 63

4.14 The rotational motion of the slider 64

4.15 The assembly of a pair of the slider and bearings 64

4.16 The C-plate design 65

4.17 The assembly of slider bearing with C-plate 66

4.18 The piston head for combustion process 67

4.19 The piston head for piston pump 67

xiii

4.20 The Crankshaft design 68

4.21 The analysis of the crankshaft balancing 69

5.1 Flowchart of the flow pattern static state analysis. 74

5.2 Schnurle type loop scavenging design 76

5.3 The simulation results of sample A 79

5.4 The simulation results of sample B 81

5.5 The simulation results of sample C 82

5.6 The simulation results of sample D 84

5.7 The simulation results of sample E 85

5.8 The simulation results of sample F 87

5.9 The sweep port design 89

5.10 The design of the upsweep degree of the port 89

5.11 The simulation results of the sample 1 91






5.17 Flow chart for flow simulation in Fluent v 6.1 100

5.18 The engine computational symmetrical domain 102

5.19 The piston surface (moving wall) of the scavenging

process 103

5.20 The convergence of the simulation Work 106

5.21 Velocity contour at 111.6° ATDC and at 136.6° ATDC 108




5.25 Mass Fraction at 111.6° ATDC and at 136.6° ATDC 113

5.26 Mass Fraction at 161.6° ATDC and at 1866° ATDC 114



5.29 The Mass fraction gas O2 versus crank angle 118

5.30 The Scavenging efficiency versus scavenging ratio 120

5.31 The trapping ratio versus scavenging ratio 120

xiv

6.1 The gasket sealing and leakage inspection 122

6.2 Leakage inspection with soap bubble 123

6.3 The machined items of the engine crank mechanism 125

6.4 The perspex material representing the intake manifold 125

6.5 The associated reed valves and cylinder head section 126

6.6 The overview of the motorized scavenging test rig 126

6.7 The gas analyzer probe and Oliver IGD gas analyzer 128

6.8 The Dewetron signal display and crank angle sensor 128

6.9 Digital manometer and Tachometer 129

6.10 Schematic diagram of the scavenging test rig set up 130

6.11 The scavenging measurement arrangement 131

6.12 The gas analyzer probe on the outflow of the system 132

6.13 The illustration of the scavenging measurement 133

6.14 The pressure inlet, P1 versus Engine speed, (rpm) 134

6.15 The Inlet velocity, V1 versus Engine Speed (rpm) 136

6.16 The pumping manifold Pressure, P2 versus Engine Speed 137

6.17 The pumping manifold velocity, V2 versus Engine Speed 138

6.18 The trapped volume ratio of Gas O2 versus Engine Speed 139

6.19 Schematic diagram of the pressure in-cylinder

measurement 140

6.20 The location of the mounting of the pressure transducer 141

6.21 Pressure Variation in chamber A versus crank Angle 143

6.22 Pressure Variation in chamber B versus crank angle 145

6.23 Pressure Variation in piston pump chamber versus crank

angle 146

6.24 The scavenging efficiency versus scavenging ratio 148

6.25 The trapping ratio versus scavenging ratio 148

xv

LIST OF TABLES

TABLE NO. TITLE PAGE

1.1 The Gantt Chart 7

2.1 The dynamic mechanism equation differences

between Conventional and Scotch-Yoke engine 15

2.2 The Packing Advantages between SYTech and

Conventional engine 16

2.3 Classification of different scavenging methods

and their applications 27

2.4 The typical values for the scavenging performance 30

2.5 The fields of application of spark ignition

two-stroke and four-stroke engine 41

3.1 Typical engine specifications 43

3.2 Prediction performance for 4 cylinders 500cc

Scotch-Yoke Engine 43

5.1 The several viscosity model application in the

CFD simulations 72

5.2 The Engine Computation Domain Detail 74

5.3 The Operation parameters for the Cosmos

FloWork 2004 75

5.4 Several samples of the main ports design 77

5.5 The study of the upsweep angle, (°) of the

transfer port 90

5.6 Specification of the optimized Schnurle loop

scavenging design. 100

xvi

5.7 The set up parameters when using Fluent v6.1 104

5.8 The simulation conditions for the scavenging

process analysis 105

5.9 Results of mass fraction 117

5.10 The dynamic results for the scavenging parameter 118

5.11 The standard data for the perfect mixing and

displacement scavenging model 119

6.1 Bill of Material for the engine model design 124

6.2 Specification of the instrumentations 127

6.3 The experimental results for volume A and

volume B 147

xvii

LIST OF SYMBOLS

Yconv = Piston replacement for conventional engine, m

Ysy = Piston replacement for Scotch-Yoke engine, m

vconv = Piston Velocity for conventional engine, m

vsy = Piston Velocity for Scotch-Yoke engine

aconv = Piston acceleration for conventional engine, m/s2

asy = Piston acceleration for Scotch-Yoke engine, m/s2

r = Crank radius, m

L = Conrod length, m

α = Crank angle after top dead center (TDC)

ω = Angular speed of the crankshaft, ˚

SE = Scavenging Efficiency

TE = Trapping Efficiency

CE = Charging Efficiency

SR = Scavenging Ratio

Vi = Intake velocity, m/s

d = Cylinder bore, mm

l , s = Stroke, mm

Vet = Swept volume from TDC to exhaust opening, mm3

Vto = Total volume in the engine cylinder, mm3

Vtr = Trapped volume during TDC, mm3

tl = Liner thickness, mm

N = Engine speed, rpm

σc = Permissible value of tension, 200 kgcm2, Cast Iron

Pm = Maximum combustion pressure, 49.07 kg.cm2

Dl = Inner liner diameter, mm

xviii

n = 1 for two stroke engine

Q = Capacity/cylinder, mm3

ηv = Volumetric efficiency

A = area, mm2

h = Height, mm

Fi = Inertia Force, N

Mi = Moment Inertia, Nm

mu = Unbalanced mass, g

rm = Radius of unbalanced mass from center, mm

Ld = The distance from center, mm

rb = The distance of mass gravity of counterweight, mm

Mref = The unbalance Moment Inertial System, Nm

AM = Angle for main port, ˚

MT = Target point for main point, mm

UPM = Upsweep angle of main port, ˚

UPR = Upsweep angle of rear port, ˚

UPS = Upsweep angle of side port, ˚

DPE = Downsweep angle of exhaust port, ˚

ρ = Density, g/ml

TDC = Top Dead Center

BDC = Bottom Dead Center

ATDC = After Top Dead Center, ˚

γ = Specific heat ratio

σ = Turbulent Prandtl numbers

k = Turbulent kinetic energy

ε = Dissipation rate of turbulent kinetic energy

atm = Pressure at atmosphere condition

CA = Crank angle, ˚

xix

LIST OF APPENDICES

APPENDIX TITLE PAGE

A Balance Shaft Design 157

B1 Orthographic view 158

B2 Exploded View of Engine Model 159

B3.1 Slider R 160

B3.2 Slider L 161

B3.3 Slider Bearing 162

B3.4 Slider Bearing 2 163

B3.5 Crankshaft bearing with thrust 164

B3.6 Crank Bearing M 165

B3.7 Crankshaft 166

B3.8 Crankcase 167

B3.9 Exhaust Manifold 168

B3.10 Crankshaft Bearing 169

B3.11 Intake manifold 170

B3.12 C-platrigmodel1 171

B3.13 C-platrigmodel2 172

B3.14 Piston55K 173

B3.15 Compression rig2 174

B3.16 Piston pump 175

B3.17 Sleeve test 176

B3.18 Reed main body 177

B3.19 Cylinder head 178

B3.20 Block 179

xx

B3.21 Linearslide 180

B3.22 Adapter block 181

C Dynamic Mesh Option Set Up 182

D Scavenging Rig Experimental Data 184

E Dewetron Signal Display Setting 188

F Pressure In-cylinder Data 189

G Tube Adaptor Specifications 194

CHAPTER 1

INTRODUCTION

1.1 Preface

The emphasis of the research project is to design a scavenging system for a

newly conceptualized small capacity (500 cc), multi-cylinder, two-stroke engine

based on the Scotch-Yoke mechanism. The research work on the Scotch-Yoke engine

concept was attempted by CMC SYTECH Corp. of Australia [2] and was proven to

have several advantages i.e. small size, perfect balance, reduction of the engine

weight compare to the conventional reciprocating engine of a same displacement.

The scavenging process in the two-stroke cycle engine has direct influent on

the performance of their combustion processes and remains one of the fundamental

important strategies towards improvement of fuel utilization efficiency and the

reduction of pollutant.

Several CFD simulation analyses have been done to characterize the

scavenging process for the port geometry optimization. In addition, an unfired test

rig for scavenging system measurement has been developed in conjunction with this

research work.

2

1.2 Objectives

The objectives of the research project are:

i. To design an external scavenging system of a two stroke Scotch-Yoke multi-

cylinder engine

ii. To develop a scavenging system test rig to optimize the scavenging process.

iii. To reduce fresh charge short-circuiting problem in the two-stroke engine.

1.3 Statement of Problem

Scavenging process is required in two-stroke engines in assuring the

appropriateness of combustion. However it will also result in the short-circuiting of

fresh charge (flow directly from the engine’s transfer to the exhaust port). The short-

circuiting phenomenon is responsible for the low fuel economy/efficiency and high-

unburned hydrocarbons emission.

1.4 Hypothesis

An external scavenging system is required to retrofit the small capacity multi-

cylinder, two-stroke horizontally opposed Scotch-Yoke engine to improve its

scavenging efficiency and overcome the mixture short-circuiting problem.

3

1.5 Scope

The scopes of work prescribed are as follows:

i. Literature reviews on the two-stroke engine, scavenging systems and

Scotch-Yoke engine concept

ii. Design of a scavenging system for the two-stroke Scotch-Yoke engine.

iii. Computational Fluid Dynamic(CFD) code simulation for the scavenging

flow analysis

iv. Development of an unfired scavenging system test rig

v. Validation of the hypotheses

4

CAD Solid Drawings

CFD simulations

Calculations and Analyses

1.6 Methodology

The methodology applied in the implementation of this project was as follows:

Figure 1.0: Flow chart of project implementation.

Literature study

Design Concept

Results

Database

Documentations and Reports

CFD Evaluation

Development of an unfired

Scavenging system test rig

Optimized

5

1.6.1 Literature Review

The review of recent works is important to provide the understanding of the

advancements of two-stroke technologies such as the scavenging systems and

Scotch-Yoke engine design itself. The previous technical references which are

published in the reputable journals such as Engineering Society for Advancing

mobility Land Sea, Air and Space (SAE) Technical Paper Series, will assist the author

in providing new research methods for the scavenging system development. Besides,

there are several books and publications on two-stroke engines which will provide

first hand knowledge on approaches to engine design and analysis.

1.6.2 Design Concept

With the knowledge obtained from literature study, a design concept of an

external scavenging system which is suited to the design of Scotch-Yoke mechanism,

as well as piston pumps design will be proposed. The loop scavenging arrangement,

which is suitable for small capacity gasoline type two-stroke engine, will then be

applied for the scavenging port geometry design work.

1.6.3 CAD Solid Drawings

It is in the opinion of the author that Computer-Aided Design (CAD)

software, (e.g. SolidWorks 2004) is suitable tool to enable engine parts be designed

and eventually developed. The specification of the engine parts will be shown in

intricate details in finalizing engineering drawings.

6

1.6.4 CFD Simulations

The Computational Fluid Dynamic (CFD) simulation work is an important

approach to predict the characteristic of the gas exchange processes particularly

during the scavenging process. The design of the porting will be improved through

the analysis of a series of simulation results.

1.6.5 Development of an Unfired Scavenging System Test Rig

The fabrication works of the unfired scavenging system test rig was done

with the assistance of a local engineering company. Prior to this, the engine

components detail drawings are prepared for the fabrication works. However, the

assembly of the components into a complete unit was not made by the said company,

but was made by the author in UTM, specifically at the Automotive Development

Center (ADC).

After the engine model was completely assembled, it was simulated for

motion analysis using a specially designed motorized control system. The

instrumentations for the scavenging measurement were installed at the engine model.

A technique call gas sampling method was applied to evaluate for the engine’s

overall scavenging system efficiency.

1.7 Gantt Chart

Planning and execution of the project indicates the milestone of the progress of the design and development work within 5 semesters.

Table 1.1: The Gantt chart

Semesters Planning and Execution

1 2 3 4 5

1. Literature Review Study on the previous technical paper

2. Design Concept Develop the design concept

1. Engine geometry design 3. Calculations and Analyses

2. External pump design

4. CAD Solid Drawings 3D model drawing for the system

5. CFD simulations CFD code simulation for the design optimization

6. Fabrication works Fabrication of the engine model

7. Test Rig Setting Up 1. Setting up the test rig

8. Experimental data analysis Investigation on the Scavenging efficiency

9. Documentations and Reports Summary of the Project

7

Planned & Execution Extend For Execution

CHAPTER 2

LITERATURE STUDY

2.1 Internal Combustion Engines

The internal combustion (IC) engine is a heat engine that converts chemical

energy in a fuel into mechanical energy. Chemical energy of the fuel is first

converted to thermal energy by means of combustion or oxidation with air inside the

engine. This thermal energy raises the temperature and pressure of the gases within

the engine. This expansion is converted by the mechanical linkages of the engine to a

rotating crankshaft, which is the output of the engine.

The most common internal-combustion engine is the piston-type gasoline

engine used in most automobiles. The confined space in which combustion occurs is

called a cylinder. In each cylinder a piston slides up and down. One end of a

connecting rod is attached to the bottom of the piston by a joint; the other end of the

rod clamps around a bearing on one of the throws, or convolutions, of a crankshaft;

the reciprocating (up-and-down) motions of the piston rotate the crankshaft, which is

connected by suitable gearing to the drive wheels of the automobile. Figure 2.1

shows the several type of cylinder arrangement which is in-line engine, V-engine,

W-engine, radial and opposed piston.

9

Figure 2.1: The different types of reciprocating engine (a: single cylinder, b: inline,

c: V-design, d: opposed cylinder, e: W-design, f: opposed piston, g: radial.) [8].

Besides, there are two categories for the internal engine design, which are

spark ignition (SI) engine and diesel engine. Spark ignition engine is an engine which

the combustion process in each cycle is started by use of a spark plug. Diesel engine

is also called as compression ignition (CI) engine which the combustion process

starts when the air-mixture self ignites due to high temperature in the combustion

chamber caused by high compression.

10

2.2 Two-Stroke Engine

Two-stroke engine is a reciprocating engine in which the piston takes over

any valve functions in order to obtain a power stroke for each revolution of the

crankshaft. This involves the use of ports in the cylinder walls which are covered and

uncovered by the movements of the piston. As the piston moves down, it clears these

ports so that the exhaust gasses can exit and fresh charge of mixture can enter at the

same time.

In a crankcase-compression engine, the fresh charge is compressed in the

crankcase by the underside of the working piston, prior to its admission to the

cylinder through the scavenge ducts. The closing and opening of the inlet, scavenge,

and exhaust ports are controlled by the piston itself, and thus in its simplest form, the

present engine requires only three moving parts for each cylinder. This engine

concept benefits greatly from this simplicity and has been used successfully as a

spark-ignition prime mover for more applications than any other two-stroke engine

type. The two-stroke SI engine operating cycle with crankcase compression is shown

in Figure 2.2. In addition, Figure 2.3 shows the gas exchange process of a crankcase

compression two-stroke engine.

In typical two-stroke engine, the air-fuel mixture enters the crankcase through

a reed valve. When the piston is near the bottom of the cylinder, a port is uncovered.

As prior movement of the piston has compressed the mixture in the crankcase, it

flows into the cylinder. Further compression in the cylinder starts as soon as the

piston reverses and covers the ports. At the same time compression is occurring in

the cylinder, movement of the piston has created a vacuum in the crankcase which

draws a fresh charge of mixture from the carburetor into the crankcase. The

compressed charge is fired as the piston reaches top dead center. As the expansion of

the burning charge forces the piston downward, the reed valve in the crankcase

closes and the mixture in the crankcase is compressed. As the piston uncovers the

ports at the bottom of the stroke, compressed mixture from the crankcase enters the

cylinder again. This incoming fresh mixture then assists in pushing the burned gasses

out of the cylinder and the cycle is repeated.

11

Figure 2.2: The two-stroke SI engine operating cycle with crankcase compression.

(a: power stroke, b. exhaust blow down, c. scavenging process, d. compression stroke,

e. combustion start) [8].

Figure 2.3: The gas exchanges process of the crankcase compression

Two-stroke engine [32].

12

The exhaust process of a two-stroke cycle engine differs from that of a four-

stroke cycle engine in that there is no exhaust stroke. Blow down is the same,

occurring when the exhaust valve opens or when the exhaust slot is uncovered near

the end of the power stroke. This is immediately followed with an intake process of

compressed air or air-fuel mixture. As the air enters the cylinder at a pressure usually

between 1.2 to 1.8 atm, it pushes the retaining lower pressure exhaust gas out the

still-open exhaust port in a scavenging process.

2.3 The Scotch-Yoke Mechanism

Scotch-Yoke mechanism converts reciprocating motion to rotary motion. It

was used in steam engines, air compressors and pumps. The horizontal-opposed

Scotch-Yoke engine differs from conventional engines in the crank and connecting

rod areas. The combustion process, fuel system, valve train, induction and ignition

system are basically identical. It replaces the arrangement of connecting rods,

gudgeon pins and pistons in conventional engines with a rigid assembly of two

pistons and two connecting rods and a bearing block (Fig. 2.4).

Figure 2.4: The crank mechanism of a Scotch-Yoke engine [2].

13

The crankpin rotates within the bearing block, which slides up and down

between the parallel surfaces formed by the bases of the two connecting rods. The

crankshaft is conventional with two pistons connected to each crankpin. The

connection of two opposing pistons determines the horizontally opposed layout of

Scotch-Yoke engines.

CMC Power Systems Ltd. Of Australia has developed a Scotch-Yoke engine

technology, called SYTech, for very compact combustion engines with 2 to 12

cylinders. The SYTech engine can be applied to all normal types of combustion

engines with reciprocating piston motion. Prototype engines have been built in two

and four stroke version, also in spark ignition as well as compression ignition

(Diesel). In two stroke engines the firing interval of 180 degrees between the

combustion strokes of opposing pistons simplifies the crank arrangement, but

increases the engine width, if the bottom side of the pistons is to be used for the gas

exchange.

The SYTech engine has shown its applications in the combustion engines

(road, water, air) and mobile power units (electric power unit, compressors). Figure

2.5 shows the application of the SYTech engines.

2.3.1 The Differences between Scotch-Yoke Engine and Conventional Engine

The difference in piston motion of conventional (Conv) and of Scotch-Yoke

(sy) engines can be described by the following equations for piston position, speed

and acceleration. The dynamic mechanism differences between the conventional

engine and Scotch-Yoke engine are shown in Table 2.1. The mechanism equations of

the conventional engine are more complex than the Scotch-Yoke engine. The

complexity of the conventional engine is caused by the distance of piston is defined

of the big end bearing rotational movement and the connecting rod length

replacement. However, the Scotch-Yoke mechanism is simply defined in simple

harmonically sinusoidal motion.

Piston engine

Combustion engine Mobile Power Unit

Road Elec. Power Generator Water Compressors Air

Mobile Bikes

Automotive

Sport boat Small aircraft Engine

Compressor

Both Small City car

Luxury vehicle

Hybrid car

Performance car

Figure 2.5: The application of the SYTech Engine [2].

14

15

Table 2.1: The dynamic mechanism equation differences between Conventional and

Scotch-Yoke engine [2].

Descriptions Conventional Engine Scotch-Yoke Engine

Piston

Replacement αα

2sin

22)(cos rLrLr

convY −++−⋅=

rrYsy −⋅= αcos

Piston

Velocity

−

⋅+⋅⋅=

α

αα

2sin

22

2sin

2sin

rL

rwr

convv

( )αsin⋅⋅= wrvsy

Piston

Acceleration ...)6cos4cos

2cos(cos

64

2

2

+⋅+⋅

+⋅+⋅⋅=

αα

αα

AA

Awraconv

L

rkand

kA

kkA

kkkA

=+=

−−=+++=

...128

9

...16

3

4...

128

15

45

6

53

4

53

2

αcos2⋅⋅= wrasy

2.3.2 Advantages of Scotch-Yoke Engine [3]

The CMC engine outperforms conventional engines in many areas. Some of

the advantages result directly from adopting the CMC SYTech engine, while others

are a secondary consequence of the reduction in weight and size of the CMC engine.

2.3.2.1 Size Reduction

The centre of gravity of the whole engine is close to the centre of the

crankshaft, which improves vehicle stability in a horizontal layout. An additional

16

advantage of the layout with horizontally opposed cylinders is the very short and

rigid crankshaft, which helps to reduce torsional crankshaft vibration, especially in

engines with a large number of cylinders. Because there are no gudgeon pins in the

CMC Scotch-Yoke engine, there is no need to prevent heat from the piston surface

being transferred directly onto the conrod. This lifts many of the traditional

constraints on piston design. Table 2.2 showed the comparison of the packing

advantages between SYTech and conventional engines.

Table 2.2: The Packing Advantages between SYTech and Conventional engine [3].

The Packaging Advantages

Conventional SYTech Differences

Type: Opel 2.0L SY 420

Bore, mm 86 86 -

Stroke, mm 86 86 -

No. of Cylinder 4 4 -

Cylinder Dist, mm 93 93 -

Bore/Stroke 1 1 -

Conrod length, mm 143 113.7 -29 mm (-20%)

Capacity, dm3 1.998 1.998 -

Deck height, mm 219 187.7 -31 mm (-14%)

Length, mm 559 297 -262 mm (-47%)

Height, mm 627 475.4 -152 mm (-24%)

Width, mm 532 667.8 136 mm (26%)

L/R Ratio (geometry) 3.33 2.64 -21%

L/R ratio(NVH) 3.33 Infinite -

Box Volume, dm3 186 94 -92dm3 (-49%)

The packaging volume of Scotch-Yoke engines using the advantages of the

CMC design shows reduced dimensional (except width) and of significant for vehicle

engine compartment packaging can have a boxed volume that ranges up to 35 per

17

cent less than conventional boxer engines, and close to 50 per cent less than in-line

and V-configuration engines as shown in Figure 2.6 [3].

Figure 2.6: The comparison of the Scotch-Yoke engine with the conventional

horizontal opposed cylinder engine [3].

2.3.2.2 Engine Balance

In CMC Scotch-Yoke engines, the horizontal opposed arrangement of the

piston movement, only first order inertial is taking into the consideration. The higher

order inertia influences is omitted because value cos2nθ is always equal to zero. The

piston and conrod assembly moves in a perfectly sinusoidal motion along the

cylinder axis, while the bearing block circles on the crank pit around the crankshaft

axis.

18

To even achieve a compromise with balancing the second order inertia forces

in conventional engines requires two balance shafts, both of which have to be driven

at twice the engine speed. Every SYTech engine can be perfectly balanced with a

maximum of one balance shaft that rotates at the same speed as the engine.

2.3.2.3 Noise, Vibration and Harshness (NVH)

Perfect balancing of the inertia forces, improved torque uniformity and

minimal piston slap – all contribute to the improvement of NVH. This is reflected in

the test results for crankcase vibration, where the CMC Scotch-Yoke engine has much

lower vibration amplitudes. It remains the case regardless of engine load and over the

whole operating range. Noise analysis tests conducted on the CMC Scotch-Yoke

engine proved that the linear bearing itself does not increase the overall noise

emissions. Full balancing means the elimination of higher order influences, which

together with improved torque uniformity and reduced piston slap, leads to the

engine’s vibration-free running and reduced noise levels. Less vibration imply fewer

secondary resonance problems. Vibration test results measured on an engine

dynamometer with acceleration sensors mounted on the generator bracket of the

conventional and the Scotch-Yoke engine demonstrate the smooth operation of the

SYTech engine and support the subjective impression already the first CMC422

prototype engine made, when it was run for the first time. The reduction in vibration

amplitudes is significant at all speeds and over the whole load range [3].

Figure 2.7 showed the significant of the lower secondary order noise level of

a 2.2-liter 4-cylinder SYTech engine at the Wide-open Throttle (WOT) acceleration

Cabin Noise compare to the conventional 4-cylinder engines.

19

Figure 2.7: The secondary order noise level of the CMC 422 SYTech engine at

WOT acceleration Cabin Noise [3].

2.3.2.4 Emission

Reaction kinetics calculations for diesel engines indicate reduced NOx at high

loads. Test results for spark ignition engines demonstrate significantly lower NOx

under part load conditions.

The comparison of results for Lambda equal to one on the conventional CMC

Scotch-Yoke engine showed an average reduction of 30 percent in the NOx emissions.

The CMC Scotch-Yoke technology allows a 4-cylinder engine to run smoothly at idle

speeds as low as 550 rpm. Therefore, the fuel consumption and emissions can be

lower than with conventional engines, which idle at between 7 and 800 rpm. Figure

2.8 shows the comparison of fuel consumption and NOx emissions advantages

between SYTech engine and the conventional engine of similar capacity [3].

20

Figure 2.8: The results of SYTech fuel consumption and NOx emissions

advantages in relation to conventional engine [3].

2.3.2.5 Efficiency

The lower frictional losses not only reduce piston and cylinder wear, but also

reduce engine fuel consumption. Combustion simulations by a German engine

Research & Development Company, FEV, as well as testing by CMC Research at the

University of Melbourne, show that improved fuel consumption can be achieved

based on a lower possible idle speed, in addition to the savings caused by lower

frictional losses. The frictional losses associated with CMC’s additional linear

bearings are more of an offset by the benefits from the reduced numbers of main and

conrod bearings, the elimination of gudgeon pins and the lower piston friction. In

motoring tests the mechanical loss in the CMC engine was less than in the

conventional boxer engine, especially at higher speeds. Even more substantial was

the improvement in the mechanical efficiency in the CMC engine over the

21

conventional engine. Figure 2.9 shows the comparison of mechanical losses of the

CMC 422 Scotch-Yoke and the conventional boxer engine.

Figure 2.9: The comparison of mechanical losses of the CMC 422

Scotch-Yoke and the conventional boxer engine [3].

2.3.2.6 Cost

Most of the manufacturing processes involved in building a SYTech engine

are the same as those used in the manufacture of conventional engines. In most

respects the engine work on the same principles as all other internal combustion

engines. The major difference is the crank mechanism, which overcomes many

disadvantages of the conventional crank mechanism, as well as being slightly

cheaper to build. CMC’s engineers have conducted a detailed cost analysis of

building the CMC-422 (four cylinders, 2.2 liter) engine, taking into account all

differences in material, machining and labor costs. Figure 2.10 shows the comparison

percentage of total engine costs between the SYTech engine and conventional engine.

22

Figure 2.10: The comparison percentage of total Engine costs between

CMC 422 SYTech and a conventional engine [3].

2.4 Scavenging Process

Scavenging process is the process where the cylinder’s burned gases are

replaced with a fresh charge using both the high blow down pressure of the expanded

combustion gases and the fluid dynamics of the incoming charges. This process

requires only a fraction of the piston’s stroke to complete, with the exhausting and

recharging events occurring simultaneously, and is critical to ensure that the cylinder

gases are adequately prepared for the next combustion cycle. Scavenging system is

defined as a method used to accomplish the charge-changing process in a two-stroke

engine.

There are two general methods of putting air into the cylinders: through

normal intake valves, and through intake slots in the cylinder walls. In a conventional

crankcase-scavenged two-stroke engine, the combustion products from the previous

23

cycle are forced from the cylinder with a new air/fuel charge. This charge is

compressed in the crankcase by the underside of the piston and then enters the

cylinder when the piston uncovers the transfer port. Unfortunately, the exhaust port

is opened during the entire time that caused the part of the air fuel mixture to “short

circuiting” through the cylinder during the scavenging process. This is the major

source of the high hydrocarbon emissions from crankcase-scavenged engines. When

short-circuiting occurs, lower scavenging efficiencies result even though the volume

occupied by the short-circuiting flow through the cylinder does displace an equal

volume of the burned gases.

Another phenomenon which reduces scavenging efficiency is the formation

of pockets or dead zones in the cylinder volume where burned gases can become

trapped and escape displacement or entrainment by the fresh scavenging flow. These

un-scavenged zones are most likely to occur in region of the cylinder that remains

secluded from the main fresh mixture flow path. Several methods for charging the

cylinder have been proposed. Scavenging arrangements are classified as illustrated in

Figure 2.11.

2.4.1 Cross-scavenged

The transfer and exhaust ports are opposite one another. A deflector on the

piston as shown in Figure 2.11(a) routes the fresh charge in the direction of the arrow

and expels the residual gases from the previous stroke. However, the flow follows

the direction of the wall at the first instant only. Proper piston head design is required

to assure that the intake air deflects up without short-circuiting and leaving a stagnant

pocket of exhaust gas at the head end of the cylinder. At piston bottom dead center, it

pursues the shortest path, with the result that a considerable amount of fresh gas is

expelled instead of the residual gas. Due to the very high charge losses, cross

scavenging is used with inexpensive, light-duty engines only.

24

Figure 2.11: Scavenging arrangements [4].

2.4.2 Loop-scavenged

The difference between the loop-scavenged two-stroke cycle engine and the

cross-scavenged is the design of the piston head. The loop-scavenged piston is flat

because the intake parts are located directly across from each other and 90˚ from the

exhaust port. The entering gas streams travel across the piston, up the far side of the

barrel and curl over and down to complete the scavenging process. This resulting

turbulence cleans the combustion chamber of all exhaust gases. The fresh gases

25

flowing into the cylinder from ports on either side of the exhaust port are directed

upward in the direction of the opposite cylinder wall and expel the exhaust gases

from the cylinder as shown in the center diagram. The scavenging losses are less than

with cross scavenging; however, a small proportion of the fresh gases are expelled

directly, in spite of the necessary diversion. A core of residual gas remains at the

center of the cylinder. Loop scavenging is more favorable for gasoline injection,

where in principle the exhaust and transfer ports are interchanged.

A different arrangement, where the exhaust ports are above the scavenge

ports (MAN-type loop-scavenging system), is shown schematically in Figure 2.11

(b). In this design, the fresh charge stream is directed toward the unported wall, flows

toward the cylinder head, changes its direction, and continues toward the exhaust

port. The long path of the entering charge requires high momentum jets and one

would expect, therefore, that this type of engine perform better at wide-open throttle

(WOT). For this reason, this MAN-loop scavenging system is well suited to diesel

engines where load is controlled by the amount of fuel injected rather than a throttle

valve.

Another method that avoids the use of the troublesome deflector piston was

developed by Schnurle in Germany about 1926. In this approach, the fresh charge is

directed toward the opposite side of the cylinder to the exhaust port, across a piston

with an essentially flat top. Instead of the single scavenge port placed diametrically

opposite the exhaust port, a pair of scavenging ports were located symmetrically

around the exhaust port on the same level as the exhaust port as shown in Fig

2.11(c). In this arrangement, the fresh charge path is shorter than in the MAN-type

loop scavenging. The Schnurle loop-scavenging system is better at throttled

conditions, and mixing between the fresh charge and burned gases is reduced. This

type of scavenging system is widely used in small-bore SI engine. Figure 2.12 shows

the various port plan layout of the Schnurle type loop scavenging.

26

Figure 2.12: Various port plan layout of Schnurle type loop scavenging [5].

2.4.3 Uniflow-scavenged

Intake ports are in the cylinder walls and exhaust valves in the head (or intake

valves are in the head and exhaust ports are in the wall, which is less common). This

is the most efficient system of scavenging but required the added cost of valves. The

exhaust gases are expelled from the cylinder longitudinally, scavenging thus being

improved still further. However, because of the high thermal loading, exhaust valves

are rarely fitted into the cylinder head, except for instance, in two stroke diesel

engines. The piston controlling the exhaust port is slightly in advance of the inlet port

piston. The exhaust time is thus shortened and displaced with respect to the transfer

time by such an amount that when the transfer port opens, the overpressure in the

cylinder has already been eliminated and the exhaust port closes well ahead of the

transfer port.

27

The classification of different scavenging methods and their applications is

shown in the Table 2.3.

Table 2.3: Classification of different scavenging methods [4].

Method

Advantages Drawbacks

Applications

Cross

Good scavenging at partial throttling and low speeds Low engine volume for multi cylinder arrangements Low manufacturing cost

High bsfc at high throttle opening and high speeds High tendency to knock limits compression ratio

Small outboard engines, and some other specific applications

Loop, MAN-type

Good scavenging at Wide Open Throttle (WOT) Low surface to volume ratio combustion chamber Low manufacturing cost

Poor scavenging at part-throttle operation

Large-bore marine CI engines

Loop, Schnurle-type

Good scavenging at WOT and medium engine speed Fair scavenging at part throttle and other than medium engine speeds Low manufacturing cost

High bsfc at part throttle operation

SI engines for a large variety of applications

Uniflow, exhaust valve

Very good scavenging at WOT for high stroke-to-bore ratio Excellent bsfc

Need for exhaust valves; thus more complex and higher manufacturing cost

Large-bore low-speed CI marine and stationary engines

Uniflow, opposed piston

Very good scavenging at WOT for high stroke-to-bore ratio

Need for mechanical coupling between two crankshafts

Sometimes used in large-bore low-speed CI marine engines

28

2.5 Scavenging Parametric

For the same power generation, more air input is required in a two-stroke

cycle engine than in four-stroke engine. This is because some of the air is lost in the

overlap period of the scavenging process. A quantitative discussion of the two-stroke

cycle scavenging process requires precise terminology and an appropriate set of

parameters. The parameters are used to describe the progress of the gas exchange

process are scavenging efficiency, trapping efficiency, charging efficiency and

delivery ratio.

The simple theories of scavenging all postulate the ideal case of scavenging a

cylinder which has a constant volume, Vcy, as shown in Figure 2.13, with a fresh air

charge in an isothermal, isobaric process. In Figure 2.13, the basic elements of flow

are presented. The incoming scavenge air can enter either a space call the

“displacement zone” where it will be quite undiluted with exhaust gas, or “mixing

zone” where it mixes with the exhaust gas, or it can be directly short circuited to the

exhaust pipe providing the worst of all scavenging situations.

Figure 2.13: Physical representation of isothermal scavenge model [6].

29

The following parameters are used to describe the progress of the gas

exchange process. [6]

i. The scavenging efficiency (SEv), which indicates to what extent the burnt

residuals have been replaced with fresh charge at any given instant.

cy

ta

exta

ta

exataa

taa

v

V

V

VV

V

VV

V

SE

=

+=

+=

=

ρρ

ρ

chargecylinder trappedof Mass

retainedair delivered of Mass

(2.1)

ii. The trapping efficiency (TEv), which defines the amount of short-

circuiting of fresh charge to the exhaust

as

ta

v

V

V

TE

=

=air delivered of Mass


(2.2)

iii. The charging efficiency (CEv), which represents the ability of the

engine to fill the cylinder trapped volume

cy

ta

cya

taa

v

V

V

V

V

CE

=

=

×=

ρ

ρ

densityambient volumetrapped


(2.3)

30

iv. The delivery (scavenging) ratio (SR), compares the actual mass of

delivered fresh charge at any given instant to the total amount

required in an ideal charging process, the reference mass.

cy

as

V

V

SR

=

=densityambient x volumetrapped

cycleper air delivered of Mass

(2.4)

In the ideal scavenging process (there is no short-circuiting of fresh

charge occurring), it is clear from manipulation of the above equations that

the charging efficiency and scavenging efficiency are identical:

vv SECE = (2.5)

v

vv

SR

SETE = (2.6)

Table 2.4: The typical values for the scavenging performance [8].

Typical scavenge performance

results

Typical values

1 Scavenging efficiency 0.6 < ηse < 0.9

2 Trapping efficiency 0.65 < ηte< 0.8

3 Charging efficiency 0.5 < ηce < 0.75

4 Delivery ratio 0.5 < ηce < 0.75

31

2.6 Scavenging Mathematical Models

There are two simple scavenge models suggested by Hopkinson, i.e. i.) Pure

displacement scavenging, and ii.) Perfect mixing scavenging [4]. These models are

used to predict realistic values for charging efficiency and the scavenging efficiency.

Both of these are based on the constant volume, isothermal ideal. Pure displacement

scavenging assumes that the fresh charge entering the cylinder displaces the residual

exhaust gas without mixing with it and without any short-circuiting of the fresh

charge until the cylinder is completely scavenged. [5] Figure 2.14 shows the

scavenging model concept of the perfect displacement scavenging and perfect

mixing scavenging.

Figure 2.14: a) Perfect displacement scavenging; b) Perfect mixing scavenging [4].

This process is also known as ‘perfect scavenging’ and may be defined as [5]:

SE = SR, when SR < 1; SE = 1, when SR > 1 (2.7)

Perfect mixing scavenging assumes that as each volume increment of fresh

charge enters the cylinder, it is instantly and completely mixed with the rest of the

cylinder contents. At the same time, an identical volume increment of the resultant

mixture exits through the exhaust port. The ‘perfect mixing’ process can be

expressed as:

SReSE

−−=1 (2.8)

32

Although neither model can be considered a true representation of the

scavenge process in a firing engine they are useful for the assessment of

experimentally attained scavenge data. It is, of course, impossible to better the

scavenge efficiency of the pure displacement model. The perfect mixing model, on

the other hand, does not represent a boundary to poor scavenging. Direct short-

circuiting will, in theory, allow the scavenging efficiency to be zero, irrespective of

the scavenge ratio. Generally, at low scavenge ratios, well-designed cylinders have

scavenging efficiencies that tend towards those calculated for pure displacement

scavenging and at high scavenge ratios the scavenging efficiency fails between that

predicted by the pure displacement and perfect mixing models.

Another theoretical model is called “Benson-Brandham” model. Benson-

Brandham model has suggested combining the perfect displacement and perfect

mixing model. That first part is to be perfect displacement until the air flow has

reached a volumetric scavenging ratio value of SRpd, then the perfect scavenge

volume is mixed together at that point, with include short-circuiting factor, σ. The

“Benson-Brandham” process can be expressed as [6]:

i. when, 0 < SRv < (1-σ) SRpd;

vv SRSE )1(1 σ−−= (2.9)

ii. When, (1-σ)SRv > SRpd;

vSRSRpd

pdv eSRSE)1((

)1(1σ−−

−−= (2.10)

Figure 2.15 showed the “Benson-Brandham” model compared to the perfect

displacement and perfect mixing model of the trapping characteristics.

33

Figure 2.15: Benson-Brandham model of trapping characteristic [6].

2.7 The Scavenging Measurement Methods

There are several types of the measurement methods to study the scavenging

flow in the cylinder chamber. We can classify these methods in three groups:

i. The global parameters measurement method

ii. The running engine parameter measurement method

iii. The computer simulation method

34

2.7.1 The Global Parameters Measurement Method

The global methods mainly deal with the cylinder itself and the inlet and

exhaust ports. The aim is firstly to characteristic the cylinder on a specific test bench

and then to test quickly design modifications before building the final design. These

methods are related to the wind tunnel, single cycle gas testing apparatus, single

cycle hydraulic testing apparatus.

Extensive research has already been conducted into optimizing the porting

layouts of two-stroke engine cylinders. One of the techniques developed at The

Queen’s University of Belfast for evaluating scavenging is a unique experimental

method described as the “single cycle scavenge test”. Although the test does not

reflect the actual scavenge process in a firing engine, it is a sufficiently useful

procedure to have become an industrial standard for scavenges evaluation [5]. Single

cycle similarity tests are frequently used to modify port geometry in order to improve

the engine’s scavenging characteristics. In configuring these tests similar geometric

ratios, as well as Reynolds and Euler number are generally used.

Gas concentration sampling provides a convenient and reliable way of

determining the scavenging and trapping efficiency of the operating engine. This

method can only be performed on scavenging systems that have been designed and

constructed, and therefore provides little direction during the design process [35].

Figure 2.16 shows the schematic diagram of single-cycle scavenge rig with

cylinder block for externally scavenged three cylinder engine in place; testing the

centre cylinder.

35

Figure 2.16: Schematic diagram of single-cycle scavenge rig with cylinder block for

externally scavenged three cylinders engine in place [5].

2.7.2 The Running Engine Parameter Measurement Method

These methods take into account a real running engine. The engine is

modified to let the tools for capture image and visualize the phenomena inside the

engine. The tools include the complicated and expensive optical access

instrumentations such as the endoscopy system, optical fibers system, Laser Doppler

Anemometry (LDA) system, or Particles Imaging System (PIV) system. The

advantages are possible visualization in several planes simultaneously, the

characterization of the evolution of scavenging with time. The disadvantage is the

measurement equipment, which allows the optical access to the engine is expensive

and complex [7]. Figure 2.17 shows the PIV imaging system on a two-stroke engine

to measure the scavenging and combustion process.

36

Figure 2.17: The PIV on the two-stroke engine [5].

2.7.3 The Computer Simulation Method

These methods are quite new, and still need complete experimental

validation. However, their use can give plenty of information. The multi-dimensional

computer simulation includes the one-dimensional, two-dimensional and three-

dimensional calculations.

The one-dimensional is to calculate the acoustic behavior of the different

areas of the engine through the burn and unburnt (air, fuel vapor, mixture) gases

passed. These codes may also provide the boundary conditions for multi-dimensional.

37

The advantages are that it allows a quick description of the engine as a whole, and

that acoustic waves are taken into account. It is a useful tool for parametric studies

(engine speed, geometric modification). However, the numerous parameters have to

be calibrated with the help of measurement.

The two-dimensional calculation is for the case of 3D geometries, which can

be considered as 2D, because the geometry is axis symmetrical or the third

dimension is large relative to the other two. The advantages are in the ability to solve

turbulent fluid mechanics equations, taking into account interaction between gas,

liquid, and model combustion development.

The three-dimensional is to calculate fully 3D geometries is terms of

aerodynamics, injection and combustion. It involves solving fluid mechanics

equations in three directions in space. The calculation being similar to that 2D, both

the models and the method of solving the equations are the same. The advantage is

that the 3D aspect of the flow is well addressed. This allows in particular the

scavenging of the burnt gases by the fresh charge to be calculated, and thus

scavenging ratio and efficiency. Figure 2.18 shows the advanced computing 3D mesh

arrangement for engine cylinder.

Figure 2.18: The 3D mesh with inlet and exhaust ducts [5].

Transfer duct

mesh

Exhaust duct

mesh

Chamber

mesh

38

2.8 Supercharger

Since the two-stroke cycle gas exchange process occurs when both the

exhaust and the scavenge ports are open, the pressure inside the cylinder is normally

above atmospheric pressure. This gas exchange or scavenging process requires the

fresh charge be supplied to the engine cylinder at a high enough pressure to displace

the burned gases from the cylinder. At the same time, the pressure should be low

enough to minimize the scavenging air pumping work.

Superchargers may be mechanical driven or be driven by the engine exhaust.

The supply of fresh air for the scavenging process is by a blower or turbocharger

(Figure 2.19) directly to the scavenging ports without use of the crankcase as an air

pumping system.

Figure 2.19: A blower and turbocharger in a fuel injected two stroke engines [6].

39

Another type of supercharger which could be applied is the displacement

pump. The engine drives the piston pump, which draws air through the carburetor

and delivers it through a pipe to the top of the combustion chamber. When the main

piston approaches BDC, the piston moves upward to compress the fresh charge

inside the pumping cylinder. When the exhaust ports are exposed, the cylinder

pressure falls, the excess pressure above the top valve causes this valve to open, and

fresh mixture is introduced into the main cylinder. One important purpose of such an

auxiliary cylinder is to create asymmetrical port timing, relative to BC, and thus

minimize any backflow through the scavenging ports [4]. Figure 2.20 shows the

piston type pumps in the engine design.

Figure 2.20: The Piston pumps [11].

2.9 Future Challenges of Two-Stroke Gasoline Engine

The exhaust of internal combustion engines is one of the major contributions

to the world’s air pollution problem. Recent research and development has made

major reductions on engine emissions, but a growing population and a greater

Pump

Combustion

chamber

Pump

Combustion

chamber

Pump

40

number of automobiles mean that the problem will exist for many years to come.

During the first half of the 1900s, automobiles emission was not recognize as a

problem, mainly due to the lower number of vehicles. As the number of automobiles

grew along with more power plants, home furnaces, and the population in general,

air pollution became an ever-increasing problem.

In two-stroke engine, several approaches have been developed. One of the

major breakthroughs has been use of the electronic fuel injector in place of the

carburetor. Sophisticated electronics is beginning to be used in two-stroke engines

for injection timing, engine management and emission control. Sensor and electronic

control units are used in the Direct Fuel Injection (DFI) system to manage injection

timing and minimize exhaust emission. The conventional carbureted two-stroke

engine will probably survive in a transitory stage with the addition of an exhaust

catalyst. Nevertheless, this solution has a limited potential in term of pollutant

emissions reduction and presents thermal problems of the exhaust and catalyst,

difficult to solve without any improvement in term of fuel economy. Therefore, the

need of a new generation of two-stroke engine will extensively use direct injection

technology to solve the problem of excessive unburned hydrocarbons emission due

to fuel short-circuiting, and to take benefit of the two-stroke cycle principle

advantages of low pumping and friction losses for high efficiency and low NOx

emissions. Figure 2.21 shows the possible routes from the existing small carburetor

two-stroke engine to clean long term future engine [43].

Figure 2.21: The future development of a two-stroke engine [37].

41

Table 2.5 showed two-stroke engines often used in non-road applications and

in transportation are used for two or three wheeler transportation.

Table 2.5: The fields of application of spark ignition engine [37].

Type of Application 2-Stroke

(%)

4-stroke

(%)

1 Chainsaw 100 -

2 Marine Outboards 100 -

3

Industrial Engines

i. 30 – 100 cm3

ii. 100 – 150 cm3

iii. > 150 cm3

100

50

-

-

50

100

4 Mopeds 50 cm3 100 -

5

Motorcycles and Scooters

i. 125 cm3

ii. 125 – 349 cm3

iii. 350 – 449 cm3

iv. 450 – 749 cm3

v. 750 cm3

70

60

10

1

-

30

40

90

99

100

6 Automotive - 100

CHAPTER 3

ENGINE DESIGN CONCEPT

3.1 Introduction

In conjunction with the conceptual design of small capacity two-stroke

Scotch-Yoke engine, the project in which the author is involved, specifically aims to

create a simple and efficient scavenging system to provide for the eventual feature of

efficient gas exchange processes. The scavenging system requires not only high

delivery ratio and raise the density of the air intake, but also considers the other

factors such as size, weight reduction and low development cost.

The two-stroke Scotch-Yoke engine design concept with external scavenging

system is definitely a new and unique engine development, due to recently the

Scotch-Yoke Engine manufacturer, CMC Power System Ltd of Australia has much

more focused on developing four-stroke cycle of Scotch-Yoke Engine.

The overall design of the Scotch-Yoke engine which has been conceptualize

and is currently being developed at the Automotive Development Center (ADC) is

shown in Figure 3.1. The engine concept is to incorporate several auxiliaries such as

the Capacitive Discharge Ignition (CDi) System, Direct Injection (DI) fuel system,

43

oil sump lubrication system and the cooling system respectively. The general

specifications of the engine design are shown in Table 3.1.

Table 3.1: The general engine specifications.

No. Descriptions Detail

1 Engine type Two-stroke gasoline engine

2 Cylinder Arrangement Horizontal Opposed

3 Number of cylinders 4

4 Total Displacement, cc

500

5 Bore x Stroke, mm 57.5 x 48.0

6 Dimension, Lx H x W mm 540.60 x 444.50 x 435.00

7 Weight, kg 43.8

Besides, the prediction of this engine performance has been simulated by the

researchers in Automotive Development Center, UTM with software GT-Power v6.

Table 3.2 shows the simulation result of engine performance for 4 cylinder 500cc

Scotch-Yoke Engine [51]

Table 3.2: Prediction performance for 4 cylinder 500cc Scotch-Yoke Engine [51].

No Specification Detail

1 Maximum Brake Power(kW) 37.8

(8000rpm)

2 Maximum Brake Torque(Nm) 51.6

(8000rpm)

3 Best Specific Fuel Consumption,

BSFC, (g/kWh) 445

(8000rpm)

4 Brake Mean Effective Pressure, bar 6.49

5 Power to Weight ratio, (kg/kW) 1.16

Direct Injection

System

Capacitive Discharge

Ignition System

Opposed Cylinder Block

Cooling

System

Alternator

Crankshaft

Figure 3.1: Design of a Two-stroke Horizontal Opposed Scotch-Yoke Engine.

Intake

System

Oil Sump

44

45

3.2 Scavenging System Design

A good scavenging system is anticipated to produce better scavenging

process inside the engine chamber. The methods which applied during the design

process include two-stroke engine design considerations, scavenging system

alternatives, as well as to run analysis and testing for prototype. In this designing

scavenging system, the valve system is not omitted for the compact design of two-

stroke engine. In addition, an external pump is required to boost the air charge intake.

The crankcase compression is not suitable for use, because it will increase the engine

block length.

3.2.1 Scavenging Arrangement

There are several types of scavenging arrangements explored for example, i.)

cross-scavenged, ii.) loop-scavenged and iii.)Uniflow-scavenged. From the literature

study, Schnurle-type loop scavenging is more favorable for SI engine application, if

compared to the MAN-type loop scavenged and Uniflow scavenged arrangement [4].

In addition, the manufacturing cost for the loop scavenging system is lower than the

Uniflow scavenging system too. In addition, the loop-scavenged losses are less than

that of cross-scavenged at the condition at the wide open throttle and high speed of

the two-stroke engine.

The Schnurle type loop scavenging arrangement is selected for the Scotch-

Yoke engine design. In general, the Schnurle type loop scavenging design is

illustrated in Figure 3.2.

46

Figure 3.2: The Schnurle loop scavenging [4].

3.2.2 External Pump Design

The two-stroke Scotch-Yoke multi-cylinder engine is to be equipped with an

external air boost pump. The pump is to be driven by the engine’s pistons linkages. It

comprises of the compression piston and cylinder that would integrate with the

Scotch-Yoke crank mechanism. The advantages of system are due its lighter material

and of small size.

The piston pump is directly connected to the crank, therefore able to produce

boost pressure at a very low rpm. The C-plate type piston linkage is able to produce

double action pumping in each cylinder block at every 180° interval. The design for

the piston-type pumping scavenging system is illustrated in Figure 3.3.

47

The multi-cylinder engine will have two pairs of opposed cylinders like any

other boxer engine, and a pair of the opposed piston-driven cylinders for charging of

mixture into the main combustion chamber. The piston-driven pump design results

the double action of air pumping for the gas exchange into the cylinder every one-

half of crankshaft revolution. Each piston pump has two sealed volumes of

compression. The compressed volume starts to pump the fresh air into the

combustion chamber when the transfer ports are opened.

Figure 3.4 shows the double action of the piston pump design at one-half of

crankshaft revolution. The fresh airflow will be split into two halves of the opposed

cylinders. The reed valves are used to control the airflow exchange. When route A is

at compression stage, route B will be at fresh charge induction stage. The piston

pumps will induce the fresh air into the combustion chamber as soon as the transfer

ports are opened. Route A and B will always switch their function for every 180° CA

interval.

Figure 3.3: The Piston pump mechanism design.

Fuel Injection

System

Crankshaft

Piston Pump

Fuel Injection

System

48

Figure 3.4: The scavenging process. 49

CHAPTER 4

CALCULATIONS AND ANALYSES

4.1 Engine Components Design

Several engine components were designed in conjunction with the small

capacity Scotch-Yoke engine such as cylinder block, cylinder head, piston, liners,

port openings, and intake and exhaust manifold and reed valves. To start the design

process, the typical design parameter i.e. the range of the bore-to-stroke ratio of

between 1.2 – 0.9 was given due to consideration [13]. In this case, the value of 1.2

was chosen due to larger bore size could reduce the overall length of engine.

mm

lld

d

Vl d

48

)2.1(

)4(125000;2.1

)4(,lengthstrokeEngine

2

2

=

==∴

=

π

π

(4.1)

Therefore the cylinder diameter is equal to 57.5 mm.

For the crank radius:

mml

r 242

48

2,radiusCrank === (4.2)

51

Piston replacement is calculated as:

24cos24

cos,treplacemenPiston

−=

−⋅=

α

α rrY sy

(4.3)

The engine’s trapped compression ratio is determined as:

( ) ( )

3

2

81.78632

35.108581.265.574

,VolumeTrapped

mm

VVV cvettr

=

+=

+=

π

(4.4)

24.7

35.10858

81.78632

ratio,nCompressioTrapped

=

=

=

cv

trtr

V

VC

(4.5)

To determine the geometry compression ratio of the engine,

51.12

35.10858

35.10858125000

ratio,ncompressioGeometry

=

+=

+=

cv

cvdr

V

VVC

(4.6)

4.1.1 Crankcase and Cylinder block

The Scotch-Yoke engine configuration is of horizontal opposed cylinder

arrangement. The engine external pumps are positioned at the middle of the engine’s

cylinder arrangement. The firing order that suits the Scotch-Yoke engine is (1, 4) - (2,

3), where there will be double combustion processes occurring at every 180°CA

degree interval. Figure 4.1 illustrates this cylinders arrangement.

52

Figure 4.1: The cylinders arrangement.

For high-power-to-weight feature the cylinder block is usually made of cast

iron or Aluminium Alloy. The same case is applied to this engine future development

work, but the unfired test rig in this research project is only applied with Perspec

material. The liners are force-pressed into the chamber slot. There are also passages,

incorporated into the engine for the pumping and water coolant passages. Reed valve

seats are located at the middle of the block for the induction and pumping process of

the piston pump. The overall design of the cylinder block is shown in Figure.4.2.

Figure 4.2: The main cylinder block design.

Cylinder Chamber slot

Piston pump chamber

slot

Coolant Passage

Reed valve

seating Pumping

duct opening

Pumping duct

opening

53

4.1.2 Cylinder Liners

Most of the gasoline engines will use grey cast iron for liners. This material

has the desired casting and machining qualities, and possesses adequate mechanical

feature plus attractive mechanical properties such as strength, toughness and wear

resistance [14]. For this work, the liner chosen is of wet-type, which is forced fitted

into the cylinder chamber slot. The liner design is shown in Figure 4.3.

Figure 4.3: Cylinder liner design.

The following is the calculation for the liner thickness, tt:

mm

DPt

c

lml

7

)200(2

)75.5(07.49

2

=

=

=σ

(4.7)

The intake velocity for a two stroke cycle engine is as below:

( )

( )

1

2

6

75.47

4

02.060

9.08000101251

60 velocity,Intake

−

−

=

×=

=

ms

xx

A

nQNV

s

vi

π

η

(4.8)

Transfer

Ports

Exhaust

port

54

For the minimum transfer port area, Atp:

270.349

7465.47

60

8000125000

,areaportTransfer

mm

x

V

QA

i

tp

=

=

=

(4.9)

In two-stroke engines, the transfer port and exhaust port opening and closing

are controlled by the piston movement. Similar feature is adopted here. The port

openings and closing for this engine is shown in Figure 4.4.

Figure 4.4: Port openings timing.

55

Transfer port height, htp:

mm

htp

20.10

125cos242448

)cos2424(48

=

+−=

−−= θ

(4.10)

Exhaust port opening height, hex:

mm

hex

90.21

95cos242448

)cos2424(48

=

+−=

−−= θ

(4.11)

Figure 4.5 illustrates the height for the ports design. The exhaust port height

is higher than the transfer ports because the exhaust port must always be open first

before the transfer ports.

Figure 4.5: The height of the transfer and exhaust ports.

BDC

TDC

48 mm

10.2 mm

Exhaust

Port Transfer

Ports

21.9 mm

56

4.1.3 Cylinder Head

The cylinder head is assembled on top of the cylinder block. For this type of

engine there is no provision for poppet valve. However it provides the housing for

fuel injectors (direct fuel injection system) for future expansion. A gasket is

sandwiched between the block and head to provide for tight sealing between these

engine parts. There are provisions for reed valve mountings for the regulation of the

air intake and pumping process. Also provided are the slots for water passages

specifically for the cooling of the cylinder head. Figure 4.6 illustrates the cylinder

head design.

Figure 4.6: The cylinder head design.

4.1.4 Chamber

In typical two-stroke engines, the hemispheric chamber geometry is the most

commonly applied for loop-scavenged system [4]. The chamber is of symmetrical

Reed Valve seats

Water

Jacket

Hemisphere

Chamber

Piston pump

chamber

57

design. It is also an open chamber due to the concavity of the cylinder head. The

hemi (abridgement of hemispherical) chamber is very popular in high performance

automobiles. This chamber geometry is applied for the reference engine.

When the piston approaches TDC (at the end of the compression stroke), the

volume around the outer edges of the combustion chamber will be reduced to a small

value. The gas mixture occupies the volume at the outer volume radius of the

cylinder is forced radial inward as this outer volume is reduced to zero. This radial

inward motion of the gas mixture is called squish. [8] During combustion, the

expansion stroke begins and the volume of the combustion chamber increases. This

reverse squish helps to spread the flame front during the latter part of combustion.

Figure 4.7 shows the view of the hemisphere chamber shape design.

Figure 4.7: The detail of the hemi-spherical chamber design.

Squish

Area

58

4.1.5 Intake and Exhaust Manifold

The engine’s intake manifold consists of a pair of intake duct and a pair of

pumping ducts. The intake ducts are for the fresh charge induction purpose, while the

pumping ducts are to pump in the fresh charge into the cylinder chamber. Figure 4.8

shows the intake manifold design for the double action pumping.

Figure 4.8: The intake manifold design.

The other manifold is the exhaust manifold, which consists of a pair of steel

exhaust ducts. The exhaust manifold is mounted to the exhaust opening of the

cylinder block for the scavenging process. Figure 4.9 shows the exhaust manifold

design.

Intake

Duct 2

Pumping

Duct 2

Intake

Duct 1

Pumping

Duct 1

Route A

Route B

59

Figure 4.9: The exhaust manifold design.

4.1.6 Reed Valves

Four pairs of reed valves are incorporated specifically to control the mixture

intake. Each pair has two-way controls of the air intake and outlet. It is designed

specifically for the double pumping of the engine’s piston pump. During induction

process, one side of the reed valve petal will lift to permit fresh charge to flow into

the pumping chamber. Consequently, during the pumping process, another reed valve

will lift to allow the fresh charge to flow into the engine cylinder. The reed petal

thickness is set at 0.2-0.4 mm, where the material could be steel, or carbon fiber. In

test rig development, the carbon fiber is applied for reed valve petals. Figure 4.10

shows the reed valve assembly, which consists of i.) Main body, ii.) limiter and iii.)

petal design. Figure 4.11 fully explains the principle of operation of the reed valve.

Openings that mount to

cylinder block

60

Figure 4.10: The overview of the reed valve assembly.

Figure 4.11: The two ways control by the reed valve design.

4.1.7 Piston Pump Design

According to typical crankcase compressed two-stroke engines, the compress

ratio for engine capacity above 500 cm3 the compression ratio always set above 1.55

[6]. In this exercise, each engine cylinder’s capacity is 125 cm3, therefore the

compression ratio is reasonably set at 1.5.

Limiter Petal

Main

body

Lifting

Induction

Pumping

61

Piston pump swept volume, Vpp = 1.5 x 125000

= 187500 mm3 (4.12)

mmx

xd p 70

48

4187500,diameter bore pumpPiston =

=

π (4.13)

The liner thickness and material type are similar to the cylinder liner.

However, the diameter size is larger than the cylinder liner because the piston pump

liner is designed to adapt to the bore piston pump. The inner diameter is set at 70mm.

It requires the openings for the reed valve seating that controls the intake and

pumping of the air charge. Figure 4.12 illustrates the design of the piston pump liner.

Figure 4.12: The piston pump liner.

Reed valve

seats

Reed valve seats

62

4.2 Scotch-Yoke Mechanism

The Scotch-Yoke mechanism consists of i.) Slider, ii.) C-plates, iii.) Piston

heads and iv.) Crankshaft. The Scotch-Yoke mechanism converts the reciprocating

motion of the piston to rotational sinusoidal motion, which allows the piston to

repeat its movement in horizontal plane. The crank mechanism directly influences

the size of the crankcase and cylinder block. The consideration of the clearance

design for the component assemblies is important to allow the free motion of the

slider and piston. The inner body of the crankshaft, slider and C-plate is drilled with

a lubrication oil passage for reduction of wear friction.

When the piston moves from TDC to BDC, fresh charge will be induced into

the chamber. Subsequently, when the piston moves from BDC toward TDC, the fresh

charge is forced into the combustion chamber. Figure 4.13 shows the piston pump

chamber which is a combination of volume A, pumping volume, and volume B.

Volume A and volume B is required for the seating of the reed valve opening.

4.2.1 Sliders

The slider moves along the locus of the rotational that would convert the

sinusoidal motion to the linear piston movement. The suitable material for slider is

high carbon steel. Figure 4.14 shows the locus of the rotational of the slider. Figure

4.15 on the other hand shows a pair of journal bearings which is mounted inside the

slider.

Figure 4.13: The volume A and B inside the piston pump chamber.

Volume

A

Volume

B C-Plate

Slider

TDC BDC

Crankcase

Piston

63

64

Figure 4.14: The rotational motion of the slider.

Figure 4.15: The assembly of a pair of the sliders and bearings.

Locus of the

rotational

Journal

Bearing

Slider

65

4.2.2 C-plates

A pair of C-plate provides the sliding plane for the slider. It is also used to

thread joint with the piston head. The suitable material for the C-plate is Alloy Steel

(Cr 0.5-1.1wt %). In test rig development, only Aluminium material is applied for C-

plates assembly. Figure 4.16 shows the C-plate design for the proposed engine. In

addition, the assembly of the slider with the C-plates is illustrated in Figure 4.16.

Figure 4.16: The C-plate design.

Sliding

plane

66

Figure 4.17: The assembly of slider bearing with C-plates.

4.2.3 Piston Heads

There are two types of piston heads design, i.e. i.) Piston head for combustion

chamber and ii.) Piston head for piston pump. The piston material should meet

certain requirements such as high hot strength, good thermal expansion and good

resistance to surface abrasion to reduce the skirt and ring groove wear. The material

for the actual piston fabrication could be either Aluminium alloy or cast Iron. In

unfired test rig, only Aluminium material is applied for piston fabrication. Figure 4.19

show the piston heads design for the engine.

Slider

bearing

A pair of C-plate

67

Figure 4.18: The piston head for combustion process.

Figure 4.19: The piston head for piston pump.

Doom

Surface

Tap for C-plate

assembly

Piston ring

groove

Piston skirt

Bore

diameter

Flat

surface

Lubrication

Oil Passage

Tap for C-plate

mounting

Piston ring

groove

68

4.2.4 Crankshaft

There are three crank journals on the crankshaft for the housing of the sliders.

The journals are suited to 180º of rotation to adapt the horizontal opposed cylinder

design. The journal radius distance from the origin of the crankshaft is equal to half

of the stroke engine design. Figure 4.20 shows the crankshaft design.

The inertial force for crankshaft balancing is given as:

)2()(

InertialsecondaryInertialprimaryFForce,Inertial

222

i

θθ Cosl

rwmCosrwm mu

mu +=

+=

(4.14)

Figure 4.20: The Crankshaft design.

The moment inertial is given as:

di rLmwM2,InertialMoment = (4.15)

Counter

Weights

Crank Journal 1

Crank Journal 2

Crank Journal 3

69

Figure 4.21: The analysis of the crankshaft balancing. (Where L1, L4, L6 = 23.5 mm, L2, L3 = 45 mm, and L5 = 59.5 mm respectively)

The secondary inertial force is neglected because the cos2θ is equal to zero

for this opposed cylinder engine. The total mass of crank mechanism for piston pump

is assumed equal to total mass for piston combustion, mp to simplify the calculation.

From the engine design, the total mass of the crank mechanism with high

Carbon Steel material, m1 is 2463.11g.

The calculation for the counterweight design:

gmmmmrw

mmrw

mmmmrw

mmmmrw

Cosrmw

FF

balanceF

x

11.2463,0}{,Therefore

}{

)}2()2{(

)]}180cos()2[()]0cos()2{[

)(

,ForceInertial

)(0,ForceInertial

2121

2

21

2

1221

2

1221

2

2

===−

−=

+−+=

+++=

Σ=

=

=

θ

(4.16)

70

For the calculation of the moment inertial at the reference plane:

2

1

1

2

1

2

1

2

21

1

2

2

2

65431431211

2

54323222

2

048.141042611.2463

62.572

SystemInertialMomentunbalanceThe

062.572

62.572;

)62.5804()5232(

)]()()([

)]()()([,InertialMoment

)(0,InertialMoment

wgm

mw

M

mwM

mwMmm

mwmwM

LLLLmLLmLLmrw

LLLmLmLmrwM

balance

ref

ref

ref

ref

b

ref

==

=

=

=−

−==

−+=

+++++++

−+++++=

=

Q

Q (4.17)

SystemInertialmomentunbalance

048.1410426

)]8675.62)(39.14(53.779[2

][2,shaftbalanceforInertialMoment

2

2

2

=

=

=

=

ω

ω

wLrmM bsbsbsbs

(4.18)

Therefore, a balance shaft at the reverse speed is required to solve the

unbalance moment Inertial. A pair of counter weight is designed with mbs=779.53g

each and rbs = 14.39mm with Lbs = 62.87675mm respectively. The detail of the

balance shaft design is illustrated in Appendix A.

CHAPTER 5

FLOW SIMULATIONS AND ANALYSES

5.1 Introduction

Computational Fluid Dynamics (CFD) is a modeling technique which has

been widely employed to describe and predict the processes that occur within IC

engines. These fluid dynamic-based techniques solve partial differential equations

for the conservation of mass, momentum, energy and species concentrations

respectively. With the recent advances in meshing techniques, boundary treatments,

and computer hardware, all have enabled more accurate computations of the gas flow

process to be done with ease and precision. Utilizing the advantages of CFD codes,

especially the ability to visualize the in-cylinder flow behavior in engine operation

has provided valuable insight into the means of optimizing the scavenging system.

In spite of these advantages it is an unfortunate fact that no single turbulence

model is universally accepted as being superior for all classes of problems. The

choice of a turbulence model will depend on several considerations such as the

physics encompassed in the flow process, the established practice for a specific class

of problem, the level of accuracy required, the available computational resources,

and the amount of time available for the simulation respectively [20].

72

Table 5.1 shows the strengths and the weaknesses of the several viscosity

model applications in CFD codes.

Table 5.1: Several viscosity model application in CFD simulations [20].

In the engine cylinder, the flow will involve a complicated combination of

turbulent shear layers, recirculation regions and wall boundary layer [4]. One

approach to the solution of turbulent flow is often referred to as k-epsilon model

which basically is to estimate the effect of the viscosity of the fluid [6]. The

application of the standard k-epsilon model is applied in engine model simulation

because of its robustness, economy, and reasonable accuracy for a wide range of

turbulent flows explain its popularity in industrial flow and heat transfer simulations.

The advantages of the approaches are that they are able to solve turbulent

fluid mechanics equations (Reynold Average Navier-Stokes (RAN) with K-epsilon

model), which take into account the interaction between gas and liquid and

subsequently model combustion development process. It is very useful for parametric

73

studies and is attractive due to reasonable cost implication [42]. Two equation

models for Standard k-ε model are defined as follows:

i.) Turbulent Kinetic Energy:

(5.1)

ii.) Dissipation Rate:

(5.2)

5.2 Flow Pattern Static Condition Analysis

The scavenging system is to be incorporated into the 500cc two-stroke

Scotch-Yoke engine. Therefore a computation domain engine model is drawn

accordance with its geometry with the CAD software, SolidWorks 2004. Table 5.2

shows the engine computational domain detail.

The Cosmos FloWork 2004 is linked to the SolidWorks 2004 user interface as

the third party software. The analysis of the flow pattern is under steady (static)

condition, where the piston is hold in stationary and set to the Bottom Dead Center

(BDC). This approach could provide a better understanding into the evaluation the

effectiveness of scavenging port arrangement.

74

Table 5.2: The Engine Computation Domain Detail

Descriptions Detail

1 Bore x Stroke, mm

57.5 X 48.0

2 Cylinder capacity/cylinder 125 cc

3 Scavenging system

Schnurle loop, 5 ports

4 Transfer port opening 125° ATDC

5 Exhaust port opening 95° ATDC

6. Trapping compression ratio 7.24

Several samples of the main port geometry design regarded to the Schnurle

loop scavenging has been done with CFD simulation. The evaluation and selection of

a high quality sample of the main port is done. The next step was the optimization of

the selected main port design with analysis of the effect of the several upsweep

degree design. The methodology for the simulation work is shown in Figure 5.1.

Figure 5.1: Flowchart of the flow pattern static state analysis.

CFD simulation on several samples of main port geometry design

Optimize the selected main port geometry design

from several upsweep degree samples analysis

Final design

Comparison of a good flow pattern from several main port geometry samples

Comparison of a good flow pattern from several upsweep degree samples

75

With Cosmos FloWork 2004, the computational domain is automatically

generated from the solid modeling. Every sample of the main ports design is defined

in separate computation domain to run the simulation analysis. The operating

parameters for each simulation works is shown in Table 5.3.

Table 5.3: The Operation parameters for the Cosmos FloWork 2004

Specifications Detail

1 Flow type Gas (air)

2 Viscosity Laminar /Turbulent model

3 Thermodynamic parameters Static Pressure: 101325 Pa

Temperature: 293.2 K

4 Turbulence parameters Turbulence intensity and length

Intensity: 10 % Length: 0.002 m

5 Input parameters Total pressure

Inlet = 1.2 atm, exhaust = 1 atm

6 Mesh Automatic generation of mesh based on Solidwork CAD

7 Solver Standard Solution Adaptive mesh generation for improve accuracy

5.2.1 The Main Port Geometry Design

The port geometry design refers to the work on the main, the side and rear

ports to generate for Schnurle loop-scavenging pattern as shown in Figure 5.2.

76

Figure 5.2: Schnurle type loop scavenging design [6].

The main port geometry is considered to be the main factor that influences

the flow pattern. The port is located just beside the exhaust port, and it has become

the first design priority. The empirical guidance provides a good starting point for

this scavenging port geometry design to be carried out. There are some potential for

empirical guidance for the author and they are as follows [6]:

i.) The upsweep angle for main port, UPM is rarely larger than 10º

ii.) The value of Angle for Main port, AM2 is usually between 50º to 55º

iii.) The target point for MT2 is at between 10 to 15% of the cylinder bore

dimension.

iv.) The target point for MT1 is approximately on the edge of the cylinder bore

v.) The port is tapered to provide an accelerating flow through the port, for

instances AM1 is greater than AM2, and AM1 is rarely larger than 70%

vi.) The larger the angle, AM1, the more the target point, MT1, the farther outside

the cylinder bore is the target point, MT1. The range of the values for AM1 is

usually more than 50º but less than 70º.

77

To understand the effectiveness of main port design, several sample design

was simulated under the similar operating condition. However, the parameter in this

main port design was the design of AM1, AM2 and MT1 and MT2 parameters. In

order to narrow down the simulation works, some parameters are set default value

with accordance to empirical guidance [6]. This includes the side and rear ports,

where the side-side bar, (sbar) at 25 mm and length of side port, Lbar is set at 10mm,

the exhaust port width is set at 55% of the bore diameter. On the other hand, the

effect of the upsweep degree is the next important parameters that influence the main

port design. The upsweep for main port is set initially at 15º, while the rear port is set

at 60°. There are six samples for the main ports design has been successfully

analyzed with CosmosFlowork 2004. The information of these samples is illustrated

in Table 5.4.

Table 5.4: Several samples of the main ports design.

Sample Main port specifications, (º)

AM1 AM2 MT1 MT2

A 65 55 21.2 7.2

B 65 50 20 10

C 60 55 22.6 8.6

D 65 50 22.6 12.3

E 60 50 22.6 8.6

F 65 50 20 12.3

78

5.2.1.1 Simulated Results

Figure 5.3 illustrates the simulated results of sample A. In Figure 5.3 (i), the

result of the velocity vector distribution has indicated that the direction of flow will

move towards the upper section of the cylinder. This phenomenon is widely

believed to scavenge the residual gas which is a by-product of the previous

combustion process.

Figure 5.3 (ii) on the other hand shows the flow pattern at the top trajectories

of the flow pattern has very disorder condition. This may due to the side flow which

was restrained by the flow coming from the main port.

Figure 5.3 (iii) shows the trajectories as viewed from the side of sample A.

The flow trajectories of the main and side ports show the flow to be towards the rear

side of chamber wall, subsequently resulted in the formation of the looping flow.

However, trajectories at the main port have relatively low lifting flow capability

which may cause the short-circuiting of the mixture to occur.

The unsatisfactory flow trajectories which were described in Figure 5.3 (ii)

and 5.3 (iii) have caused the fresh charge not being able to reach the upper side of

chamber thus fail to flush the residual gases. The fresh charges did not lift up, and

this may resulted in AMT 2 degree being too large. The upper flow design is

important because it reflects the quantity of fresh charge to replacement the residual

at the upper chamber.

i. Velocity vector distribution at symmetrical plane ii. The trajectories (top view) of the flow pattern

iii. Trajectories (side view) of flow pattern

Figure 5.3: Simulated results of sample A.

Good Flow toward upper chamber

Low lifting of flow

Flow pattern at

disorder condition

Good Looping flow trajectories

79

80

In Figure 5.4 (i), the result shows the velocity vector distribution of the flow.

It has illustrated the direction of flow toward upper part of the cylinder. The upper

flow helps to scavenge the residual gas due to the combustion process of the previous

engine cycle.

Figure 5.4 (ii) shows the flow pattern at the top trajectories of the flow pattern

which depicts quite symmetrical flow pattern condition. But, there is a reverse flow

at the right main port which may contribute to the worse situation of the short-

circuiting phenomenon. The design of AMT1 for main port may be inadequate.

Besides, Figure 5.4 (iii) shows the trajectories from the side view of the

sample B. The trajectories flowing in the main and side ports have vectored towards

the rear side of chamber wall and generate looping flow. However, there is an un-

scavenged zone present in this chamber. This un-scavenged zone may cause a

portion of the residual gas to trap, thus attributed to the decrease in the scavenging

efficiency.

Figure 5.5 shows the simulation results of sample C. In Figure 5.5 (i), the

results of velocity vector distribution have illustrated the direction of flow toward

upper cylinder. This upper flow helps to scavenge the residual gas which was again

due to the combustion process.

Besides, Figure 5.5 (ii) shows a reverse flow at the right main port that has

caused the slight short-circuiting problem. This is because by AMT 1 of the main

port is not enough to create the upper flow toward the chamber.

On the other hand, Figure 5.5 (iii) shows the trajectories from side view of

sample C. The looping flow towards the upper part of chamber reflects that the

porting design is satisfactory.

i. Velocity vector distribution at symmetrical plane ii. The trajectories of the flow pattern iii. Trajectories(side) of flow pattern

Figure 5.4: Simulated results of sample B.

Un-scavenged

Zone

Good Upper Flow toward chamber

Reverse flow

toward exhaust port

81

i. Velocity vector distribution at symmetrical plane ii. The trajectories of the flow pattern iii. Trajectories(side) of flow pattern

Figure 5.5: Simulated results of sample C.

Reverse flow

toward exhaust

Good upper flow

of blow down

Good Upper flow

toward chamber

82

83

Figure 5.6 shows the simulation results of sample D. In Figure 5.6 (i), the

result of velocity vector distribution illustrates the direction of flow toward the upper

part of the cylinder. The upper flow helps to scavenge the residual gas which was

resulted from the combustion process.

Besides, Figure 5.6 (ii) shows the flow pattern at the top trajectories of the

flow pattern has very disorder condition. There is low flow lifting and reverse flow

to exhaust port. The low lifting flow also has resulted insufficient fresh charge flow

to upper chamber to scavenge the residual gas, thus the scavenging efficiency is

decreased.

Figure 5.6 (iii) shows the trajectories at side view of the sample D. The

trajectories flow of the main port and side port has flow towards the rear side of

chamber wall and generated looping flow. The right side port has generated the good

looping flow pattern. There is low lifting and reverse flow at the main port causes the

short-circuiting problem and decreases the scavenging efficiency.

Figure 5.7 shows the simulation results of sample E. In Figure 5.6 (i), the

result of velocity vector distribution has illustrated the direction of flow toward upper

cylinder. This upper flow helps to scavenge the residual gas which resulted from the

combustion process.


flow pattern is at symmetrical condition. But, there is reverse flow at the right main

port that has caused the worse situation of the short-circuiting.

Figure 5.7 (iii) shows the trajectories at side view of the sample E. The

trajectories flow of the main port and side port has flow towards the rear side of

chamber wall and generated looping flow. The looping flow trajectories have shown

in satisfactory condition. But, the reverse flow generated by the main port has caused

the short-circuiting problem.

Figure 5.6: Simulated results of sample D.

Low flow lifting

Good upper flow

toward chamber

Good flow toward

upper chamber

Reverse flow to exhaust port

84

i. Velocity vector distribution at symmetrical plane ii. The trajectories (top)of the flow pattern iii. Trajectories(side) of flow pattern

Figure 5.7: Simulated results of sample E.

Good looping of scavenging flow


Symmetrical

flow pattern Reverse flow

to exhaust port

85

86

Figure 5.8 shows the simulation results of sample F. In Figure 5.8 (i), the

result of velocity vector distribution has illustrated the direction of flow toward upper


combustion process.


flow pattern also have symmetrical condition. Besides, this sample has resulted with

good lifting flow toward the upper part of chamber. This flow pattern helps to

scavenge the residual gases which resulted from the combustion process.

Figure 5.8 (iii) shows the trajectories of the main port and side port has flow

towards the rear side of chamber wall to generate a looping flow pattern. The looping

flow trajectories have resulted the better scavenging process where the fresh charge

scavenges the residual gas.

i. Velocity vector distribution at symmetrical plane ii. The trajectories (top)of the flow pattern iii. Trajectories(side) of flow pattern

Figure 5.8: Simulated results of sample F.

Symmetrical

Flow pattern Good lifting flow

toward upper chamber


87

88

5.2.1.2 Conclusion of the Simulated Main Port Design Results

From the samples results shown all sample A to E have encountered the same

problem of the reverse flow, toward the exhaust port which caused the short-

circuiting problem to occur. Sample F (in Figure 5.8) has produced in good upper

flow toward chamber, and upper flow for the blow down process. As such, the main

and rear ports are hereby considered have achieved the adequate design of main port

geometry to achieve the adequate scavenging flow pattern. The upper flow is

important in this engine port design. This is because how well the flow toward the

upper chamber, reflected that the fresh charge is going to replacement the residual at

the upper chamber, and how much the fresh charge will remain at the upper chamber

during the scavenging process has great influence to increase the scavenging

efficiency. The upper flow will generate a swirl at the chamber, the flow thereby will

then withdraw the residual at the upper chamber, and pass through to the exhaust

port. Due to the higher degree of ATM2, the entry of the flow is thereby flow upper

to chamber.

5.2.2 The Upsweep Angle Design

The next step is the optimization of the upsweep angle of sample F, which

was obtained from the main port simulation results. The upsweep angle design is

important to generate high lifting flow toward the upper chamber, and improve

looping flow pattern. These will help to improve the overall engine’s scavenging

efficiency. In the simulation work attempted, the upsweep degree of side port, (UPS)

and downsweep degree of exhaust port, (DPE) were set at default values to minimize

the operating parameters. Accordance to the empirical guidance [6], the UPS and

DPE are usually set at 20°. Therefore, only the Upsweep degree of main port, (UPM)

and Rear port, (UPR) factors are analyzed for the port design optimization purpose.

Figure 5.9 and 5.10 shows the upsweep port design in further details.

89

Figure 5.9: The sweep port design.

i. Upsweep Degree of Main port,

(UPM) ii. Upsweep Degree of Rear Port,

(UPS)

Figure 5.10: The design of the upsweep degree of the port. Samples (as shown in Table 5.5) are studied for the optimization of the

upsweep degree for the main port design. The range of the main port upsweep

degree, (UPM) is varies from 10 to 20° while Rear port is varies from 55 to 60°. The

samples are set to vary at every 5° interval and are simulated by the Cosmos FloWork

2004 under the same operating parameters of main port simulation.

Main

Port Rear

Port

90

Table 5.5: The study of the upsweep angle, (°) of the transfer port.

Sample Main port,

UPM, (°)

Rear port,

UPR, (°)

Side port,

UPS, (°)

Exhaust port,

DPE, (°)

1 10 55 20 20

2 10 60 20 20

3 15 55 20 20

4 15 60 20 20

5 20 55 20 20

6 20 60 20 20

5.2.2.1 The Simulation Results

In Figure 5.11 (i), the result of velocity vector distribution for sample 1 has

illustrated the direction of flow toward upper cylinder. This upper flow helps to

scavenge the residual gas which resulted from the combustion process.


flow pattern has quite symmetrical flow pattern condition. Besides, this sample has

resulted with good lifting flow toward the chamber. This upper flow helps to

scavenge the residual gases which resulted from the combustion process. Thus, this

sample has successfully increased the scavenging efficiency.

Figure 5.11 (iii) shows the trajectories the main port and side port flow

towards the rear side of chamber wall to create looping flow. The looping flow

trajectories however, in this design sample, show the low lifting flow at the main

port, that causes the short-circuiting problem.

i. Velocity vector distribution at symmetrical plane ii. The trajectories(top) of the flow pattern iii. Trajectories(side) of flow pattern

Figure 5.11: Simulated results of sample 1.

Good flow toward

upper chamber

Low flow

lifting

Low lifting

flow

91

92

Next, velocity vector distribution for sample 2 shows flow toward upper part

of chamber in Figure 5.12(i). This upper flow helps to scavenge the residual gas

which resulted from the combustion process.

Besides, Figure 5.12 (ii) shows the top trajectories of the flow pattern is at

symmetrical condition. Besides, this sample has resulted with good lifting flow

toward upper part of the chamber to scavenge the residual gases which resulted from

the combustion process. Thus, this sample has successfully increased the scavenging

efficiency.

Figure 5.12 (iii) shows the trajectories at side view of the sample 2. The

trajectories flow of the main port and side port flow towards the rear side of chamber

wall to generate the looping flow. The looping flow trajectories show the better

scavenging process where the fresh charge will loop inside the chamber to scavenge

the residual gas.

In Figure 5.13 (i), the sample 3 has illustrated the direction of flow towards

the upper portion of the cylinder. This upper flow is as similar to simulation in

sample 2.

Besides, Figure 5.13 (ii) shows the top trajectories of the flow pattern is not at

symmetrical condition. The left side flow trajectories show the low lifting that would

cause the short-circuiting problem.

Figure 5.13 (iii) shows the trajectories flow of the main port and side port

flow towards the rear side of chamber wall to generate looping flow. The looping

flow pattern acceptably helps to scavenge the residual gases.

i. Velocity vector distribution at symmetrical plane ii. The trajectories(top) of the flow pattern iii. Trajectories(side) of flow pattern


Good flow toward

upper chamber Good flow

lifting

93

Low lifting

flow

i. Velocity vector distribution at symmetrical plane

ii. The trajectories(top) of the flow pattern iii. Trajectories(side) of flow pattern


Low flow lifting Good flow toward

upper chamber

94

95

Figure 5.14 shows the simulation results of sample 4. In Figure 5.14 (i), the

result of velocity vector distribution illustrates the direction of flow toward upper


previous combustion process.

Besides, Figure 5.14 (ii) shows the top trajectories is also having symmetrical

flow pattern condition. Besides, this sample has resulted with good lifting flow

toward the chamber. This upper flow helps to scavenge the residual gases which

resulted from the combustion process.

Figure 5.14 (iii) shows the trajectories of the main port and side port has flow

towards the rear side of chamber wall to generate looping flow. The looping flow

trajectories have shown in satisfactory condition. The looping flow has resulted in

improved scavenging process where the fresh charge will loop inside the chamber to

scavenge the residual gases.

In Figure 5.15 (i), sample 5 shows good upper flow toward upper part of

chamber. But, Figure 5.15 (ii) shows the top trajectories of the flow pattern did not

have symmetrical flow pattern condition. There is unsatisfied of the left side flow

trajectories show the low lifting condition. This will cause the short-circuiting

problem to happen.

Figure 5.15 (iii) shows the trajectories are in satisfactory condition. The

looping flow has resulted the better scavenging process where the fresh charge will

loop inside the chamber to scavenge the residual gas.


ii. The trajectories(top) of the flow pattern iii. Trajectories(side) of flow pattern


High flow lifting Good flow toward

upper chamber High lifting flow

96


ii. The trajectories (top) of the flow pattern iii. Trajectories (side) of flow pattern


Low Flow lifting Good Flow toward

Upper Chamber Good flow looping

97

98

Figure 5.16 (i), (ii) and (iii) show the computer simulated results of sample 6.

Figure 5.16 (i) depicts the result of velocity vector distribution indicating the

direction of flow towards the upper cylinder. The scavenging will help to flush out

residual gases, which are by-products of the combustion process.

Figure 5.16 (ii) shows the flow pattern of the top trajectories. Here the flow

pattern has been quite symmetrical in nature. Besides, it seems here that there is a

slightly low lift in comparison to sample 4 and 5. The low lift feature will decrease

the scavenging efficiency, where there will be insufficient intake fresh charge to

replace the residual gases.

Figure 5.16 (iii) on the other hand shows the trajectories of the main port and

side port has flow towards the rear side of chamber wall and generated looping flow.

The looping flow has resulted the better scavenging process where the fresh charge

will loop inside the chamber to scavenge the residual gas.


ii. The trajectories (top) of the flow pattern iii. Trajectories (side) of flow pattern


Low flow lift Good Flow toward

Upper Chamber Good flow looping

99

100

5.2.2.2 Conclusion of the Results on Upsweep Angle Design Simulation

From the simulation results, sample 4 demonstrated a near perfect flow lift

and flow toward the upper part of the chamber. The similarity among them is that the

rear port is set at 60° and therefore the upsweep degree for rear port (UPR) is

considered to be optimized at 60°. For other samples, results have demonstrated that

low quality of flow patterns was obtained. Therefore, the optimized design model for

the Schnurle loop scavenging is having the following specifications shown in Table

5.6.

Table 5.6: Specification of the optimized Schnurle loop scavenging design.

No Specifications Detail

1 AM1 65°

2 AM2 50°

3 MT1 20°

4 MT2 12.3°

5 UPM 15°

6 UPR 60°

7 UPS 20°

5.3 Analysis of the Simulated Scavenging Process

An alternative CFD code, Fluent v6.1 was used to predict for the scavenging

efficiency. There were two different transport species used in Fluent v6.1 to

represent the fresh charge and residual burned gas. The dynamic meshing method in

Fluent v6.1 was to simulate the piston movement at any range of engine speed.

Owing to this, the simulation work could perform the analysis on the engine

101

computation domain at the different port timings. Figure 5.17 shows the flow chart

for the simulation works carried out.

Figure 5.17: Flow chart showing the sequence of processes involved when using Fluent v 6.1.

The engine model is drawn in a symmetrical domain, and is transferred to the

mesh tool Gambit v 2.0 in an ACIS file format. The engine domain meshing consists

of quadrilateral and triangular pave typeface mesh, and Het/Hybrid volume mesh

respectively. The definition of the boundary conditions such as the pressure inlet and

exhaust, the symmetrical plane, the piston wall, volume fluid type is made in Gambit

solver.

A computation domain is produced in SolidWorks 2004 and is exported to

Gambit v 2.0 in ACIS format

The mesh option and boundary settings are executed in Gambit v2.0. The file is

exported in mesh file format to Fluent v6.1

The operating parameters are set in the Fluent v6.1, which includes flow properties, model type, viscosity, and the input value of pressure inlet and outlet are made.

The flow simulation results are analyzed in term of flow pattern velocity and pressure contour and mass fraction.

Define the dynamic mesh condition in the use of the unsteady state condition.

The gas sampling method is defined by the transport species without undergoing reaction.

102

Figure 5.18 shows the symmetrical computational domain of the engine model. The

symmetrical domain is used only for the symmetrical chamber design, and this

reduces the iteration time of simulation works.

Figure 5.18: The engine computational symmetrical domain.

During the simulation, the transfer port is assigned as intake and the exhaust

port as exhaust for all the combustion chamber models produced. Similarly, the

transfer port is defined as pressure inlet while the exhaust port is defined as pressure

outlet. The interior of the chamber is designated as the fluid, while the surface of the

chamber is designated as the wall. For the piston surface, it is defined as the moving

wall and this is to simulate the piston movement during the simulation process.

Ambient

Air

Tracer

gas O2

103

Figure 5.19 illustrates the concept of the moving wall of the piston surface,

which could be interfaced with the transfer ports and exhaust port during it motion

from TDC toward to BDC. The scavenging flow starts when the piston wall reaches

the transfer port openings, and the fresh charge flows toward the inner chamber, and

during blow down the residual gases will exit through the exhaust port.

a.) Piston at TDC

b.) Piston moves toward BDC

Figure 5.19: The piston surface (moving wall) of the scavenging process.

For the analysis of the dynamic flow condition, the following assumptions are

appropriately made [4]:

1. No mass or heat is allowed to across the interface between the fresh charge

and burnt gas.

2. The cylinder walls are adiabatic.

3. The two gases involved obey the ideal gas law and have the same molecular

weights, with identical and constant specific heats.

4. The process occurs at a constant cylinder volume and pressure.

The unfired method simulating the scavenging process is applied. In dynamic

gas sampling method, the transport species are defined as non-reacting gas, which

represent the burned and unburned gases. The tracer gas oxygen, O2 is applied as the

unburned gases or fresh charge. The ambient air represents the unburned gas inside

the cylinder. When the transfer ports open during piston descend to BDC, the gas

exchange process is noted to occur. The O2 gas will replace the internal zone of

104

chamber, and will blow down the ambient air through the exhaust port. Table 5.4

shows the parameters set up using Fluent v 6.1.

In the case of this exercise (i.e. scavenging simulation), the engine speed is

set at 8000rpm. The gauge inlet pressure is set at 500kPa as the initial pressure of the

flow intake into the chamber. The initial internal pressure of the chamber, before the

exhaust port opens, is set in accordance to the situation where the expansion process

occurred. The simulation conditions are presented in Table 5.7.

Table 5.7: The set up parameters when using Fluent v6.1.

Specifications Detail

1 Model 3D, segregate, Double precise,

Unsteady, 1st-Order Implicit

2 Viscosity Standard k-epsilon turbulence model

3 Species Transport Non-Reacting (2 species): O2 and Air

Material: mixture, incompressible ideal gas

4 Boundary Conditions

Pressure inlet / outlet:

Turbulence Intensity: 10%

Turbulence Length Scale: 2 mm

5 Meshing

i. Quadrilateral and triangular pave face mesh

ii. Het/Hybrid volume mesh

ii. Dynamic mesh

6. Accuracy check

Convergence when residual reach:

1. at velocity, k- epsilon, continuity = 1 x 10-3

2. at energy = 1 x 10-6

105

Table 5.8: The simulation conditions for the scavenging process analysis.

Parameter Value

1. Engine Speed 8000 rpm

2. Gauge Pressure inlet, (constant) 500 kPa

3. Pressure in-cylinder

before exhaust port open, at 86.6º ATDC (kPa)

600 kPa

There were 38 time steps for the simulation to reach convergence. The overall

duration for the dynamic mesh iteration to converge takes at least 12 hours for each

of the simulation works. After several trials, the dynamic scavenging model which

simulating the whole scavenging the in-cylinder process have been successfully

implemented. The simulation is stopped when the results reaches the residuals for

accuracy as showed in Table 5.7. Figure 5.20 illustrates an example of duration of

the iteration done to reach convergence.

Figure 5.20: An example of a convergence of the dynamic scavenging model simulation work.

106

107

5.3.1 The Simulation Results

The simulation results are elucidated in term of velocity distributions and

mass fraction of species transport. The velocity distribution results have illustrated

the flow pattern during the scavenging process with the mass fraction results have

showed the trapping efficiency.

5.3.1.1. Velocity distribution

Figure 5.21 (a) shows piston moves from 86.6° to 116.6° ATDC. The

expansion volume has created the vacuum inside the chamber. Therefore, the

phenomenon of back flow occurred at the exhaust port. Besides, Figure 5.21 (b)

shows the flow has started entering the chamber.

Figure 5.22 and Figure 5.23 show the scavenging flow went to upper top

chamber. The looping flow pattern is seen to blow down the residual toward the

exhaust port during piston moves from 116.6 to 236.6 ° ATDC.

Figure 5.24 (a) shows the piston continues to move upward to 261.6°ATDC.

The transfer ports are closed, while the compression chamber volume has expedited

the velocity at the zone near to the exhaust port. Figure 5.24 (b) shows both the

transfer port and exhaust port are closed, the internal chamber remains in static

condition.

a. At 111.6° ATDC b. At 136.6° ATDC

Figure 5.21: Velocity contour at 111.6° ATDC and at 136.6° ATDC.

8000rpm

8000rpm

Back Flow Occurred Intake Flow starts

108



8000rpm 8000rpm

Upper Flow Flow filled

chamber

109



8000rpm 8000rpm

110



8000rpm

8000rpm

111

112

5.3.1.2. Species Transport Mass Fraction Distribution

Figure 5.25 (a) shows the event at 111.6° ATDC. There is no any mass

fraction present inside the chamber as the transfer port is closed. As the piston

reached at 136.6° ATDC, the transport species has gradually filled in the transfer

port.

Figure 5.26 (a) illustrate the event at 161.6° ATDC, the transfer ports start to

open, and the initial transport species will enter the chamber. While, Figure 5.26 (b)

shows the piston at 186.6°ATDC, seen to have much more transport species has

entered the chamber.

Figure 5.27(a) (b) shows when the piston at 211.6 and 236.6 ATDC, the

transport species continues flow into the chamber. At the same time, the fresh charge

replaced the residual gas inside the chamber.

Figure 5.28(a) (b) shows that piston at 261.6 and 271.6 ATDC. The intake of

the transport species ceased after the transfer port closed. The chamber was filled

with a quantity of transfer species. Also shown is the compression volume chamber

that continue blow down the residual gas until all the ports were closed.


Figure 5.25: Mass Fraction distribution at 111.6° ATDC and at 136.6° ATDC.

8000rpm 8000rpm

113



8000rpm 8000rpm

114



8000rpm 8000rpm

115



8000rpm 8000rpm

116

117

5.3.2 Discussion on the Results of the Dynamic Simulation

The dynamic simulation work has enabled the mass fraction of the trapped

transport species inside the chamber during the scavenging process be obtained.

Table 5.9 shows the mass fraction results.

Table 5.9: Results of Mass fraction

Crank angle

(ºATDC)

Mass Fraction

Of Transport Species

86.6 0

111.6 0.00024

136.6 0.04456

186.6 0.22452

211.6 0.44534

236.6 0.58822

261.6 0.61538

271.6 0.62946

Figure 5.29 shows the graph of the mass fraction versus the crank angle. The

mass fraction value has increased with the crank angle. This is due to the medium of

the transport species, which increase when the transfer port is opened from 86.6°

ATDC until 266.6°ATDC position.

118

Dynamic Simulation of Scavenging Process

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

86.6 111.6 136.6 186.6 211.6 236.6 261.6 271.6

ATDC

Ma

ss

Fra

cti

on

Figure 5.29 The Mass fraction gas O2 versus crank angle.

From the results shown, the highest trapping efficiency for the proposed

scavenging system was obtained as 63% at 271.6° ATDC. As mentioned in Section

2.5, scavenging efficiency can be calculated by multiplying the trapping efficiency

with the scavenging ratio. Table 5.10 shows the parametric results for the scavenging

system.

Table 5.10: The dynamic results for the scavenging parameter.

Parametric Value

1 Trapping efficiency, 0.63

2 Scavenging ratio

(pump volume ratio) 1.5

3 Scavenging efficiency

(simulation) 0.945

119

In Section 2.6, there are perfect displacements and perfect mixing models for

scavenging system were used to evaluate the simulation results. Table 5.11 shows the

standard data (according to the mathematical model) of perfect displacement and

perfect mixing scavenging model.

Table 5.11: The standard data for the perfect mixing and displacement scavenging

model [4].

Scavenging

efficiency

Trapping

Efficiency

Scavenging

ratio

Perfect

mixing

Perfect

displacement

Perfect

mixing

Perfect

displacement

0 0 1 1 0

0.0952 0.1 0.9516 1 0.1

0.1813 0.2 0.9063 1 0.2

0.2592 0.3 0.8639 1 0.3

0.3297 0.4 0.8242 1 0.4

0.3935 0.5 0.7869 1 0.5

0.4512 0.6 0.7520 1 0.6

0.5034 0.7 0.7192 1 0.7

0.5507 0.8 0.6883 1 0.8

0.5934 0.9 0.6594 1 0.9

0.6321 1 0.6321 1 1

0.6671 1 0.6065 0.9091 1.1

0.6988 1 0.5823 0.8333 1.2

0.7275 1 0.5596 0.7692 1.3

0.7534 1 0.5381 0.7143 1.4

0.7767 1 0.5179 0.6667 1.5

120

Scavenging efficiency Vs Scavenging ratio

0

0.2

0.4

0.6

0.8

1

1.2

0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6

scavenging ratio

sca

ven

gin

g e

ffic

ien

cy

perfect mixing

perfect displacement

dynamic simulation result

In Figure 5.30 and Figure 5.31, the dynamic simulation results are plotted

together with the perfect mixing and perfect displacement standard data. The

dynamic results were closed to the perfect mixing standard, and thus we can

conclude that the simulation results have shown that the scavenging system design is

satisfactory.

Figure 5.30: Scavenging efficiency versus scavenging ratio.

trapping efficiency Vs Scavenging Ratio

0

0.2

0.4

0.6

0.8

1

1.2

0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6

Scavenging Ratio

Trap

pin

g R

ati

o

Perfect mixing

Perfect displacement

Dynamic result

Figure 5.31: Trapping ratio versus scavenging ratio.

CHAPTER 6

FABRICATION OF SCAVENGING SYSTEM TEST RIG

6.1 Introduction

The scavenging test rig was developed to analyze the actual scavenging

process. The unfired condition with the tracer gases of oxygen is applied to analyze

for the scavenging efficiency. In addition, the Scotch-Yoke mechanisms such as the

C-plates, sliders and pistons were constructed to simulate the linear piston motion.

The fabrication work took about six months to finish. During fabrication

process, the motion parts such as i) bearing sliders, ii) piston liners (with the cylinder

block) and iii) crank mechanism have encountered problem of unfit joint. However,

improving the clearance of the machined items solved the problems.

Other constraints were the piston speed. The system speed has to be set low

because most of the rig components such as the cylinder block, liners are made by

Perspec material. To overcome this, a 3-phase motor controlled the system with

speed at 1450 rpm, and the gear box set to 1: 20, which has reduced the speed to 72.5

rpm.

122

6.2 Test Rig Components

The test rig components are first drawn in SolidWorks 2004. The detail

drawings provide the information of the components dimensions, orthographic view

and the isometric view of the assembly as well as the exploded view of the

components. The physical dimensions are illustrated in Appendix B.

During assemblies, the liners, slider bearing, cylinder block and pistons

needed much adjustments and this is to ensure that of the pistons could run smoothly.

Besides this, the gasket seal, oil rings and Teflon are added to mitigate the leakage

problems. The soap bubble inspection method was applied to check for the leakage.

Figure 6.1 shows the gasket sealing and Figure 6.2 shows the soap bubble method

performed for leakage test on the rig.

Figure 6.1: The gasket sealing and leakage inspection.

Gasket

Piston

Crankcase

Piston Liner

123

Figure 6.2: Leakage inspection with soap bubble.

BOM is a product structure to describe what raw materials or components are

required, and in what quantities, to produce the engine model. Table 6.1 shows the

Bill of Material (BOM) for the test rig model.

The Orthographic Drawing and Exploded Assembly Drawing for the test rig

are illustrated in Appendix B1 and B2. There are in total 35 items, which made up the

test rig components. The crank-mechanism components are fabricated using

aluminum, whilst the cylinder block and crankcase are fabricated in Perspec

material. The machined components are illustrated from Figure 6.3 to Figure 6.5.

Bubble

Soap

Intake

Manifold

Cylinder

Block

124

Table 6.1: Bill of Material (BOM) for the test rig model.

Item

no. Qty Part No. Material Drawing No.

Appendix

1 3 Slider R Aluminium 1 B3.1

2 3 Slider L Aluminium 2 B3.2

3 3 Slider Bearing Brass 3 B3.3

4 3 Slider Bearing 2 Brass 4 B3.4

5 2 Bearing Withthrust Brass 5 B3.5

6 2 Crank Bearing M Brass 6 B3.6

7 1 Crankshaft Aluminium 7 B3.7

10 2 Crankcase Aluminum 8 B3.8

11 1 Exhaust Manifold Perspex 9 B3.9

12 4 Crankshaft Bearing Brass 10 B3.10

13 1 Intake manifold Perspec 11 B3.11

14 6 c-platrigmodel1 Aluminium 12 B3.12

15 4 c-platrigmodel2 Aluminium 13 B3.13

16 4 Piston55K Aluminium 14 B3.14

17 2 Compression rig2 Aluminium 15 B3.15

18 1 Piston pump Aluminium 16 B3.16

19 2 Sleeve test Perspec 17 B3.17

20 2 reed main body Perspec 18 B3.19

21 4 reed petal Fiber glass 18 B3.19

22 4 reed limiter Brass 18 B3.19

23 8 screw, M2 x 0.4 x 3

- - -

24 1 Block gasket Paper

gasket - -

25 1 Cylinder head Perspec 19 B3.20

26 1 Block Perspec 20 B3.20

27 1 linertest Perspec - -

28 1 gasket1 Paper

gasket - -

29 1 linearslide Brass 21 B3.21

30 2 sparkplug - - -

31 2 gasket exhaust Paper

gasket - -

32 2 gasket intake Paper

gasket - -

33 2 adapter block Perspec 22 B3.22

34 4 Piston ring Oil ring - -

35 2 Pump ring Oil ring - -

125

Figure 6.3: The machined items of the engine crank mechanism.

Figure 6.4: The Perspec material representing the intake manifold.

C-plate

Piston

Slider

Intake

Manifold

Probe connector

126

Figure 6.5: The associated reed valves and cylinder head section.

The overview of the motored scavenging test rig is shown in Figure 6.6. The

motored assembly consists of an AC Inverter, a 2-phase motor, a coupling, and a

gearbox. The 3-phase motor will convert electric energy to mechanical turning

torque to run the engine model at a predetermined speed.

Figure 6.6: The overview of the motorized scavenging test rig.

AC Inverter

(50HZ)

3-phase Motor 50Hz

Gear Box

(1:20) Coupling

Engine

Model

Reed valve

with gasket

Cylinder

Block

127

Besides this, the main components for the scavenging system test rig (i.e. the

motored system), there are several instrumentations used for the measurement

purposes. These instrumentations and consumable are listed in Table 6.2.

Figure 6.7 illustrates the gas analyzer probe used. The probe was used to

collect the samples of the trapped gas. It is linked to the Oliver IGD gas analyzer to

display the respective constituents of the exhaust gas.

Figure 6.8 illustrated the Dewetron high-speed data acquisition unit and the

crank encoder used. The instrumentations are used to measure the crank angle during

the scavenging process in conjunction with the chamber pressure.

Table 6.2: Specification of the instrumentations.

Description of Instrumentations Detail

1 Cylinder Gas 02 and N2 MOX, 7.2m3, 145bar,

Purity = O2 (95%), N2 (99.99%)

2 Pressure regular 1 & 2 1. Comet 700, BOC, 0 – 10 bar

2. CONCOA, Range 0- 25 bar

3 Inverter TECO, 220V, 1Hp, 0-50 Hz

4 3-phase motor TECO, 1450Rpm, 50Hz,

Power: 0.56kW(0.75hp)

5 Speed Reducer GONG TZYH, TKB50, Ratio 1:20

6 Pressure Transducer KISTLER, type 6117BCD15, SN

127479, Measure range: 0-50bar.

7 Exhaust Analyzer

TOCSIN IGD 300 GA

Response time = 30s

Sample rate = 1L/s

8 Pressure and crank angle Signal Monitor the signal relation between pressure and crank angle

9 Crank angle encoder KISTLER type 2613B

10 Manometer Micro manometer, model 8702,

DP-CALC

128

i. Gas analyzer probe ii. Oliver IGD gas Analyzer

Figure 6.7: The gas analyzer probe and Oliver IGD gas analyzer.

i. Dewetron signal Display

ii. Crank angle sensor

Figure 6.8: The Dewetron high-speed data acquisition and crank shaft encoder.

Figure 6.9 (i) shows a digital manometer, which was used to measure for

pressure and velocity at the intake manifold. Figure 6.9 (ii) shows a Tachometer,

which was used to check for the rotational speed of the system during trials.

129

i. Digital manometer

ii. Tachometer

Figure 6.9: Digital Manometer and Tachometer.

6.3 Test Rig Set-up

In the scavenging system model, only half of the model was developed for

the testing purpose. As such, only a piston pump chamber and two combustion

chambers (chamber A and B) were constructed. The intake manifold is connected to

the piston pump chamber while the piston pump chamber supplied the intake charges

to each side of the combustion chamber at every 180° interval.

The scavenging measurement method includes:

1. Pressure and velocity measurements at intake and pumping manifolds (at

both Chamber A and Chamber B) during the scavenging process.

2. Measurements of the internal pressure of pumping and combustion

chambers during the scavenging process.

3. Measurements of the trapped volume fraction inside the chamber during

the scavenging process.

130

Figure 6.10 illustrates the schematic diagram of the scavenging test rig. There

is inlet from tracer gas Oxygen (O2) as the unburned gas, while another inlet gas

Nitrogen (N2) represented the combustion residual gas. There are two outlets, exhaust

port and a one-way control valve mounted at the top of the chamber. During the end

of the blow down process (after exhaust port closed), the trapped volume inside

chamber will be compressed by the piston, and force the trapped volume to pass the

control valve for the sampling collection.

Figure 6.10: Schematic diagram of the scavenging test rig set up.

The experimental procedures for the measurement of the scavenging efficiency

are as following:

1. Install the instrumentations and check for the leakage.

2. Start to run the motor to intake the charging gas O2.

3. While the transfer port opens, the gas O2 will enter the chamber.

4. While the reach the TDC, exhaust port is closed, and the trapping volume

flow through the one-way valve to the exhaust analyzer.

5. Gas N2 is drawn into the chamber to fill the vacuum in-cylinder.

6. The gas sampling is collected by exhaust analyzer.

Gas O2 cylinder

Exhaust Analyzer

Inverter Speed

Reducer

Pressure Regulator 1

3-phase Motor

Gas Box Manometer

Testing Rig

Engine

Model

Pressure Regulator 2

Gas N2

Gas Box

Inlet

Inlet

Exhaust port

(During

compression)

131

Figure 6.11 shows the picture of the measurement of pressure and velocity

points and cylinder A and B.

Figure 6.12 shows the picture of the analyzer probe measurement. The gas

analyzer probe collected the species for is putting at the opening of the outflow of

this trapped volume to obtain the samplings. Besides, another tracer gas Nitrogen is

applied as the ambient gas in the chamber. Only the gas O2 is analyzer with Oliver

IGD analyzer, which sets the range of reading at 20 – 100 % with tolerance of 0.01%

and with accuracy at Forecast Standard Deviation (FSD) of 1%.

Figure 6.13 showed the piston pumping will draw the tracer gas O2 from gas

box, while the combustion chamber will draw the gas N2 during the expansion.

Figure 6.11: The scavenging measurement arrangement.

Cylinder B

Cylinder A

P1, V1

Point

P2, V2

Point

Figure 6.12: The gas analyzer probe on the outflow of the system.

Analyzer Probe

Supply of gas N2

Check Valve Spark plug as blind plug

Cylinder Block

132

Figure 6.13: The illustration of the scavenging measurement. 133

Volume

A

Volume

B

P2, V2

P1,V1

Tracer gas O2

Tracer gas N2

To gas analyzer

134

6.3.1 Scavenging Measurement Results

The measurement parameters for the scavenging process were in term of the

pressure and velocity at manifold, as well as the mass fraction of tracer species. The

pressure and velocity distribution showed the flow rate, and the mass fraction

showed the scavenging efficiency.

Figure 6.14 showed the pressure inlet, P1 with the engine speed. The pressure

P1 is increased when the engine speed increased. The Volume A with pump and

tracer gas has shown the higher pressure than the volume B. At speed 74.5 rpm, P1

for volume A is 12 mmHg (0.016 bar), volume B is 8 mmHg (0.011 bar) and the

pressure P1 without pump is only 2.26mmHg (0.003 bar). The differences in pressure

inlet of volume A and volume B may due to the geometrical design is different in

between the each side of pumping manifold due to the double action pumping design

constraint.

Pressure Inlet, P1 Versus Engine Speed, Rpm

0.00

2.00

4.00

6.00

8.00

10.00

12.00

14.00

0 10 20 30 40 50 60 70 80

Engine Speed, RPM

Pre

ssu

re (

mm

Hg

)

Without pump

Volume A withpump,and tracer gasO2

volume B withpump,and tracer gasO2

Figure 6.14: Pressure inlet P1 versus Engine speed (rpm).

135

The Figure 6.15 shows the variation of the inlet velocity, V1 with engine

speed. The maximum velocity volume A (with pump) and tracer gas could reach

50.70m/s. The velocity for volume B (with pump) and tracer gas is at about

42.38m/s. The velocity is observed to increase with the engine speed. The condition

of the velocity V1 for the testing without piston pump and the condition has shown

the lower results (20.9m/s). Both the volume A and volume B has received the

pumping charge, this showed that the piston pump design has been successfully

provided the double action of pumping in every 180° interval.

Figure 6.16 shows the pumping manifold pressure versus engine speed. The

pumping pressure for volume A has a small drop of pressure when the engine speed

increases. This may caused by the unsatisfactory lifting of the reed valves. In

addition to this volume B shows the increase of the pressure P2 with the engine

speed. But, the pressure P2 for volume B is noted to be lower than P2 at volume A.

This may due to the discharge coefficient is higher during the increasing of engines

speed, and thus the gas leakage problem through the clearance between the piston

pump and the cylinder liner is substantially reduced.

Figure 6.17 shows the pumping manifold velocity versus engine speed. The

velocity V2 for volume A is slightly higher than volume B. This maybe also

influenced by the differential of the pressure P2, as mentioned in Figure 6.10. The

maximum for V2 could be archived by volume A is 36.55 m/s, while volume B is at

30 m/s.

Figure 6.18 shows the profile of the trapped volume ratio (of gas O2) versus

engine speed. Both the trapped volume for Volume A and B are noted to increase

with the increase in engine speed. This is most likely due to increase in the pumping

work and proportionately reduction in the pressure lost at higher speed region. Here

the maximum trapped volume ratio for volume A is 0.75, while the cylinder B is at

0.70.

Inlet Velocity, V1 versus Engine Speed, RPM

0.00

10.00

20.00

30.00

40.00

50.00

60.00

0 10 20 30 40 50 60 70 80

Engine Speed, RPM

Inle

t V

elo

city

, m

/s

without pump

volume A with pump,and tracer gas O2

volume B with pump,and tracer gas O2

Figure 6.15: Inlet velocity, V1 versus Engine Speed (rpm).

136

Pumping Manifold Pressure , P2 Versus Engine Speed, RPM

0.00

1.00

2.00

3.00

4.00

5.00

6.00

7.00

0 10 20 30 40 50 60 70 80

Engine Speed, RPM

Pres

sure

, m

mH

g

volume A with pumpand tracer gas O2

volume B with pump,and gas tracer

Figure 6.16: Pumping manifold Pressure, P2 versus Engine Speed (rpm). 137

Pumping manifold velocity, V2 versus Engine Speed, RPM

0.00

5.00

10.00

15.00

20.00

25.00

30.00

35.00

40.00

0 10 20 30 40 50 60 70 80

Engine Speed, RPM

Vel

ocit

y,

m/s

volume A withpump and tracergas

volume B withpump and tracergas

Figure 6.17: Pumping manifold velocity, V2 versus Engine Speed (rpm).

138

Trapped volume ratio of gas O2 versus Engine Speed, RPM

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

0 10 20 30 40 50 60 70 80

Engine Speed, RPM

Tra

pp

ed v

olu

me

rati

o o

f g

as

O2 Trapped volume

ratio of gas O2for volume A

Trapped volumeratio of gas O2for volume B

Figure 6.18: Trapped volume ratio of Gas O2 versus Engine Speed (rpm).

139

140

6.4 The Pressure inside the chamber

The in-cylinder pressures of cylinder chamber A and B, and cylinder pump

are measured with the use of a pressure transducer (refer to Table 6.2). Apart from

the pressure measurements, volume displacement is measured using crank angle

encoder. The reference cylinder TDC is set to 0º crank angle in the DeweCA

software. The in-cylinder pressure for volume A and B are found to depict similar

gauge pressure distribution profile. Figure 6.19 shows the schematic diagram of the

pressure in-cylinder measurement.

The pressure measurement signals are acquired from the Dewetron signal

display. The software DeweCA v2.2 is applied for the signal display and data reading.

The parameters set up of DeweCA v2.2 is shown at Appendix E.

Figure 6.19: Schematic diagram of the pressure in-cylinder measurement.

Dewetron high speed data acquisition

Test Rig

Engine Model Pressure

Transducer Crank Angle

encoder

Inverter Speed Reducer

3-phase Motor

Signal 1

Speed input

Signal 2

Data Collection

141

The mounting position of the pressure transducer is centrally at the spark plug

hole. In most occasions, the Kistler type pressure transducer is mainly use for

measuring pressure profile of a combustion cycle in a reciprocating engine. Figure

6.20 shows the location of the mounting of the pressure transducer.

Figure 6.20: The location of the mounting of the Pressure Transducer.

6.4.1 Results of Pressure-In-Cylinder Analysis

The comparison between the cylinder A, B and pump with the engine speed

is illustrated Figure 6.21 to 6.23. The TDC is set to 0° for this simulation results

discussion. The crank angle in between -120° to -80° is the expansion process, while

Pressure Transducer Cylinder

A Cylinder Pump

Cylinder B

Cylinder Block

142

the crank angle in between -80° to 40° is the compression process inside the

chamber.

The pressure inside cylinder A and B has dramatically drop during the period

at timing from crank angle from -120º to -80º, this may because of expansion volume

chamber during the piston moves to BDC. However, the pressure started to increase

during period of crank angle from -80º to 0º. This is because of the compression

stage of the piston movement.

In Figure 6.21, the pressure in-cylinder A is found has same distribution with

the engine speed. The maximum value for the gauge pressure for the compression is

0.8 bar at 74.5rpm. The pressure chamber dropped during the period of crank angle

from -120° to -80°, the lowest negative gauge pressure is 2.6 bar (vacuum). This

vacuum condition occurred due to the expansion volume at unfired condition.

Figure 6.21: Pressure Variation in chamber A versus Crank Angle.

143

144

In Figure 6.22, the pressure in-cylinder B is found also has pressure

distribution with the chamber A. The maximum value for the gauge pressure for the

compression is 0.8 bar at 74.5rpm. The pressure chamber also dropped during the

period of crank angle from -120° to -80°, the lowest negative gauge pressure is 3.1

bar (vacuum). This vacuum condition occurred also due to the expansion volume at

unfired condition.

Figure 6.23 showed the pressure in piston pumping chamber, the negative

gauge is found lower than ambient pressure. This is because of the piston pump

always expanded its volume for the induction process. The expansion volume caused

the vacuum and drew the new intake charge. Another reason for the negative gauge

pressure is that reed valve always allowed the medium to flow through to pumping

manifold during the piston moves upward TDC to compress the medium. However,

to understand the actual pumping process, the pumping manifold was instigated. The

Figure 6.16 has showed that highest velocity of pumping manifold is 36.55 m/s,

while the velocities for intake charge without pump is 20.9 m/s. There is a relatively

increase of 15.65 m/s with the piston pump usage. This has proven that the negative

gauge pressure inside the pumping chamber was the process of induction of new

charge. The highest of negative gauge pressure for the pumping chamber could reach

at 0.245 bar (vacuum).

Figure 6.22: Pressure Variation in chamber B versus crank angle.

145

Figure 6.23: Pressure Variation in piston pump chamber versus Crank angle.

146

147

6.5 Scavenging Performance Analysis

Due to the test rig limitation at the running speed, the experimental results

could only provide the result of the scavenging performance at the 72.5RPM speed

range. The experimental data for the testing is attached in Appendix D. Table 6.3

shows the experimental results for volume A and volume B. The volume A obtained

the scavenging efficiency at 0.75, while the volume B obtained the scavenging

efficiency at 0.74. To compare the typical value in section 2.5(Table 2.4), the

scavenging efficiency is in between 0.6 to 0.9. Therefore, the scavenging efficiency

result is within the range of typical value.

Table 6.3: The experimental results for volume A and volume B.

Scavenging ratio

(pump volume ratio)

Scavenging

efficiency

( experimental)

Trapping

efficiency

Volume A 1.5 0.75 0.50

Volume B 1.5 0.74 0.49

To understand the effectiveness of the trapped volume value in the test rig,

the comparison with the ideal scavenging model is required. Figure 6.24 and Figure

6.25 show that the situation of the scavenging parametric for this engine model is

close to the perfect mixing model curve line. This showed the engine model, which

employs the gas sampling method has met the good mixing process during the

scavenging process.

148

Scavenging efficiency vs Scavenging ratio

0

0.2

0.4

0.6

0.8

1

1.2

0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6

scavenging ratio

sca

ven

gin

g e

ffic

ien

cy

perfect mixing


Volume A

volume B

Trapping efficiency vs Scavenging ratio

0

0.2

0.4

0.6

0.8

1

1.2

0 0.2 0.4 0.6 0.8 1 1.2 1.4 1.6

scavenging ratio

tra

pp

ing

eff

icie

ncy

Perfect mixing


volume A

volume B

Figure 6.24: Scavenging efficiency versus scavenging ratio

Figure 6.25: Trapping ratio versus scavenging ratio.

CHAPTER 7

CONCLUSIONS AND RECOMMENDATIONS

FOR FURTHER WORK

7.1 Conclusions

The following salient points are the major outcomes of the work carried out:

1. The scavenging system with external piston pump has successfully proposed

and developed. The external piston pump is designed by adapting the Scotch-

Yoke mechanism, which allows creating the double action of the pumping

process. The piston pump is capable of a delivery ratio reaching at 1.5. The

double action-pumping feature has been proven to be successfully developed

for the engine model.

2. In static condition, an optimized Schnurle loop scavenging system was

proven to have a better scavenging flow pattern inside the cylinder. With the

optimization of the port geometry design, it manages to achieve the goal of

low fuel consumption and reduction of exhaust emission.

150

3. The dynamic model test of the engine domain was successfully investigated

with the CFD code Fluent v 6.1. The results have shown promising results of

the inlet charge flowing toward the upper chamber during the blow down

process. In addition, the gas sampling method has shown that the simulation

result for the engine could reach the scavenging efficiency at 0.945 and

trapping efficiency at 0.63 at the maximum engine speed of 8000 rpm.

4. In the test rig, the scavenging performance experimental results have shown

the scavenging results are near to the perfect mixing condition. The

scavenging efficiency could achieve 75% and trapping efficiency at 50%.

This will allow for the reduction of the pollutant emission, and will minimize

short-circuiting problem of the two-stroke engine.

This project is of significant important to produce a scavenging system for a

newly designed multi-cylinder two-stroke Scotch-Yoke engine, which will contribute

to the prototype development of a lean burn, stratified-charge, and two-stroke engine.

The reduction of the short-circuiting problem (for this engine model) will ultimately

reduce fuel lost and mitigate pollutant emission caused by the incomplete

combustion process normally associated with conventional two-stroke engine.

151

7.2 Recommendation for Further Works

Some of the immediate work to be carried put to further enhance the performance of

the engine prototype will be:

1. The engine model has been simulated with the unfired condition, and the

port geometry design is considered in optimized design condition.

Therefore, for the fired condition purposes, the consideration of the piston

and the engine chamber shapes are to be made to adapt the latest

technology of direct fuel injection system and ignition timing system.

2. Further simulation work on the scavenging process must include in-

cylinder combustion process at various operating conditions. The

investigation of the combustion and scavenging processes will be the next

challenge to further validate the virtual prediction of the engine

combustion analysis made earlier on.

3. Further design optimization of the engine intake and exhaust systems is

required especially the tuning of the engine exhaust for optimize engine

output.

4. Further design optimization of the reed valve and its materials used is also

required to ensure that the valves are able to withstand the extreme

pressure and temperature fluctuation in engine’s transient and steady-state

operating condition.

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APPENDICES

157

Ap

pen

dix

A

158

Ap

pen

dix

B1

159

Ap

pen

dix

B2

160

Ap

pen

dix

B3.1

161

Ap

pen

dix

B3.2

162

Ap

pen

dix

B3.3

163

Ap

pen

dix

B3.4

164

Ap

pen

dix

B3.5

165

Ap

pen

dix

B3.6

166

Ap

pen

dix

B3.7

167

Ap

pen

dix

B3.8

168

Ap

pen

dix

B3.9

169

Ap

pen

dix

B3.1

0

170

Ap

pen

dix

B3.1

1

171

Ap

pen

dix

B3.1

2

172

Ap

pen

dix

B3.1

3

173

Ap

pen

dix

B3.1

4

174

Ap

pen

dix

B3.1

5

175

Ap

pen

dix

B3.1

6

176

Ap

pen

dix

B3.1

7

177

Ap

pen

dix

B3.1

8

178

Ap

pen

dix

B3.1

9

179

Ap

pen

dix

B3.2

0

180

Ap

pen

dix

B3.2

1

181

Ap

pen

dix

B3.2

2

182

APPENDIX C

Dynamic Mesh Option Setting

The Scotch-Yoke mechanism piston movement step is needed to adapt the

conventional engine setting in the Fluent v6.1 mesh option set up.

The calculation for the Dynamic mesh for crank angle:

( )

°=

=

=

+−=

+−=+−

+−=−−

+−=−

−=−

−=−

−+−+=

−+−+=

63.86

05885.0

4728

25.278cos

5761000025.9702cos4728

cos576cos472825.9702cos57657610000

cos576cos472825.9702)cos1(57610000

cos576cos472825.9702sin57610000

cos245.98sin24100

cos245.98sin24100

`)sin24100cos24(241005.25

)sincos(

22

22

22

2222

222

222

222

θ

θ

θ

θθθ

θθθ

θθθ

θθ

θθ

θθ

θθ araarD

Where,

D = 25.5mm,

a = crank offset, 24 mm

r = connecting rod length, 100 mm

θ = crank angle, which is measured from the cylinder centerline and is zero when the

piston is at TDC.

183

The dynamic mesh In-cylinder setting is as following:

Appendix D

Scavenging Rig Experimental Data

Date: 1/6/2005

Testing 1: Measurement at cylinder chamber without pump,

and at natural aspirated condition

Inverter (Hz)

Tachometer (rpm) P1 V1

1 2 3 4 min 1 2 3 4 min

10 14.8 0.83 1.03 0.95 0.91 0.93 11.47 13.14 10.84 13.73 12.30

20 29.7 2.05 2.49 2.45 1.93 2.23 17.92 15.12 16.03 24.61 18.42

30 44.6 2.13 2.13 2.3 2.46 2.26 19.33 20.02 21.82 20.58 20.44

40 60.5 2.15 2.11 2.07 1.98 2.08 21.26 20.05 20.49 21.84 20.91

50 74.5 2.1 2.16 2.09 2.68 2.26 22.03 20.53 21.41 19.6 20.89

1

84

Date: 9/6/2005

Testing 2: Measurement at cylinder A with piston pump,

and supply with gas sampling O2

i.) The measurement of the Pressure and velocity

Inverter (Hz)


1 2 3 4 min 1 2 3 4 min

10 14.8 11.91 9.16 10.42 12.01 10.88 43.86 50.56 49.23 44.32 46.99

20 29.7 12.65 11.79 11.66 10.6 11.68 49.23 51.87 47.48 47.49 49.02

30 44.6 11.66 12.22 12.32 12.63 12.21 49.64 49.61 51.84 44.01 48.78

40 60.5 11.77 11.64 12.9 12.68 12.25 49.23 50.13 49.94 49.94 49.81

50 74.5 12.41 11.2 12.42 13.08 12.28 51.4 51.32 52.54 47.53 50.70

Inverter (Hz)


1 2 3 4 min 1 2 3 4 min

10 14.8 6.56 7.19 6.09 7.48 6.83 21.06 34.97 36.36 34.62 31.75

20 29.7 6.2 7.11 6.84 7.49 6.91 24.38 35.6 37.81 22.4 30.05

30 44.6 4.87 6.44 5.94 4.76 5.50 33.59 32.07 36.22 36.77 34.66

40 60.5 5.41 4.99 4.32 5.07 4.95 32.48 32.21 31.07 33.01 32.19

50 74.5 5.4 5.48 5.25 5.41 5.39 33.98 40.85 34.92 36.46 36.55

185

ii.) The measurement of the Scavenging Efficiency

Inverter (Hz)

Tachometer (rpm)

Gas box sampling

O2, %vol Analyzer sampling

O2, % vol Scavenging Efficiency

10 14.8 56.8 31.59 0.539078498

20 29.7 56.8 36.88 0.629351536

30 44.6 56.8 40.85 0.697098976

40 60.5 56.8 41.68 0.711262799

50 74.5 56.8 44.39 0.757508532

Date: 13/6/2005

Testing 3: Measurement at cylinder B with piston pump,

and supply of gas sampling O2

i.) The measurement of the Pressure and velocity

Inverter (Hz)


1 2 3 4 min 1 2 3 4 min

10 14.8 8.45 8.31 9.31 6.65 8.18 36.7 39.94 36.78 40.1 38.38

20 29.7 8.78 9.21 8.97 8.84 8.95 40.36 40.81 40.23 40.58 40.50

30 44.6 7.46 6.8 6.31 6.86 6.86 42.24 42.23 42.09 42.06 42.16

40 60.5 7.55 8.01 7.56 7.51 7.66 41.65 42.24 42.13 42.67 42.17

50 74.5 7.19 8.58 7.27 8.84 7.97 41.73 41.51 41.81 44.47 42.38

186

Inverter (Hz)


1 2 3 4 min 1 2 3 4 min

10 14.8 5.31 2.83 3.34 3.25 3.68 23.46 20.46 28.87 27.66 25.11

20 29.7 3.47 3.15 3.3 3.65 3.39 24.38 24.01 28.4 29.46 26.56

30 44.6 4.04 4.04 3.08 3.68 3.71 26.9 28.15 29.36 30.22 28.66

40 60.5 4.64 4.48 4.55 4.28 4.49 27.98 26.04 27.07 27.64 27.18

50 74.5 4.1 4.08 4.58 4.58 4.34 30.36 30.32 28.79 30.42 29.97

ii.) The measurement of the Scavenging Efficiency

Inverter (Hz)

Tachometer (rpm)

Gas box sampling O2, %v

Analyzer sampling O2, % v Scavenging Efficiency

10 14.8 56.8 28.06 0.47883959

20 29.7 56.8 36.45 0.622013652

30 44.6 56.8 39.21 0.669112628

40 60.5 56.8 36.45 0.622013652

50 74.5 56.8 43.53 0.742832765

187

188

Appendix E

Dewetron Signal Display Setting

i. The setting of the engine geometry

ii. The channel for the pressure transducer detection

189

Appendix F

Pressure In-cylinder Data

Measurement of the Pressure in-cylinder A at different rpm

Crank degree(°)

20Hz ( 14.8 rpm)

30Hz ( 29.8 rpm )

40Hz (44.6 rpm )

50Hz (74.5rpm)

-180 -0.099487 -0.00044 0.079346 0.115967

-175 -0.100708 -0.0061 0.075073 0.113678

-170 -0.10376 -0.00828 0.07019 0.106812

-165 -0.106201 -0.01221 0.065918 0.106812

-160 -0.109863 -0.01482 0.064087 0.102234

-155 -0.111694 -0.02093 0.057373 0.092316

-150 -0.12207 -0.03749 0.030518 0.069428

-145.001 -0.134277 -0.05668 0.004272 0.031281

-140.001 -0.140991 -0.0715 -0.01099 0.010681

-135.001 -0.147705 -0.08458 -0.03906 -0.04959

-130.001 -0.158691 -0.09766 -0.0592 -0.01907

-125.001 -0.222168 -0.17003 -0.14282 -0.1297

-120.001 -0.36499 -0.32131 -0.30701 -0.29678

-115.001 -0.510254 -0.49221 -0.49561 -0.49667

-110.001 -0.679321 -0.69013 -0.71045 -0.72479

-105.001 -0.847778 -0.90114 -0.95337 -0.9819

-100.001 -1.02478 -1.13395 -1.22131 -1.26953

-95.0013 -1.19446 -1.37373 -1.51123 -1.58386

-90.0014 -1.34766 -1.60697 -1.80359 -1.90887

-85.0014 -1.46545 -1.82408 -2.08557 -2.22168

-80.0015 -1.52283 -1.97449 -2.29431 -2.4498

-75.0016 -1.51001 -2.04119 -2.39075 -2.55051

-70.0017 -1.41663 -2.005 -2.35962 -2.49252

-65.0018 -1.24695 -1.85983 -2.21252 -2.34756

-60.0018 -1.00525 -1.61264 -1.95618 -2.08817

-55.0019 -0.730591 -1.30048 -1.63574 -1.76926

-50.002 -0.435181 -0.92991 -1.23657 -1.37711

-45.0021 -0.162964 -0.57068 -0.82947 -0.96054

-40.0021 0.0567627 -0.25504 -0.46692 -0.57068

-35.0022 0.195313 0.009591 -0.15625 -0.24109

-30.0023 0.286255 0.20752 0.092163 0.028229

-25.0024 0.394287 0.322178 0.281372 0.234985

-20.0024 0.472412 0.430298 0.392456 0.375366

-15.0025 0.524292 0.525338 0.497437 0.46463

-10.0026 0.563354 0.595093 0.59021 0.579071

-5.00267 0.592651 0.650024 0.662842 0.654602

-0.00274658 0.608521 0.688825 0.719604 0.719452

5.19717 0.623169 0.717163 0.761108 0.772095

10.1971 0.567017 0.701468 0.759888 0.785065

15.197 0.369263 0.569807 0.673828 0.722504

20.1969 0.230103 0.457328 0.599976 0.668335

190

25.1969 0.120239 0.366647 0.523682 0.610352

30.1968 0.0476074 0.25504 0.432129 0.526428

35.1967 0.0170898 0.164795 0.291748 0.392151

40.1966 0.0109863 0.142997 0.222778 0.257111

45.1966 0.0048828 0.134713 0.219116 0.25177

50.1965 0.0024414 0.130354 0.210571 0.244904

55.1964 -0.003662 0.121634 0.206299 0.241089

60.1963 -0.008545 0.115967 0.199585 0.2388

65.1963 -0.012207 0.112043 0.193481 0.230408

70.1962 -0.020142 0.102888 0.187378 0.226593

75.1961 -0.025024 0.100272 0.180054 0.219727

80.196 -0.029297 0.092861 0.177002 0.217438

85.196 -0.033569 0.088065 0.169678 0.211334

90.1959 -0.036621 0.082833 0.167236 0.20752

95.1958 -0.040283 0.078474 0.158691 0.200653

100.196 -0.046387 0.071498 0.151978 0.197601

105.196 -0.053711 0.064523 0.149536 0.192261

110.196 -0.053101 0.062343 0.144653 0.185394

115.195 -0.057983 0.054496 0.139771 0.180817

120.195 -0.061646 0.052752 0.135498 0.177002

125.195 -0.065918 0.047084 0.128174 0.175476

130.195 -0.068359 0.040109 0.125122 0.16861

135.195 -0.075684 0.036621 0.118408 0.167084

140.195 -0.076904 0.032261 0.112915 0.160217

145.195 -0.079956 0.029646 0.106812 0.156403

150.195 -0.083008 0.021798 0.10498 0.152588

155.195 -0.084839 0.02049 0.100708 0.145721

160.195 -0.087891 0.014823 0.096436 0.144958

165.195 -0.091553 0.012643 0.091553 0.138855

170.195 -0.093994 0.005668 0.087891 0.13504

175.195 -0.095825 0.002616 0.083008 0.131989

179.795 -0.098267 0 0.078125 0.126648

Measurement of the Pressure in-cylinder B at different rpm

crank degree(°)

20Hz ( 14.8 rpm)

30Hz ( 29.8 rpm )

40Hz (44.6 rpm )

50Hz (74.5rpm)

-180 -0.0274658 0.067139 0.15564 0.462341

-175 -0.0350952 0.060018 0.151571 0.457764

-170 -0.038147 0.052389 0.143433 0.454712

-165 -0.0427246 0.044759 0.136312 0.45166

-160 -0.0518799 0.039673 0.126139 0.448608

-155 -0.0579834 0.02594 0.11495 0.431824

-150 -0.0854492 -0.00966 0.0671387 0.38147

-145.001 -0.109863 -0.05188 0.0142415 0.32959

-140.001 -0.126648 -0.08494 -0.0325521 0.271606

-135.001 -0.137329 -0.10173 -0.0620524 0.224304

191

-130.001 -0.143433 -0.12004 -0.0895182 0.18158

-125.001 -0.152588 -0.13682 -0.112915 0.140381

-120.001 -0.268555 -0.26499 -0.24821 -0.00305176

-115.001 -0.415039 -0.43538 -0.431315 -0.204468

-110.001 -0.585938 -0.62561 -0.638835 -0.430298

-105.001 -0.778198 -0.84127 -0.879924 -0.70343

-100.001 -0.98114 -1.08795 -1.1556 -1.01624

-95.0013 -1.20087 -1.35142 -1.45264 -1.35345

-90.0014 -1.42975 -1.6276 -1.77104 -1.73645

-84.0015 -1.69373 -1.9104 -2.09757 -2.13623

-80.0015 -1.83563 -2.16064 -2.3936 -2.52075

-75.0016 -1.96838 -2.34833 -2.63774 -2.85645

-70.0017 -2.034 -2.44904 -2.771 -3.07617

-65.0018 -1.98669 -2.43022 -2.77608 -3.13721

-60.0018 -1.80969 -2.28068 -2.62044 -3.00446

-55.0019 -1.5625 -2.0284 -2.35291 -2.75269

-50.002 -1.24512 -1.67898 -1.99483 -2.40173

-45.0021 -0.91095 -1.30361 -1.58183 -1.98364

-40.0021 -0.578308 -0.91705 -1.17188 -1.53198

-35.0022 -0.289917 -0.56966 -0.782267 -1.10779

-30.0023 -0.0411987 -0.26805 -0.442505 -0.720215

-25.0024 0.161743 -0.01831 -0.154622 -0.384521

-20.0024 0.315857 0.183614 0.0762939 -0.10376

-15.0025 0.437927 0.341288 0.262451 0.128174

-10.0026 0.50354 0.463867 0.411987 0.312805

-5.00267 0.558472 0.545756 0.523885 0.463867

-0.00274658 0.621033 0.606283 0.586955 0.585938

4.99718 0.660706 0.661723 0.656128 0.662231

9.9971 0.576782 0.632731 0.655111 0.680542

14.997 0.485229 0.608826 0.665283 0.727844

19.9969 0.411987 0.564067 0.663249 0.765991

24.9969 0.350952 0.472514 0.579834 0.793457

29.9968 0.306702 0.396729 0.486247 0.799561

34.9967 0.273132 0.361633 0.430298 0.701904

39.7966 0.259399 0.344849 0.416056 0.58136

44.9966 0.244141 0.33315 0.404867 0.572205

49.9965 0.22583 0.319417 0.391642 0.570679

54.9964 0.215149 0.306193 0.380452 0.564575

59.9963 0.201416 0.293986 0.37028 0.566101

64.9963 0.184631 0.26652 0.352987 0.556946

69.9962 0.167847 0.252787 0.343831 0.550842

74.9961 0.15564 0.24058 0.334676 0.547791

79.996 0.143433 0.22939 0.320435 0.541687

84.996 0.131226 0.218201 0.310262 0.537109

89.9959 0.119019 0.205485 0.30009 0.534058

94.9958 0.106812 0.19633 0.287882 0.526428

99.9957 0.0961304 0.18514 0.275675 0.520325

104.996 0.088501 0.177511 0.272624 0.515747

192

109.996 0.0778198 0.165304 0.263468 0.512695

114.995 0.0701904 0.158183 0.252279 0.508118

119.995 0.0595093 0.150553 0.244141 0.498962

124.995 0.0549316 0.142924 0.240072 0.50354

129.995 0.0442505 0.13682 0.231934 0.500488

134.995 0.0320435 0.126648 0.221761 0.495911

139.995 0.0289917 0.120544 0.218709 0.489807

144.995 0.0213623 0.113424 0.209554 0.485229

149.995 0.0183105 0.102743 0.203451 0.483704

154.995 0.00610352 0.09257 0.194295 0.4776

159.995 0 0.088501 0.184123 0.473022

164.995 -0.00915527 0.082398 0.168864 0.469971

169.995 -0.0167847 0.072225 0.172933 0.466919

174.995 -0.0213623 0.065613 0.16276 0.462341

179.795 -0.0274658 0.158691

Measurement of the Pressure in-cylinder C (Piston Pump) at different rpm

crank degree(°) 20Hz ( 14.8 rpm)

30Hz ( 29.8 rpm )

40Hz (44.6 rpm )

50Hz (74.5rpm)

-180 -0.23651 -0.24262 -0.24719 -0.24282

-175 -0.23689 -0.24223 -0.24894 -0.24206

-170 -0.23651 -0.24262 -0.24588 -0.24348

-165 -0.23727 -0.24414 -0.2533 -0.24419

-160 -0.23651 -0.24376 -0.25068 -0.24495

-155 -0.23575 -0.2449 -0.25112 -0.2448

-150 -0.23575 -0.24223 -0.24806 -0.24536

-145.001 -0.23613 -0.24109 -0.24806 -0.24323

-140.001 -0.23422 -0.23804 -0.24283 -0.23972

-135.001 -0.23384 -0.24147 -0.24283 -0.23773

-130.001 -0.2327 -0.23651 -0.24022 -0.23183

-125.001 -0.23346 -0.23689 -0.24022 -0.23321

-120.001 -0.23041 -0.23613 -0.23891 -0.22985

-115.001 -0.23346 -0.23422 -0.23629 -0.227

-110.001 -0.23003 -0.2346 -0.23717 -0.22608

-105.001 -0.23041 -0.23384 -0.23499 -0.22629

-100.001 -0.22926 -0.23308 -0.23542 -0.22502

-95.0013 -0.22888 -0.23232 -0.23629 -0.2238

-90.0014 -0.22926 -0.23155 -0.23324 -0.22603

-85.0014 -0.2285 -0.2327 -0.23411 -0.22471

-80.0015 -0.22774 -0.22965 -0.23368 -0.2243

-75.0016 -0.22926 -0.23117 -0.23281 -0.22288

-70.0017 -0.22965 -0.22965 -0.23237 -0.22161

-65.0018 -0.22736 -0.23079 -0.23237 -0.22247

-60.0018 -0.22812 -0.22888 -0.22888 -0.21983

-55.0019 -0.22812 -0.22812 -0.22932 -0.21769

-50.002 -0.22583 -0.22697 -0.22975 -0.21907

193

-45.0021 -0.22697 -0.2285 -0.23193 -0.21815

-40.0021 -0.22621 -0.22736 -0.22714 -0.21744

-35.0022 -0.22392 -0.22736 -0.22714 -0.21525

-30.0023 -0.22621 -0.22888 -0.2267 -0.21652

-25.0024 -0.2243 -0.22736 -0.22888 -0.21673

-20.0024 -0.22736 -0.22659 -0.22627 -0.21805

-15.0025 -0.22697 -0.22583 -0.22757 -0.21566

-10.0026 -0.22583 -0.22659 -0.2267 -0.21566

-5.00267 -0.22697 -0.22583 -0.22627 -0.21515

-0.00275 -0.22621 -0.22697 -0.22888 -0.2179

4.99718 -0.22659 -0.22812 -0.22801 -0.21495

9.9971 -0.22659 -0.22697 -0.22932 -0.21759

14.997 -0.22812 -0.22697 -0.22975 -0.21256

19.9969 -0.22736 -0.22888 -0.2267 -0.2154

24.9969 -0.22774 -0.22697 -0.22757 -0.21545

29.9968 -0.22545 -0.22812 -0.22757 -0.21566

34.9967 -0.22888 -0.22545 -0.22714 -0.21723

39.9966 -0.22736 -0.23117 -0.2267 -0.21403

44.9966 -0.22621 -0.22965 -0.2267 -0.21586

49.9965 -0.2285 -0.22926 -0.22627 -0.22156

54.9964 -0.22888 -0.22736 -0.22845 -0.22191

59.9963 -0.22888 -0.22697 -0.22888 -0.21729

64.9963 -0.22888 -0.23003 -0.22975 -0.21846

69.9962 -0.22888 -0.22965 -0.22888 -0.21952

74.9961 -0.23003 -0.22926 -0.23019 -0.21947

79.996 -0.2285 -0.23079 -0.23106 -0.21973

84.996 -0.23117 -0.2285 -0.2315 -0.22054

89.9959 -0.23041 -0.23346 -0.2315 -0.22024

94.9958 -0.22926 -0.23155 -0.23063 -0.22003

99.9957 -0.23041 -0.2327 -0.23193 -0.22146

104.996 -0.22965 -0.23155 -0.23411 -0.2211

109.996 -0.23079 -0.23346 -0.23281 -0.22273

114.995 -0.23117 -0.23422 -0.23455 -0.2242

119.995 -0.23117 -0.23499 -0.23411 -0.22476

124.995 -0.23232 -0.23422 -0.23542 -0.2268

129.995 -0.23155 -0.23537 -0.23629 -0.22771

134.995 -0.23193 -0.23575 -0.23673 -0.23377

139.995 -0.23346 -0.2388 -0.23847 -0.23402

144.995 -0.23537 -0.23804 -0.24153 -0.23438

149.995 -0.2346 -0.24071 -0.24109 -0.23275

154.995 -0.23613 -0.23804 -0.24414 -0.23697

159.995 -0.23613 -0.23994 -0.24632 -0.23977

164.995 -0.23689 -0.24185 -0.24632 -0.239

169.995 -0.23575 -0.24376 -0.24719 -0.23992

174.995 -0.2346 -0.24223 -0.24763 -0.24216

179.795 -0.23575 -0.24262 -0.24763 -0.242

194

Appendix G

Tube Adaptor Specifications

NO Tube Accessory specifications quantity purchase

1 tube(inlet) Stainless steel OD = 6 mm, thickness = 1mm

length = 1.5 m

-

2 Male connector Brass Male Connector, 6 mm OD - 1/4 in. Male ISO Tapered Threads

3 units

Swagelok

B-6M0-1-4RT

3 Female connector Brass Female Connector, 6 mm OD - 1/4 in. Female ISO Tapered

1 unit

Swagelok

B-6M0-7-4RT

4 Branch Tee connector

Brass Street Tee, 1/4 in. FNPT - 1/4 in. MNPT - 1/4 in. FNPT

1 unit

Swagelok

B-4-ST

5 Check valve Brass 1-Piece Check Valve, 1/4 in. Male NPT, 1 PSIG Spring

2 unit Swagelok

B-4CP2-1

engine seal fluent

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