Effects of Intake Valve Timing and Injection Timing

EFFECTS OF INTAKE VALVE TIMING AND INJECTION TIMING

IN A NATURAL GAS DEDICATED DIESEL ENGINE

MR.CHEDTHAWUT POOMPIPATPONG

A THESIS SUBMITED IN PARTIAL FULFILLMENT OF THE REQUIREMENTS

FOR THE MASTER OF SCIENCE IN AUTOMOTIVE ENGINEERING

THE SIRINDHORN INTERNATIONAL THAI-GERMAN

GRADUATE SCHOOL OF ENGINEERING

KING MONGKUT'S INSTITUTE OF TECHNOLOGY NORTH BANGKOK

ACADEMIC YEAR 2007

COPYRIGHT OF KING MONGKUT'S INSTITUTE OF TECHNOLOGY NORTH BANGKOK

Name : Mr.Chedthawut Poompipatpong

Thesis Title : Effects of Intake Valve Timing and Injection Timing in a Natural

Gas Dedicated Diesel Engine

Major Field : Automotive Engineering

King Mongkut’s Institute of Technology North Bangkok

Thesis Advisors : Professor Dr.Choi Gyeung Ho

Assistant Professor Dr.Saiprasit Koetniyom

Academic Year : 2007

Abstract

The objective of the research was to study the effects of intake valve timings

(Miller cycle), injection timings and ignition timings on the efficiencies and emissions

in a natural gas dedicated diesel engine. The engine was dedicated to natural gas

usage by modifying piston, fuel system and ignition system. The engine was installed

on a dynamometer and attached with various sensors and controllers. Intake valve

timing, engine speed, load, injection timing and ignition timing are main parameters.

The results of engine performances and emissions are present in form of graphs.

Miller Cycle without supercharging can increase brake thermal efficiency

1.08% and reduce brake specific fuel consumption 4.58%. The injection timing must

be synchronous with valve timing, speed and load to control the performances,

emissions and knock margin. Throughout these tested speeds, camshaft no.1 is

recommended to obtain high volumetric efficiency. Retard ignition timing can reduce

NOx emissions up to 14.5% while maintaining high brake thermal efficiency.

(Total 107 pages)

Keywords : Miller Cycle, Intake Valve Timing, Injection Timing, Ignition Timing,

Emissions and Natural Gas Dedicated Diesel Engine

______________________________________________________________ Advisor

ii

ช่ือ : นายเชษฐวุฒิ ภูมิพิพฒันพงศ ช่ือวิทยานพินธ : ผลกระทบของจังหวะวาลวไอดแีละจังหวะการฉดีเชื้อเพลิงตอเครื่องยนต

ดีเซลที่ดัดแปลงเพื่อใชแกสธรรมชาติ สาขาวิชา : วิศวกรรมยานยนต สถาบันเทคโนโลยีพระจอมเกลาพระนครเหนือ ที่ปรึกษาวิทยานิพนธ : ศาสตราจารย ดร.เชว กยอง โฮ ผูชวยศาสตราจารย ดร.สายประสิทธิ์ เกิดนยิม ปการศึกษา : 2550

บทคัดยอ งานวิจยันี้มจีุดประสงคเพื่อศึกษาผลกระทบของจังหวะวาลวไอดีและจงัหวะการฉีดเชื้อเพลิงตอประสิทธิภาพและมลภาวะจากเครื่องยนตแกสธรรมชาติที่ไดรับการพัฒนามาจากเครื่องยนตดีเซล เครื่องยนตนี้ไดรับการเปลี่ยนแปลงที่ลูกสูบ ระบบการจายเชื้อเพลิง และระบบการจุดระเบดิ เครื่องยนตที่ใชทดสอบนั้นถูกติดตั้งบนไดนาโมมิตเตอรและไดเชื่อมตอเขากับเซ็นเซอรและเครื่องควบคุมหลายชนิด โดยที่จงัหวะวาลวไอดี ความเร็วรอบ โหลด จังหวะการฉีดเชื้อเพลิง และจังหวะการจดุระเบิดเปนปจจัยหลักในการทดสอบ โดยผลการทดสอบดานสมรรถนะและมลภาวะถูกนําเสนอในรูปแบบของกราฟ วัฏจักรมิลเลอรแบบไมมีซูเปอรชารจนั้นสามารถเพิ่มประสิทธิภาพเชงิความรอนไดรอยละ 1.08 และลดการสิ้นเปลืองเชื้อเพลิงจําเพาะไดรอยละ 4.58 ผลการทดลองแสดงใหเหน็ไดวาจังหวะการฉีดเชื้อเพลิงตองสัมพันธกับจังหวะวาลว ความเร็วรอบ และโหลดเพื่อใหไดมาซึ่งสมรรถนะ มลภาวะและการหลีกเลี่ยงการน็อกที่ดี ในชวงความเรว็รอบที่ใชในงานวจิัยนี้พบวาจังหวะวาลวไอดีของเพลาลูกเบี้ยวที่ 1 นั้นเหมาะสมที่สุดในทรรศนะของประสิทธิภาพเชิงปริมาตร การศึกษายังพบดวยวาการจุดระเบิดลานั้น สามารถลดสวนประกอบออกไซดของไนโตรเจนไดถึงรอยละ 14.5 โดยยังคงไวซ่ึงประสิทธิภาพเชิงความรอนที่สูง

(วิทยานิพนธมีจํานวนทั้งส้ิน 107 หนา) คําสําคัญ : วฏัจักรมิลเลอร จังหวะวาลวไอดี จังหวะการฉีดเชื้อเพลิง จังหวะการจดุระเบิด มลภาวะ เครื่องยนตดีเซลที่ดัดแปลงเพือ่ใชแกสธรรมชาติ ________________________________________________ อาจารยที่ปรึกษาวิทยานิพนธ

iii

ACKNOWLEDGEMENTS

I would like to express my sincere gratitude to Professor Dr.Choi Gyeung Ho

and Assistant Professor Dr.Saiprasit Koetniyom for their helpful guidance, suggestion

and encouragement throughout this study. I am grateful to all Power-Train Laboratory

members, Mr. Lee Hyun Woo, Mr. Baek Seung Yup, Mr. Seo Min Su, Mr. Lee Ju

Hee, Mr. Byun Chang Hee, Mr. Jeon Kang Hyo, Mr. Seo Jong Woo, Mr. Park Sam

Hoon, Mr. Kim Gyung Taek and Mr. Kim Eun Taek and Mr. Kim Young Jo. I would

like to thank all my teachers, family, friends and the staffs of The Sirindhorn

International Thai-German Graduate School of Engineering, King Mongkut’s

Institute of Technology North Bangkok, for their valuable assistance throughout the

entire research. Finally, I am also indebted to Keimyung University and EROOM

Company for the academic supports.

Chedthawut Poompipatpong

iv

TABLE OF CONTENTS Page

Abstract (in English) ii

Abstract (in Thai) iii

Acknowledgements iv

List of Tables vii

List of Figures viii

List of Abbreviations and Symbols xii

Chapter 1 Introduction 1

1.1 Introduction 1

1.2 Literature Review 1

1.3 Objective and Scope of Study 6

1.4 Benefit for this Research 7

Chapter 2 Theory 9

2.1 The Four-stroke Spark-Ignition Engine (SI) 9

2.2 Emissions in SI Engine 17

2.3 Parameters of SI Engine 20

2.4 Calculation Parameters 27

Chapter 3 Experimental Methodology 33

3.1 Equipment and Instruments 33

3.2 Testing Procedure 46

Chapter 4 Results and Discussions 51

4.1 Effects of Loads 51

4.2 Effects of Speeds 56

4.3 Effects of Intake Valve Timings 64

4.4 Effects of Injection Timings 72

4.5 Effects of Ignition Timing 81

Chapter 5 Conclusions and Recommendations 91

5.1 Conclusions 91

5.2 Recommendations for Future Works 92

References 95

v

TABLE OF CONTENTS (CONTINUED) Page

Appendix A Natural Gas Property 99

Appendix B Experimental Calculation 103

Biography 107

vi

LIST OF TABLES

Table Page

3-1 Natural Gas Diesel Engine Specification 34

4-1 MBT Timing at 25% Load 83

4-2 MBT Timing at 50% Load 83

4-3 MBT Timing at WOT 84

4-4 Comparison between MBT Ignition Timing and Retard Ignition Timing

for Camshaft no.1 88





vii

LIST OF FIGURES

Figure Page

2-1 P-V Diagram of the Ideal Four- Stroke Otto Cycle 10

2-2 P-V Diagram of the Ideal Two- Stroke Otto Cycle 10

2-3 P-V Diagram of the Mechanical Four- Stroke Otto Cycle 11

2-4 P-V Diagram of the Ideal Four- Stroke Miller Cycle 13

2-5 P-V Diagram of the Ideal Two- Stroke Miller Cycle 13

2-6 P-V Diagram of Four- Stroke Miller Cycle with Supercharger 14

2-7 P-V Diagram of Two- Stroke Miller Cycle with Supercharger 15

2-8 Comparison of Fired and Unfired Cycle 16

2-9 Emissions in Different Air-Fuel Ratio in General SI Engine 17

2-10 Sources of Emissions in SI Engine 17

2-11 SFC Plotted against Power Output for Varying Throttle Setting 20

2-12 Effect of Engine Speed 21

2-13 Effect of Engine Speed and Equivalence Ratio on the NOx 22

2-14 Valve Timing: Small Valve Overlap 23

2-15 Valve Timing: Large Valve Overlap 23

2-16 Effect of Intake Valve Timing on the Unit Air Charge 24

2-17 Effect of Ignition Timing 25

2-18 Effect of Ignition Timing on Pressure-Crank Angle 26

2-19 Effect of Ignition Timing on Pressure-Volume Diagram 26

2-20 Power and Losses 28

2-21 Indicated Mean Effective Pressure 30

3-1 The Original and Modified Pistons 34

3-2 The Natural Gas Diesel Engine Modified from

Daedong 4A220A-S1 Diesel Engine 35

3-3 Dynamometer 36

3-4 Dynamometer Controller 37

3-5 Exhaust Gas Analyzer 38

viii

LIST OF FIGURES (CONTINUED)

Figure Page

3-6 (a) Two Pipes Connected to the Exhaust Gas Analyzer

(b) Position of the Pipes 39

3-7 Motec ECU 40

3-8 Motec ECU Computer Control Program 40

3-9 Sensors – ECU – Actuators 41

3-10 Position of ISC, TP Sensor and Injectors 41

3-11 MAP Sensor, Pressure Sensor and

Engine Water Temperature Sensor 42

3-12 Engine Oil Temperature Sensor 43

3-13 Exhaust Gas Temperature Sensor and Lambda Sensor 43

3-14 TDC Sensor and Crank Angle Sensor 44

3-15 Laminar Flow Meter 45

3-16 Gas Flow Meter 45

3-17 Overall System 46

3-18 Data Collecting Arrangement for a Camshaft 47

3-19 Dynamometer Control Program 48

3-20 MOTEC ECU Manager 48

3-21 Data from the Exhaust Gas Analyzer 49

3-22 Intake Valve Timing 50

4-1 Effect of Load on Power at High Speed 51

4-2 Effect of Load on Torque 52

4-3 Effect of Load on Power at Low Speed 52

4-4 Effect of Load on SFC 53

4-5 Effect of Load on Brake Thermal Efficiency 53

4-6 Effect of Load on Volumetric Efficiency 54

4-7 Effect of Load on THC 54

4-8 Effect of Load on NOx 55

4-9 Effect of Load on CO 55

4-10 Effect of Load on O2 56

ix


Figure Page

4-11 Effect of Load on CO2 56

4-12 Effect of Speed on the Power Output 57

4-13 Effect of Speed on the Volumetric Efficiency at WOT 58

4-14 Effect of Speed on the Torque Output at WOT 58

4-15 Effect of Speed on the Volumetric Efficiency at 25% Load 59

4-16 Effect of Speed on the Torque Output at 25% Load 59

4-17 Effect of Speed on the Brake Thermal Efficiency at 25% Load 60

4-18 Effect of Speed on the Brake Thermal Efficiency at WOT 60

4-19 Effect of Speed on the Brake Thermal Efficiency at WOT 61

4-20 Effect of Speed on the Volumetric Efficiency at 25% Load 62

4-21 Volumetric Efficiency at WOT for Camshaft no.1 62

4-22 Effect of Speed on THC Emission 63

4-23 Effect of Speed on NOx Emission 63

4-24 NOx at 25% Load for Camshaft No.3 64

4-25 Effect of Intake Valve Timing on the Power Output 65

4-26 Effect of Intake Valve Timing on the Torque Output 65

4-27 Effect of Intake Valve Timing on the SFC 66

4-28 Effect of Intake Valve Timing on the Brake Thermal Efficiency 66

4-29 Volumetric Efficiency versus Speed 67

4-30 Effect of Intake Valve Timing on the THC Emission 68

4-31 Effect of Intake Valve Timing on the NOx 68

4-32 Effect of Intake Valve Timing on the CO 69

4-33 Effect of Intake Valve Timing on the CO2 at 1500 rpm. 69

4-34 Effect of Intake Valve Timing on the CO2 at 2000 rpm. 70

4-35 Brake Thermal Efficiency versus Ignition Timing

2500 rpm. 25% Load at Injection Timing 40ºBTDC 71


2500 rpm. 25% Load at Injection Timing 8ºBTDC 72

x


Figure Page


2500 rpm. 25% Load at Injection Timing 103.5ºATDC 72

4-38 Effect of Injection Timing on Power 73

4-39 Effect of Injection Timing on Torque 74

4-40 Effect of Injection Timing on Volumetric Efficiency 74

4-41 Effect of Injection Timing on SFC 75

4-42 Effect of Injection Timing on Brake Thermal Efficiency 76

4-43 Effect of Injection Timing on THC 76

4-44 Effect of Injection Timing on NOx 77

4-45 Knocking at 1500 rpm in Camshaft No.3 78

4-46 Knocking at 2000 rpm in Camshaft No.3 78

4-47 CO2 at 1500 rpm 25% Load for Camshaft No.2 79

4-48 CO2 at 1500 rpm 50% Load for Camshaft No.2 79

4-49 CO at 1500 rpm 25% Load for Camshaft No.2 80

4-50 CO at 1500 rpm 50% Load for Camshaft No.2 80

4-51 CO2 Concentration according to the Knocking 82

4-52 THC Concentration according to the Knocking 82

4-53 O2 Concentration according to the Knocking 83

4-54 MBT at 2000 rpm and WOT versus Torque 85

4-55 MBT at 2000 rpm and WOT versus Power 85

4-56 MBT at 2000 rpm and WOT versus SFC 86

4-57 MBT at 2000 rpm and WOT versus Brake Thermal Efficiency 86

4-58 MBT at 2000 rpm and WOT versus THC 87

4-59 MBT at 2000 rpm and WOT versus NOx 87

xi

LIST OF ABBREVIATIONS AND SYMBOLS

ABDC After Bottom Dead Center

ATDC After Top Dead Center

BDC Bottom Dead Center

BMEP Brake Mean Effective Pressure

BSFC Brake Specific Fuel Consumption

BTDC Before Top Dead Center

CA Crank Angle

CO Carbon Monoxide

CO2 Carbon Dioxide

ECU Electronic Control Unit

EIVC Early Intake Valve Closure

FMEP Friction Mean Effective Pressure

HC Hydrocarbon

IMEP Indicated Mean Effective Pressure

ISC Idle Speed Controller

LIVC Late Intake Valve Closure

LPG Liquefied Petroleum Gas

MAP Manifold Absolute Pressure

MEP Mean Effective Pressure

MBT Maximum Brake Torque

m Mass

N engine speed

NCMH Normal Cubic Meter per Hour

NGV Natural Gas Vehicle

NOx Oxides of Nitrogen

P Power

PS Pferde Starke

Pb Brake Power

Pi Indicated Power

xii

LIST OF ABBREVIATIONS AND SYMBOLS (CONTINUED)

Q Heat

S Surface Area

SFC Specific Fuel Consumption

SI Spark-Ignition

t Time

T Temperature

TDC Top Dead Center

THC Total Hydrocarbon

TP Throttle Position

U Internal Energy

V Volume

vd Cylinder Swept Volume

W Work

WOT Wide Open Throttle

ε Compression Ratio or Expansion Ratio

ηb Brake Thermal Efficiency

ηi Indicated Thermal Efficiency

ηm Mechanical Efficiency

ηth Thermal Efficiency

ηv Volumetric Efficiency

ρ Density

xiii

CHAPTER 1

INTRODUCTION

1.1 Introduction

Natural gas is the very new alternative fuel for Thai society. In the past, Thai

people knew only LPG (Liquefied Petroleum Gas) which is used in taxis and

households. In fact, the Thai government has been doing research on NGV (Natural

Gas Vehicle) for a long time [1]. After the oil crisis, natural gas has potential to be

promoted as a new alternative fuel.

Price of natural gas is definitely cheaper than that of diesel or gasoline [1]. This

advantage can draw attention to many people. The benefits of using natural gas are to

reduce the imported fuel and to have much less pollution.

1.2 Literature Review

There are many previous studies about natural gas engine. In addition, most of

them focus on the studying of improving the engine performance and emission.

Yusoff et al. [2] worked on finding the effects in different valve timing and

ignition issues in compressed natural gas direct injection. The intake and exhaust

valves must open and close at the right time. Otherwise, the efficiency, fuel

consumption and emission will be poor. Injection time and ignition time also have to

be in exactly right time to produce maximum power and minimum pollution.

Moreover, they found the better spray characteristics at proper pressure and

temperature can accelerate the air-fuel mixing.

Kalam et al. [3] tried to improve a natural gas engine. They compared between

the gasoline and natural gas in three difference situations based on the same engine.

They found that: natural gas gave 15% - 20% lower power than gasoline but the

specific fuel consumption was also 18% less. This was testing at the same throttle

position. For the same output power, natural gas also had lower fuel flow rate and

better emission except the NOx (Oxides of Nitrogen). Finally, they set up the output

2

power of natural gas 10% higher. They found the fuel consumption of natural gas was

little higher but the emission was much better except the NOx.

Department of Mechanical Engineering, Federal University of Technology [4]

compensated the longer ignition delays and slower burning rates by advanced

injection timing. The testing was undergoing a natural gas diesel engine (compressed

ignition engine). The standard injection timing was 30º BTDC (Before Top Dead

Center). The advanced injection timing was 33.5º BTDC. Result found that advanced

injection timing was not recommended for high load condition because of high HC

(Hydrocarbon). The test was continued by advancing another 1.5º more. But the

engine could not run smoothly.

Michael et al. [5] investigated on naturally aspirated Miller Cycle SI engine

(Spark-Ignition Engine) with LIVC (Late Intake Valve Closure) based on first and

second law analyses. Their analytical methodology was on two computer-modeling

tools. They assumed that the cylinder was divided into two zones, unburned and

burned zone. Each zone was uniform. Combustion was modeled as a turbulent flame.

Heat transfer, homogeneous mixture, temperature etc. were considered as well. They

found that LIVC required less fuel to produce the same output and could achieve up

to 6.3% higher indicated thermal efficiency at part load. LIVC had thermomechanical

advantage due to higher intake manifold pressure.

Yorihiro et al. [6] applied the Miller cycle to a lean-burn gas engine

cogeneration. They tested two types of combustion chamber shape. One was high

turbulent type and the other was low unburnt type. They found the low unburnt type

less likely to cause knocking. Therefore, advance ignition timing could be applied

which improved the exhaust of total hydrocarbon concentration and thermal

efficiency. The higher swirl ratio was, the higher temperature and heat loss were.

Bassett et al. [7] simulated a simple and cheap mechanism that allows two-state

LIVC control. This device allowed the engine to operate with wider than normal

throttle settings at low load, which reduced pumping losses. They located a reed valve

in the intake manifold. At full load, reed valve prevented the charge from being

rejected out from the cylinder. At low load, the reed valve allowed the charge to

return freely. This can reduce BSFC (Brake Specific Fuel Consumption) around 7%

and also reduce NOx.

3

Shiga et al. [8] found that the intake capacity chamber installation reduced the

pumping loss by applying LC (Late Closing). They varied the valve timing and

compression ratio. They found that the pumping loss trend was not really affected by

the expansion ratio but it was mainly affected by intake valve timing. And pumping

loss could be decreased by LC. They could not clearly conclude the effect of intake

valve timing on the BSFC. But BSFC decreased with the increasing of expansion

ratio. The experiment results could be explained by calculations that the expansion

ratio was ten times as effective as the compression ratio in increasing the thermal

efficiency.

Chih Wu et al. [9] used the computer simulation the Miller cycle comparing to

Otto cycle based on thermodynamic method. They simulate both Miller cycle with

and without supercharger. The Miller cycle without supercharger processed lower

mass than Otto cycle without supercharger. The pressure and temperature at the end

of compression process were lower. Then they assumed the intake pressure to be 110

kPa for supercharge Miller cycle. They still found that temperature at the end of

compression stroke was lower than that of Otto cycle without supercharger. The net

work, MEP (Mean Effective Pressure) and mass inside the cylinder output of Miller

cycle were also lower than that of Otto cycle without supercharger. Then they

simulated the Mazda engine that operated on Miller cycle. The pressure of

supercharger was 196.5 kPa which higher than they simulated. The result was that

there was more mass in the cylinder, higher MEP and more net work output. They

suggested that Miller cycle should operate with supercharger.

Gyeung Ho Choi et al. [10] simulated the Miller cycle through the computer

simulation according to the EIVC (Early Intake Valve Closure) and LIVC (Late

Intake Valve Closure) method by construction the test engine using the engine

analysis program and by changing the valve close timing. The real engine was also

tested. They observed that the error from the simulation was 5 Pferde Starke (PS)

Finally, they found that the intake valve closing at 55 degree ABDC increased power,

torque and brake thermal efficiency around 2 PS, 1.5 kg·m and 2% respectively.

Wang et al. [11] studied the Miller cycle to reduce NOx emission in a diesel

engine. They compare the original valve timing with three different Miller cycles.

Late intake valve opens and early intake valve closures are used as follow. Miller 1,

4

the intake valve opened 20º late and closed 20º earlier. Miller 2, the intake valve

opened 25º late and closed 25º earlier. Miller 3, the intake valve opened 10º late and

closed 10º earlier. They found that the different output powers were quite small.

Miller cycle 1 was the best for reducing NOx, which can reduce more than 10%. The

exhaust gas temperatures of Miller cycle were lower than normal.

Alla et al. [12] researched on effect of injection timing on the performance of

dual fuel engine. They worked on a single cylinder indirect injection diesel engine

fueled with gaseous fuel. Diesel fuel was used as the pilot fuel and methane or

propane was used as the main fuel, which was inducted in the intake manifold to be

mixed with the intake air. Three values of injection timings of 25º, 27.5 º and 30º

BTDC were used in the test. They found that retarding injection timing (at 25º BTDC)

delayed combustion. The temperature of mixture is not enough to propagate in the

whole mixture. The amount of unburned hydrocarbon and CO (Carbon Monoxide)

increase as injection timing retards. While NOx and thermal efficiency increases with

the advanced injection timing

Takagaki and Raine [13] used a single cylinder, spark ignition engine to study

effects of the compression ratio on nitric oxide emissions using natural gas. They

found that for fixed ignition timing nitric oxide emissions increased with increasing

compression ratio. But for Maximum Brake Torque (MBT) timing, nitric oxide

emissions first increased and then decreased.

Koichi et al. [14] investigated the effect of Miller cycle on MEP for high-

pressure supercharged gasoline engine. Intake valve closing timing was set at 75

degrees in the case of the late intake valve closure. They found that the exhaust gas

temperature did not increase and the maximum BMEP (Brake Mean Effective

Pressure) increased because of knocking limit improvement. Miller-cycle with a

supercharger, which is highly efficient at high-pressure ratio and an intercooler, with

high efficiency, can increase IMEP (Indicated Mean Effective Pressure).

Caton [15] simulated the nitric oxide emissions in spark-ignited automotive

engine using a cycle simulation, which employed three zones for the combustion

process: unburned gas, adiabatic core region and boundary layer gas. The effects of

engine parameters such as equivalence ratio, ignition timing, inlet manifold pressure

and engine speed were examined. He found that maximum nitric oxide was at about

5

equivalent ratio of 0.9. Nitric oxide increased as advanced ignition timing and higher

inlet manifold pressures. For an equivalent ratio of 0.9, the decreasing available time

as engine speed increases dominates the increase of gas temperature.

Caton [16] focused on the effect of compression ratio on nitric oxide emissions

for a spark ignition engine. The study completed for a commercial, 5.7 liters spark

ignition V-8 engine operating at a part load condition at 1400 rpm with an

equivalence ratio of one and MBT (Maximum Brake Torque) ignition timing. He

mentioned that there are many researches on this effect, which showed different

results. A number of previous studies indicated that the increment of compression

ratio increased nitric oxide emissions. However, other studies showed the opposite.

He expected that results might be affected by uncontrolled and variable condition

(temperature, pressure and humidity) and the ignition timing. Furthermore, the

conclusion might be different depending on whether the ignition timing was constant

or set to MBT ignition timing and equivalence ratio. For his investigation, he adjusted

to provide MBT timing and constant throttle position (constant load). He found that

increasing the compression ratio resulted in decreasing brake specific nitric oxide

value due to the changes of gas temperature, cylinder pressure and brake specific fuel

consumption. However, it could decrease as compression ratio increased at high

compression ratio, which might involve with the burn duration.

Engineers at Tokyo Gas Co., Ltd., and Yanmar Diesel Engine Co., Ltd. [17]

modified a diesel engine to Miller-Cycle natural gas engine. They both designed for

EIVC and LIVC for this 23.15-liter engine. LIVC required lower cost and fewer

design changes. The engine operated on premixed natural gas with turbocharger. It

was also a close loop control. The engine could achieve 36.1% brake thermal

efficiency. Moreover, the cogeneration system produced 300 kW of electric and

achieved 83.5% energy efficiency.

Sarkhi et al. [18] modeled the efficiency of a Miller engine in term of

thermodynamics calculation. They found that the effects of the temperature-dependent

specific heat of the working fluid on the cycle performance were significant and

should have been considered in design. A slight increase in some parameter would

have an impact on the thermal efficiency of the cycle.

6

Akira et al. [19] developed Miller gas engine for the purpose of attaining

electrical efficiency equivalent to that of a diesel engine on the basis of the lean burn

gas engine for high efficiency and low NOx emission. Miller cycled gas engine

cogeneration package improved efficiency to 40% level by the Miller cycle.

Mohamed [20] used propane as a fuel. He tested at the speed of 1500 to 3000

rpm with the interval of 500. He varied load of 50%, 75% and 100%. He also tested at

different ignition timings and found the relation among BMEP, speed, load, ignition

timing, MBT, BSFC and emission. Results showed that the engine could be operated

with propane over a wide range of air-fuel ratios with less carbon dioxide, carbon

monoxide and hydrocarbon emissions compare to operation with gasoline. The

differences in fuel characteristics, the operation of the engine on propane were

accompanied with some power loss. The fuel economy of the engine on propane got

poor with increase in speed from 2500 to 3000 rpm. HC and CO of the propane was

lower comparing to gasoline. But CO2 (Carbon Dioxide) was higher.

Lee Ju Hee [21] researched on the thermal efficiency on an industrial engine

with Miller cycle. A diesel engine was retrofitted to natural gas engine for better

duration. He changed the closing time of intake valve for adapting Miller cycle.

Intake cam lift compensation test was added on the EIVC test and also effective

compression pressure compensation test was added on the LIVC test. He found that

EIVC had less thermal efficiency than the basic cam experiment. LIVC test at 51

degree-ABDC (After Bottom Dead Center) bettered the fuel consumption ratio around

5-8% and brake thermal efficiency around 2-3%. LIVC test at 77 degree-ABDC

bettered the fuel consumption ratio and brake thermal efficiency around 3-7% and

1-2% respectively. The quantity of NOx was reduced about 5-10%.

1.3 Objective and Scope of Study

The objectives of the work are to study the influences of intake valve timing and

injection timing in a natural gas diesel engine. In addition, finding the tendencies of

engine efficiencies in different intake valve closures and injection timings is one of

the purposes.

In this research, the effects of loads, speeds, intake valve timings and injection

timings on the efficiencies and emissions will be studied under the compression ratio

7

(expansion ratio) of nine, speed of 1500 rpm, 2000 rpm and 2500 rpm with the

equivalent air-fuel ratio of 1.0.

1.4 Benefit for this Research

This research shows the influences of loads, engine speeds, intake valve

timings, gas injection timings and ignition timings. The result will show how affective

each parameter is. Therefore, the benefit of the investigation would result a clearer

way of improving any retrofitted engines. Engine development procedure will be

shortened and become more efficient in the future.

CHAPTER 2

THEORY

2.1 The Four-Stroke Spark-Ignition Engine

The most common internal combustion engine is the four-stroke Otto engine,

which was invented by a German engineer, Nikolaus August Otto, in 1876 [22].

The four stroke ideal Otto cycle, as shown in figure 2-1, models the intake air-

fuel mixture as piston moves from TDC (Top Dead Center) to BDC (Bottom Dead

Center) during the intake valves open (isobaric process 1-2). Then the air-fuel mixture

is compressed when the piston moves upward to TDC with the intake valve closure

(isentropic process 2-3). At the TDC, The spark suddenly ignites the air-fuel mixture

to provide the heat energy input (isochoric process 3-4). The air-fuel mixture expands

and pushes the piston downward to BDC as usually called power stroke (isentropic

process 4-5). At BDC, the exhaust valves open so the pressure drops (isochoric

process 5-6). Lastly, the piston moves upward to pump out the combustion products

with the open exhaust valves (isobaric process 6-1).

The area of the graph represents both work (W) done and work added. The net

work output is W45 – W23. While W12 – W61 is zero because they are ideally equal.

The important processes are compression stroke and power stroke (process 2-3,

process 3-4, process 4-5 and process 5-6). This can be focused as two-stroke Otto

cycle, as shown in figure 2-2, and assumed as a closed system.

In fact, the intake and exhaust valves do not open and close right at the TDC

and BDC. The valves normally have early open and late closure. Therefore, the real

situation of figure 2-1 is presented in figure 2-3. This represents the actual four-stroke

Otto cycle. The pressure of the intake stroke is normally lower than exhaust stroke in

a naturally aspirated engine.

This is to show the differences of actual Otto cycle and ideal two-stroke Otto

cycle, which is commonly used in the calculation.

http://techni.tachemie.uni-leipzig.de/otto/otto_g0_eng.html

10

Otto cycle assumes that air is a perfect gas with constant specific heat and all

the processes are fully reversible.

Otto cycle assumes that there is no intake and exhaust process that means

quantity of air is fixed in closed system.

Otto cycle assumes that the heat addition process has no internal combustion but

heat is from external source.

Otto cycle assumes that heat rejection is to the environment, which is different

from blow-down and exhaust process.

FIGURE 2-1 P-V Diagram of the Ideal Four- Stroke Otto Cycle

FIGURE 2-2 P-V Diagram of the Ideal Two- Stroke Otto cycle

11

FIGURE 2-3 P-V Diagram of the Mechanical Four- Stroke Otto cycle

The thermal efficiency of this cycle is considered as:

H

L

H

LHth Q

QQ

QQ−=

−= 1η Eq.2-1

)1()1(1

)()(1

343

252

34

25

−−

−=−−

−=TTTTTT

TTmCTTmC

v

vthη Eq.2-2

Note that:

5

4

1

4

5

1

3

2

2

3

VV

VV

TT

TT

kk

=⎟⎟⎠

⎞⎜⎜⎝

⎛=⎟⎟

⎠

⎞⎜⎜⎝

⎛=

−−

This makes:

2

5

3

4

TT

TT

= Eq.2-3

Substitution of Eq. 2-3 into Eq. 2-2 gives:

12

)1(

3

2 11 kth T

T −−=−= εη Eq.2-4

Where ε is the compression ratio.

The thermal efficiency of Otto cycle can be obviously increased by increasing

the value of compression ratio (ε). But in practical, the increase of compression ratio

raises the temperature in the cylinder. This causes knocking because of auto ignition,

which makes the temperature and pressure inside cylinder severely rise. The engine

can get damage. In order to solve this problem, increasing the value of ε by

maintaining the limit of compression ratio is the way.

The Miller cycle was patented by Ralph Miller, an American engineer, in the

1940s. This cycle has the potential to increase the efficiency and net power in spark

ignition internal combustion engine. This takes advantages on Otto cycle by

maintaining the compression ratio and increasing the expansion ratio. Therefore, this

Miller cycle has high value of V5/V4 (which is the value of ε) and operates without

knocking.

The basic of Miller cycle is almost the same as Otto cycle. It is also a four-

stroke cycle with a little difference in intake stroke or compression stroke. Intake

valve closes before or after the piston reach BDC.

EIVC is the Miller cycle that intake valves close before the piston reaches the

BDC in intake stroke. LIVC is the Miller cycle that intake valves close after the piston

start moving upward in compression stroke.

The P-V diagram is shown in figure 2-4. The process 1-2 is an isobaric intake

process. The process 2-3 is an isobaric compression process because the intake valves

still open. The process 3-4 is an isentropic compression process. The process 4-5 is an

isochoric heat adding process. The process 5-6 is an isentropic expansion process. The

process 6-7 is an isochoric cooling process. Finally, the process 7-1 is an isobaric

exhaust process.

Notify that this P-V diagram represents the Miller cycle without supercharger.

Comparing to the Otto cycle, the two-stroke P-V diagram is shown in figure 2-5.

http://en.wikipedia.org/w/index.php?title=Ralph_Miller_%28engineer%29&action=edit

http://en.wikipedia.org/wiki/1940s

13

FIGURE 2-4 P-V Diagram of the Ideal Four- Stroke Miller Cycle

FIGURE 2-5 P-V Diagram of the Ideal Two- Stroke Miller Cycle

The thermodynamic analysis of this two-stroke Miller cycle is:

W12 = U1-U2 Eq.2-5

Q23 = U3-U2 Eq.2-6

W34 = U3-U4 Eq.2-7

Q45 = U5-U4 Eq.2-8

14

Q56 – W56 = U6- U5 Eq.2-9

W56 = P5 (V6 – V5) Eq.2-10

The net work of this cycle is:

Wnet = W12 + W34 + W56 Eq.2-11

The cycle efficiency is:

η = Wnet / Q23 Eq.2-12

As the piston moves upward while the intake valve opens. Some of air-fuel

mixture is pumped out that causes lower amount of mixture, maximum temperature

and maximum pressure at the compression top dead center. This causes the

combustion efficiency to be reduced. To increase the mass of mixture, supercharger is

used. Supercharger increases the intake pressure to be higher as shown in figure 2-6

and figure 2-7

FIGURE 2-6 P-V Diagram of Four- Stroke Miller Cycle with Supercharger

15

FIGURE 2-7 P-V Diagram of Two- Stroke Miller Cycle with Supercharger

Combustion can occur normally and abnormally. In spark ignition, the frame

front should propagate throughout the mixture steadily. Nevertheless, if pre-ignition

or self-ignition occurs, abnormal combustion is started. Pre-ignition is a situation that

the air-fuel mixture ignites by hot spot such as exhaust valve. Self-ignition is a

situation that air-fuel mixture ignites by the temperature and pressure.

Normal Combustion in SI engines with homogeneous mixture, the combustion

process can be divided into three periods. Firstly, a spark (or called as high-

temperature plasma [23]) is discharged between the spark plug electrodes. The spark

causes a small nucleus of flame that propagates into unburnt gas. Combustion starts

very slowly because the frame size is small. It does not generate enough energy to

heat the surrounding gas quickly. This causes a delay period as shown in figure 2-8.

Delay period is usually about 7.5 degrees of crank angle after spark plug firing [24].

This delay period depends on the temperature, cylinder pressure and composition of

the air-fuel mixture.

The second stage of combustion is known as “frame propagation period” as

shown in figure 2-8. By the time the first 5-10% of mixture is burned, the combustion

process is well set up and frame moves quickly in the combustion chamber due to

induced turbulence. The chemical reaction time in this period is very short and the

16

frame front speed increases. Figure 2-8 also shows that the maximum pressure usually

occurs around 5 – 20 degrees ATDC (After Top Dead Center) [24]. The reason can be

explained as follow, since the mixture is ignited before top dead center (at the end of

compression stroke), there is a pressure rise from the combustion before the end of

compression stroke, and an increase from the compression (negative work).

Advancing the ignition timing causes both the pressure to rise before top dead center

and also the compression work to increase. However, the high pressure at top dead

center leads to higher expansion pressure (positive work). The optimization between

these two effects is the “minimum ignition advance for best torque” or Maximum

Brake Torque (MBT) ignition timing.

The last period is the flame termination. Even though, the piston already moved

down, the combustion volume increases. Cylinder temperature and pressure decrease.

The reaction still occurs in slow rate and adds a little more work to the piston.

FIGURE 2-8 Comparison of Fired and Unfired Cycle [24]

Abnormal Combustion, when the mixture contacts with hot area such as exhaust

valve, pre-ignition occurs. Pre-ignition causes an increase in temperature and

pressure. This increases the compression work and decreases the power. Moreover,

17

because of high temperature and pressure, pre-ignition brings the system to self-

ignition.

Self-ignition occurs when the unburned gas instinctively ignites. This is a

rapidly rise in pressure and causes knocking. If the pre-ignition occurs early, the self-

ignition will occur early and give a severe knock

2.2 Emissions in SI Engine

Improving the performance of engine to be the highest is not enough for

engineering work. The concern of air pollution is a very important issue. The exhaust

gases from spark ignition engine consist of oxides of nitrogen, carbon monoxide and

unburned hydrocarbons. These emissions are worse spark ignition engine more than

from compress ignition engine because emissions from compress ignition engine are

primarily soot and odour associated with certain hydrocarbons [24].

Figures 2-9 and 2-10 show the variations of emissions with air-fuel ratio and

main sources of emissions in SI engine.

FIGURE 2-9 Emissions in Different Air-Fuel Ratio in General SI Engine

18

FIGURE 2-10 Sources of Emissions in SI Engine [24]

Stoichiometric Combustion, the reaction between fuel and air becomes the

composition of products. Fuel is usually in form of hydrocarbon. If there is enough

oxygen, the entire hydrocarbon will be completely burned and become carbon dioxide

and water (H2O). Actually, oxygen in the combustion process cannot be gotten purely.

Nitrogen in the air also becomes a reactant. However, nitrogen does not really affect

the reaction because of the consideration at low temperature.

Stoichiometric combustion can be called as chemically correct or theoretical.

This means there is just enough oxygen for the combustion. All the fuel can be

converted to product without the rest of oxygen. So the stoichiometric air-fuel ratio

depends on the type of fuel.

Ideally, the products of combustion are only carbon dioxide and water. The

percentage of carbon dioxide in the product (exhaust gas) shows how the fuel and air

efficiently combust. Moreover, this can be related to the level of power output.

The stoichiometric combustion should not have exhaust oxygen. Practically, the

mixing between air and fuel can be imperfect. Because of this, some part is unburned

or not completely burned. The amount of exhaust oxygen indicates the quality of air-

fuel mixture that was combusted.

The exhaust hydrocarbon emissions come from incomplete combustion. The

unburned hydrocarbon in the exhaust was a useful fuel that could be burned and give

19

out some more work. It presents the waste of fuel. Therefore, the amount of exhaust

hydrocarbon can indicate the trend of combustion efficiency.

The unburned hydrocarbon emissions are from different sources. In

compression and combustion, the high-pressure forces some mixture into crevice like

piston ring grooves. These crevices are too narrow for frame to enter. And this

mixture is one case of unburned hydrocarbon. “. . . To reduce hydrocarbon emission,

excess air should be supplied until the reduced flammability of the mixture causes a

net decrease in hydrocarbon emission . . .” (Mohamed, 1998: 23) [20]. The heating

value is an important factor. Because the higher heating value is, the higher energy for

work avails. The reduction of unburned hydrocarbon increases the power output and

engine efficiency.

There are some other causes of hydrocarbon emissions. Figure 2-9 shows that

hydrocarbon emission levels are a strong function of air-fuel ratio. At the rich mixture

zone, there is not enough oxygen to react with the fuel (hydrocarbon). This leads to

high level of hydrocarbon and carbon monoxide in the exhaust. If the air-fuel mixture

is lean, poor combustion occurs because of misfire.

During compression and combustion, some of mixture is forced through the

crevice around valves and valve seats. Or else during valve overlap, the mixture flows

directly into the exhaust. A well-design engine reduces these problems.

The exhaust carbon monoxide emission is controlled by air-fuel ratio as

obviously shown in figure 2-9. When there is not enough oxygen in the combustion.

Some carbon ends as CO. The exhaust carbon monoxide emission also indicates the

fuel conversion efficiency. It represents the lost of chemical energy that is not

perfectly utilized. CO can be said as a fuel that can be combusted to supply thermal

energy (CO + 1/2 O2 CO2 + heat).

Increasing the rate of carbon dioxide reduces the carbon monoxide and makes

the combustion process approaches the theory.

Exhaust gas of an engine can have oxides of nitrogen. Most of that will be

nitrogen oxide (NO), with a little amount of nitrogen dioxide (NO2) and other

nitrogen-oxygen combinations. These are not desirable emission. NOx is from

nitrogen in the air.

20

Oxides of nitrogen increase with increasing frame temperature. Slightly rich

mixture should give the highest oxides of nitrogen. But the formation of NO needs

oxygen. The maximum point in figure 2-9 locates at lean mixture area. Another factor

is the frame speed. Lower speed with lean mixture gives longer time for NOx

formation. Retarding the ignition timing is a way to reduce the highest pressure and

temperature. This causes reduction in NOx, output power and economy.

2.3 Parameters of SI Engine

2.3.1 Effects of Loads

Figure 2- 11 shows a curve called “fish-hook”. This presents the effects of load

on specific fuel consumption and brake mean effective pressure (because BMEP is

independent of engine size). In each throttle position, the curve also shows the

influence of air-fuel ratio on SFC (Specific Fuel Consumption) and BMEP.

FIGURE 2-11 SFC Plotted against Power Output for Varying Throttle Setting [23]

Load is one of basic parameters. BMEP and torque increase as load increases

while SFC decreases, comparing at a same speed. Andrew J.K. and group [25] found

that CO emission factor and NOx emission factor are higher in higher load.

21

2.3.2 Effects of Speeds

Figure 2-12 shows the effect of engine speed on output torque, power and fuel

consumption. At low speed, main energy loss is the heat transfer (heat loss) while

friction loss is dominant at high speed. Therefore, a particular engine model has

maximum torque at a specific engine speed.

FIGURE 2-12 Effect of Engine Speed [23]

Engine speed affects the NOx emission. One of the previous researches was

studied by Maher A.R. and Sadiq Al-Baghdadi [26]. The effects of engine speed and

equivalent ratio with the compression ratio of eleven were investigated, as shown in

figure 2-13. At the optimum spark timing for the maximum brake torque, the NOx

emission increases as the engine speed increases for all equivalence ratios less than

0.8. This is due to the increment in the maximum temperature in the cycle with

excessive oxygen. However, the NOx emission decreases as the engine speed

increases for all equivalence ratios more than 0.8 due to a decreasing amount of

oxygen.

The effect of speed that involves with intake valve timing will be also discussed

in section 2.3.3.

22

FIGURE 2-13 Effect of Engine Speed and Equivalence Ratio on the NOx [26]

2.3.3 Effects of Valve Timings

The valve timing is basically controlled by the camshaft. Figure 2-14 shows the

typical engine but figure 2-15 shows the high performance engine. The longer valve

overlap period has advantage in high-speed engine. Both the intake and exhaust valve

must have the appropriate timing for their open and closure.

Yusoff et al. [2] researched on valve timing. Practically, the intake and exhaust

valve do not open and close exactly at the TDC or BDC. They explain the effects of

inappropriate timing as follow. If the exhaust valve opens too late, the volume of

exhaust gas increases. This leads to higher pumping losses. If the exhaust valve opens

too early, some work available from expanding gas would be lost. Engine does not

take full advantage on power stroke. Moreover, there is less inertia at TDC.

Therefore, there is less force on incoming air during overlap period and lower

volumetric efficiency. If exhaust valve closes too late, piston sucks the exhaust gas re-

enter or air-fuel mixture flows out directly to exhaust valve, which causes poor

economy and emission. If intake valves open too early, the exhaust gas will go back

to intake port and block the fresh air. If the intake valve closes too late, maximum

pressure and temperature will be low and lead to low combustion efficiency.

23

FIGURE 2-14 Valve Timing: Small Valve Overlap [24]

FIGURE 2-15 Valve Timing: Large Valve Overlap [24]

The work from the engine depends directly on the amount of energy released

when air-fuel mixture burns. Therefore, both air and fuel are equally important. For

24

this reason, the induction of air becomes one of the greatest problems. The weight of

air inducted to the engine on one intake stroke is normally called “unit air charge”.

Figure 2-16 shows three different intake valve closure timing with an opening point.

In all cases, the intake valve starts to open before TDC. Therefore, the piston

descends during the intake stroke with large valve-opening area that is not to throttle

the intake airflow.

FIGURE 2-16 Effect of Intake Valve Timing on the Unit Air Charge [27]

Assuming that the piston moves very slow as the average speed approaches zero,

the throttle loss would be small because the air velocity through the valve would be

very low. The air charge will be maximum if the intake valve closes at BDC. As the

piston speed increases, the velocity of air would increase. The throttle loss also

increases. So the curve AB is shown.

In the same case of very low piston speed, if the exhaust valve closes after BDC.

A part of the charge will be pushed back into the intake manifold. A unit air charge

will reach only point C. When speed is increased, the momentum of incoming air

increases. It continues to charge the cylinder even though the piston reaches BDC and

begin the compression stroke. This is the advantage of late intake valve closure as

shown in line CD. Beyond the point D, the fluid friction is more than the gain from

the momentum charging of the cylinder. The unit air charge decreases to point E.

Similarly, if the intake valve closes at point F, the maximum point of unit air

charge can be shift to higher speed.

25

2.3.4 Effect of Injection Timing

For spark-ignition engine, the injection timing has much less effect to the engine

performance than in the compress-ignition engine. Because advancing or retarding the

injection timing in compress-ignition engine means that the combustion occurs earlier

or later in the cycle. Injection timing in diesel engine performs a great role as same as

ignition timing in spark-ignition engine. Many previous research showed that

adjusting injection timing only few degrees can give an obviously improve engine

performances and emissions. On the other hand, spark-ignition engine does not

meticulously focus on this parameter. However, injection timing in spark-ignition

engine can affect the temperature of intake air, mixture distribution and engine output

especially in the liquid phase gaseous fuel [28].

2.3.5 Effect of Ignition Timing

Ignition timing is a definitely important factor that must be controlled

accurately. Power output, efficiency and emission are the responses of it. If the

ignition is too late, the piston work in power stroke is low because of lower pressure.

Moreover, the probability of incomplete combustion by the time exhaust valve opens

is high. On the other hand, if the ignition timing is too early, there is too much

pressure before end of compression stroke. This increases the work in compression

stroke instead of power stroke (negative work instead of positive work). The early

ignition timing can also cause knock as shown in figure 2-17.

FIGURE 2-17 Effect of Ignition Timing [24]

26

Richard S. [24] modeled the effect of ignition timing by a computer model.

Figures 2-18 and 2-19 show the effects of 15 degrees different from MBT point.

There are differences in maximum pressure and the angles of the peak. The ignition at

15-degree before MBT provides very high pressure, which should give high torque. In

fact, this moves the peak point near to the TDC. That means the compression work

increases. While ignition at 15-degree after MBT ignition timing needs less

compression work but the piston already moves down, this leads to low cylinder

pressure.

FIGURE 2-18 Effect of Ignition Timing on Pressure-Crank Angle [24]

FIGURE 2-19 Effect of Ignition Timing on Pressure-Volume Diagram [24]

27

2.3.6 Other Parameters of SI Engine

2.3.6.1 Air-Fuel Ratio; generally, the exhaust emission varies with the air-

fuel ratio. CO is mainly affected by the air-fuel ratio. Other factor has very small

effect. In case of HC, the lean mixture increases HC because of frame propagating

imperfection. While rich mixture does not have enough O2 for the combustion.

Therefore, HC also increases. NOx is highest at near stoichiometric because of high

temperature.

2.3.6.2 Ignition Timing; ignition timing influences the NOx and HC

concentrations. Ten-degree-late ignition point can decrease NOx up to 30%. NOx is a

result of combustion temperature and ignition timing. Late ignition timing also

decreases HC because there is not enough time for combustion. Therefore,

combustion continues to the exhaust stroke and in the exhaust pipe. This leads to

higher exhaust temperature.

2.3.6.3 Intake Air Condition; high temperature incoming fresh air increases

NOx and is a reason of lower air density, which leads to lower output power.

Incoming air temperature has very small effect on HC emission.

2.3.6.4 Engine S/V Ratio; The ratio of combustion chamber surface area to

volume affects the concentration of HC and NOx. Generally, this valve is low because

engineers try to reduce heat loss. This leads to low HC and high NOx.

2.4 Calculation Parameters

2.4.1 Theoretical Efficiency

It is also called thermodynamic efficiency or air-standard efficiency. It is a

function of compression ratio and method of combustion. This cycle assumes that air

is the working substance.

2.4.2 Ideal Efficiency

This is the efficiency of ideal engine. It is assumed that there is no heat loss to

the walls. The same working substances as real engine are considered.

2.4.3 Indicated Thermal Efficiency

The comparison between the performances of engines sometimes has to ignore

the effect of mechanical losses. This is to use the indicated efficiency for the means of

examining the thermodynamic process in an engine.

28

consumedheatpowerindicated

i =η Eq.2-13

2.4.4 Mechanical Efficiency

The difference between indicated power (ip or Pi) and brake power (bp of Pb) is

the concern of friction power. This means the combination of brake power and

friction power equals to the indicated power as shown in figure 2-20. If the

mechanical losses are low, the value of brake power will close to indicated power.

This obviously leads to the high mechanical efficiency.

powerindicatedpowerfriction

powerindicatedpowerbrake

m −== 1η Eq.2-14

Since indicated power = brake power + friction power Eq.2-15

Energy losses (exhaust, coolant,

radiation)

Mechanical losses

Useful Energy

FIGURE 2-20 Power and Losses

2.4.5 Brake Thermal Efficiency

This is usually called overall efficiency. It shows the final output efficiency.

From figure 2-20, the energy from fuel converts to useful energy with many losses.

The brake thermal efficiency shows how much the energy can be used from 100% of

available energy from fuel.

29

consumedheatpowerbrake

b =η Eq.2-16

And also mib ηηη = Eq.2-17

2.4.6 Volumetric Efficiency

The volumetric efficiency of an engine defined as the ratio of the mass of air

inducted by the engine on the intake stroke to the theoretical mass of air that should

be inducted by filling the piston displacement volume with air at atmospheric

condition.

Volumetric efficiency is a measure that shows the effectiveness of induction and

exhaust processes.

t

av m

m=η Eq.2-18

Where: ma ; actual mass of air inducted per intake stroke.

mt ; theoretical mass of air to fill the piston displacement volume under

atmospheric condition

“… The name “volumetric efficiency” is a misnomer because actually it is a

mass and not a volume ratio…” (Edward, 1973: 48) [27].

2.4.7 Mean Effective Pressure (MEP)

The mean effective pressure represents the ratio of work per combustion cycle

to the displacement volume of piston as shown in figure 2-21. This is also called

specific work.

cylinderpervolumeSweptcyclemechanicalpercylinderperWorkMEP = Eq.2-19

Indicated Mean Effective Pressure (IMEP) is the area under p-v diagram. The

IMEP measures the indicated work per swept volume.

30

cylinderpervolumeSweptcyclemechanicalpercylinderperworkIndicatedIMEP = Eq.2-20

FIGURE 2-21 Indicated Mean Effective Pressure

Brake Mean Effective Pressure (BMEP), the work output from the engine is

usually measured by a dynamometer. This is more important than indicated mean

effective pressure.

cylinderpervolumeSweptcyclemechanicalpercylinderperworkBrakeBMEP = Eq.2-21

According to Eq. 2-15, it can be re-written to

FMEPBMEPIMEP += Eq.2-22

2.4.8 Specific Fuel Consumption (SFC)

In engine testing, another parameter that should be determined is the fuel

consumption. Fuel consumption is mass flow per unit time (mass flow rate).

31

However, the more useful parameter is specific fuel consumption. It tells how the

engine uses the fuel for producing work.

outputPowerrateflowFuelSFC = Eq.2-23

CHAPTER 3

EXPERIMENTAL METHODOLOGY

A diesel engine was dedicated for using with natural gas by modifying the

pistons. Fuel pump and fuel injectors are replaced by spark plugs. Compression ratio

has been reduced to 9 : 1. In fact, natural gas has higher octane number than that of

gasoline. The compression ratio for the operation should be higher than typical

gasoline engines. However, the purpose of this investigation was not focusing on the

value but the experiment was set to investigate the influences of each parameter.

Choosing a relatively low compression ratio is a way to provide long torque curve

without knocking. The data would be more appropriate for the analyzing.

The engine was installed to an eddy current dynamometer. The dynamometer

measured torque at the flywheel directly without losses from driveline. All the sensors

were connected, which would be shown in this chapter.

This experiment was mainly to compare the differences among three intake

valve closures. Notify that the intake valve opening time and exhaust valve timing

were not changed. Changing camshaft profiles was a way to this experiment.

Therefore, this experiment needed three different camshafts.

Each camshaft was also tested in various loads. Every load, three different

injection timings were tested to achieve the objective. In each injection time, many

ignition timings were tested to find the MBT.

3.1 Equipment and Instruments

A diesel engine, Daedong 4A220A-S1, was totally dedicated to natural gas

diesel engine with natural gas injectors and close loop controller. The pistons were

redesigned from the diesel compression ratio of twenty-two to the compression of

nine as shown in figure 3-1. Diesel pump and injectors were replaced by spark plugs.

Table 3-1 shows the dedicated engine specification.

Figure 3-2 demonstrates the Daedong 4A220A-S1 natural gas diesel engine

located on the dynamometer and attached with several sensors in the engine test

34

laboratory, Department of Mechanical and Automotive Engineering, Keimyung

University, Republic of South Korea.

(a) Compression Ratio of 22 (b) Compression Ratio of 9

FIGURE 3-1 The Original and Modified Pistons

TABLE 3-1 Natural Gas Diesel Engine Specification

Item Natural Gas Diesel Engine

(dedicated engine)

Type 4-cylinder, 4-stroke engine

Displacement 2,197 cc.

Bore (mm.) 87

Stroke (mm.) 92.4

Compression Ratio 9.0

Fuel Supply System Gas Injectors

35

FIGURE 3-2 The Natural Gas Diesel Engine modified from

Daedong 4A220A-S1 Diesel Engine

3.1.1 Dynamometer

The dynamometer, shown in figure 3-3, is an eddy current dynamometer. It

measures output at the flywheel without transmission or driveline. There are no losses

affected to the results. Eddy current dynamometer operates on the principle of slip

loss that occurs when electrically conductive drum rotates against a stationery and

non-uniform flux distribution around its periphery. The relative speed causes the flow

of eddy currents in drum material by the law of electromagnetic induction. The

reactive magnetic field, resulted from induced currents, is responsible for the braking

torque. This dynamometer is operated under the dynamometer controller, shown in

figure 3-4, and a dynamometer control program.

Calibration of the dynamometer makes all the data reliable. Dynamometer

calibration can be done by this following method.

3.1.1.1 Warm up the dynamometer following the dynamometer

manufacturer’s specifications.

3.1.1.2 Determine the dynamometer calibration moment arm.

Dynamometer manufacturer’s data, actual measurement, or the value recorded from

the previous calibration.

36

3.1.1.3 When calibrating the engine flywheel torque transducer, any lever

arm used to convert a weight or a force through a distance into a torque must be in a

horizontal position.

3.1.1.4 Calculate the indicated torque for each calibration weight to be

used by calibration weight multiplies by calibration moment arm.

3.1.1.5 Attach each calibration weight specified into the moment arm at

the calibration distance determined in 3.1.1.2 of this section. Record the power

measurement equipment response to each weight.

3.1.1.6 For each calibration weight, compare the torque value measured

in paragraph e) to the calculated torque from 3.1.1.4.

3.1.1.7 The measured torque must be within 2 percent of point or 1

percent of the engine maximum torque of the calculated torque.

3.1.1.8 If the measured torque is not within the requirements, adjust or

repair the system. Repeat steps 3.1.1.1 to 3.1.1.7 again.

Dynamometer control program provides several useful output data such as

power, torque, atmospheric pressure, exhaust gas temperature and intake air

temperature.

FIGURE 3-3 Dynamometer

37

FIGURE 3-4 Dynamometer Controller

3.1.2 Exhaust Gas Analyzer

The machine, shown in figure 3-5, is used for measuring amount of exhaust

gases that are NOx, HC, CO, CO2 and O2. The exhaust gas analyzer is connected to

two pipes. First pipe is connected between the muffler and catalyst (exit of catalyst

but entrance of muffler). The second pipe is connected at the entrance of catalyst as

shown in figure 3-6. Therefore, the experiment can compare the emissions between

entrance and exit of catalyst if needed.

38

FIGURE 3-5 Exhaust Gas Analyzer

39

(a)

Second pipeFirst pipe

Muffler ----------- Catalyst ---------- Engine

(b)

FIGURE 3-6 (a) Two Pipes connected to the Exhaust Gas Analyzer

(b) Position of the two Pipes

40

3.1.3 ECU (Electronic Control Unit)

The ECU, shown in figure 3-7, receives signals from sensors and after that, this

ECU will send the signal to actuators. While the engine is being tested, a computer

program that connects directly to this Motec ECU controls the ignition timing and

injection timing.

Figure 3-8 shows the screen of Motec computer program. The ignition timing is

inserted into this program.

FIGURE 3-7 Motec ECU

FIGURE 3-8 Motec ECU Computer Control Program

41

3.1.4 Sensors and Other Instruments

Sensors are the major equipments in the control loop. MAP sensor (Manifold

Absolute Pressure), TDC sensor, Crank Position sensor, Throttle position sensor,

Lambda sensor, and temperature sensors are used. The signals of these sensors are

sent to the ECU. Output from the ECU is to control the idle speed, ignition timing and

fuel injection timing as shown in figure 3-9.

FIGURE 3-9 Sensors - ECU - Actuators

Idle Speed Controller (ISC)

Injectors

Throttle Position Sensor

FIGURE 3-10 Position of ISC, TP Sensor and Injectors

42

Figure 3-10 is focusing on the fuel system. There is a main natural gas pipe

feeding gas to the four injectors. Idle speed controller is also in this fuel supply

system while the throttle position sensor is a very important equipment for controlling

the load in the experiment.

Manifold absolute pressure sensor, cylinder pressure sensor, engine water

temperature sensor, engine oil temperature sensor, exhaust gas temperature sensor,

lambda sensor, TDC sensor and crank angle sensor are attached in different positions

as shown in figures 3-11 to 3-14.

Engine Water Temperature Sensor

MAP SensorPressure Sensor

FIGURE 3-11 MAP Sensor, Pressure Sensor and

Engine Water Temperature Sensor

43

FIGURE 3-12 Engine Oil Temperature Sensor

Exhaust Gas Temperature Sensor

Lambda Sensor

FIGURE 3-13 Exhaust Gas Temperature Sensor and Lambda Sensor

44

Crank Angle Sensor

TDC Sensor

FIGURE 3-14 TDC Sensor and Crank Angle Sensor

To find the brake thermal efficiency, the data of airflow rate and fuel flow rate

must be determined. Figures 3-15 and 3-16 are laminar flow element and gas flow

meter respectively. Laminar flow element tries to control the intake airflow quality

and uses the principle of pressure to provide the airflow rate (in term of volume flow

rate). Therefore, air mass flow rate can be finally calculated. Gas flow meter is

connected between the natural gas pipe and gas injectors. The data is given in the unit

of NCMH (Normal Cubic Meter per Hour), which is the natural gas volume flow rate.

Finally, figure 3-17 shows the overall system.

45

FIGURE 3-15 Laminar Flow Element

FIGURE 3-16 Gas Flow Meter

46

FIGURE 3-17 Overall System

3.2 Testing Procedure

This research was focusing on the intake valve timing and injection timing.

Therefore, three camshafts are used for giving three different valve timings. Each

valve timing was tested in 25%, 50% and 100% loads. The speeds of 1500, 2000 and

2500 rpm. are experimented in each load. Three different injection timings are tested

in every speed. MBT was found by changing the ignition timing. The compression

ratio of nine and equivalent air-fuel ratio of 1 are the test condition. Figure 3-18

shows the data collecting arrangement for each intake valve timing. The ignition

timings were varied between 15 and 54 degree BTDC with the interval of 3 degrees.

Therefore, to complete the experiment, engine testing must go over all the processes

of figure 3-18 three times.

47

FIGURE 3-18 Data Collecting Arrangement for a Camshaft

Firstly, the first camshaft was installed in the natural gas diesel engine. All the

sensors were connected. The engine then started warming up until the cooling water

temperature reached 80ºC.

The dynamometer controller was set to 1500 rpm. and 25% load as shown in

figure 3-19 and figure 3-20. In the main area of MOTEC ECU control screen,

injection timing and ignition timing can be inserted. Figure 3-20 also shows an

example of injection timing and ignition timing controlling. From the figure, injection

timing is 368 degrees (before the start of power stroke, which means 8 degrees

BTDC, or the intake valve opening time) and ignition timing is 48 degrees BTDC.

48

FIGURE 3-19 Dynamometer Control Program

FIGURE 3-20 MOTEC ECU Manager

49

Before collecting the data in each ignition timing, the engine must be running

under the condition of equivalent air-fuel ratio of 1 (air-fuel ratio of 16.83). So

adjusting the amount of injected natural gas was needed. Figure 3-21 shows the air-

fuel ratio of 23.74 from the exhaust gas analyzer. After the ratio was set to 16.83

already, the data collection started. The dynamometer control program collected the

output data such as power, torque, engine speed, temperature, etc. While the exhaust

gas analyzer collected the data of CO, CO2, NOx, O2, THC, airflow and fuel flow.

The experiment finished after collecting all data from figure 3-18.

FIGURE 3-21 Data from the Exhaust Gas Analyzer

Figure 3-22 shows the intake valve timings both open and closure of all

camshafts. The three camshafts have exactly same intake valve opening time and

exhaust valve timing. The differences are the intake valve closure as follow.

50

51º ABDC 145.5° ATDC

119.5° ATDC

103.5° ATDC

Camshaft No.3

Camshaft No.2

CamshaftNo.1

FIGURE 3-22 Intake Valve Timing

Camshaft no.1, intake valves start opening at 8 degree BTDC during the exhaust

stroke. The maximum lift is at 103.5 degree ATDC in the intake stroke. The intake

valves close 35 degree ABDC.


stroke. The intake valves close at 51 degree ABDC (16 degree later than camshaft

no.1). Therefore, the maximum valve lift period is between 103.5 and 119.5 degree

ATDC in the intake stroke.


stroke. The intake valves close at 77 degree ABDC (42 degree later than camshaft

no.1). Therefore, the maximum valve lift period is between 103.5 and 145.5 degree

ATDC in the intake stroke.

CHAPTER 4

RESULTS AND DISCUSSIONS

This section deals with the analysis, presentation and discussion of results

obtained in the investigation. All analyses and presentations were done using

Microsoft Excel program. For meaningful and ease of comparison, trends from the

tests are presented on the same charts. The four parameters are presented versus

ignition timing. In additional, ignition timing is another important parameter.

Therefore, the topic of “Effects of the Ignition Timings” is put in section 4.5.

4.1 Effects of Loads

Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC

32

37

42

47

52

15 18 21 24 27 30 33 36 39 42 45 48 51 54

Ignition timing (ºBTDC)

Pow

er (P

s) 25%50%100%

FIGURE 4-1 Effect of Load on Power at High Speed

Figure 4-1, shows effect of loads on power, is harmonious with figure 2-11. The

higher load has higher power and torque output. This experiment found that the

change in low load has much effect than the change in high load as shown in figures

4-1 and 4-2; there is much difference between 25% load and 50% load while 50%

load and full load does not show much difference. At the speed of 1500 rpm, the

52

difference of power output between 25% load and full load is about 3 PS

(Pferde Starke), as shown in figure 2-3, whereas the speed of 2500 rpm presents the

difference of 12 PS. This validates that load has more effect at higher speed.


89

1011

12131415

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Torq

ue (k

g m

)

25%50%100%

FIGURE 4-2 Effect of Load on Torque


22232425

26272829

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Pow

er (P

s)

25%50%100%

FIGURE 4-3 Effect of Load on Power at Low Speed

Specific fuel consumption and brake thermal efficiency are perhaps the most

important parameter in the evaluation the overall performance of the engine operating

on the given fuel. Enumerating again, brake specific fuel consumption and brake

thermal efficiency base on the same data but they present in different visions.

53

Figure 2-11 also mentions the effect of load on SFC. At low load, the fuel flow is

relatively low. However, after comparing to the output power, the graphs show an

obviously high specific fuel consumption and low brake thermal efficiency, which

corresponds to figure 2-11, section 2.3.1. Figures 4-4 and 4-5 are plotted at the speed

of 2500 rpm. Therefore, the difference between a quarter load and full load is

noticeable. If the speed of 1500 rpm were presented, all the curve lines would be very

close.


7

7.5

8

8.5

9

9.5

10

15 18 21 24 27 30 33 36 39 42 45 48 51 54


SFC

(10-5

g/J

)

25%50%100%

FIGURE 4-4 Effect of Load on SFC


19

20

21

22

23

24

25

15 18 21 24 27 30 33 36 39 42 45 48 51 54


SFC

(10-5

g/J

)

25%50%100%

FIGURE 4-5 Effect of Load on Brake Thermal Efficiency

54

The higher load allows more air-fuel mixture to come into the cylinder. The

value of volumetric efficiency represents on this behavior. In low speed, three curve

lines are very close to each other. While the speed increases, the 50% and 100% curve

lines seem to separate from 25% curve line as shown in figure 4-6.


50

60

70

80

90

100

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Volu

met

ric

effic

ienc

y (%

)

25%50%100%

FIGURE 4-6 Effect of Load on Volumetric Efficiency

Effects of load on THC (Total Hydrocarbon) and NOx emissions are not so

much huge as the effects of valve timing, which will be shown in the forthcoming

topics. There is not obvious difference in the THC emission. However, figures 4-7

and 4-8 show that THC and NOx emissions increase as the load increases.


0

200

400

600

800

1000

15 18 21 24 27 30 33 36 39 42 45 48 51 54


THC

(ppm

)

25%50%100%

FIGURE 4-7 Effect of Load on THC

55


0

500

1000

1500

2000

2500

3000

15 18 21 24 27 30 33 36 39 42 45 48 51 54


NO

x (pp

m)

25%50%100%

FIGURE 4-8 Effect of Load on NOx

Andrew et al. [22] concluded their research on the effects of vehicle speed and

engine load on motor vehicle emissions that, the higher load leads higher NOx and CO

emissions. Comparing to this experiment, the result is going in the same way but the

reason that, this experiment did not show palpable differences, could be because of

the fuel type. Figures 4-9 to 4-11 support the research of Andrew and group by

displaying the high CO and O2 with low CO2 .


0

0.2

0.4

0.6

0.8

1

1.2

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO (%

) 25%50%100%

FIGURE 4-9 Effect of Load on CO

56


00.10.20.3

0.40.50.60.7

15 18 21 24 27 30 33 36 39 42 45 48 51 54


O2 (

%) 25%

50%100%

FIGURE 4-10 Effect of Load on O2


10.3

10.4

10.5

10.6

10.7

10.8

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO

2 (%

) 25%50%100%

FIGURE 4-11 Effect of Load on CO2

4.2 Effects of Speeds

The result, displayed in figure 4-12, shows that every camshaft between the

speed 1500 and 2500 rpm, the higher speed gives higher output power than the lower

ones.

57

Camshaft No.2; WOT at Injection timing 40ºBTDC

20

25

30

35

40

45

50

15 18 21 24 27 30 33 36 39 42 45 48 51 54

Ignition Timing (ºBTDC)

Pow

er (P

s) 1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-12 Effect of Speed on the Power Output

Figures 4-13 to 4-16 present the torque output and volumetric efficiency of

camshaft no.2 at WOT load and 25% load. Camshaft no.2 has high volumetric

efficiency at the speed of 2000 rpm for WOT load condition and the lowest

volumetric efficiency at the speed of 1500 rpm as shown in figure 4-13. The curve

order is the same as torque output curve in figure 4-14 that camshaft no.2 has the

highest output torque at 2000 rpm and the lowest at 1500 rpm.

Figures 4-15 and 4-16 are from the condition of 25% load. They show the

advantage of lower speed on volumetric efficiency and output torque. They also show

the similarity of curve order. Practically, the useful point is only the maximum brake

torque ignition timing. If this point is focused, it does not always show that higher

volumetric efficiency condition has higher brake torque. However, these figures

proved that torque curve in figure 2-13 does not depend only on heat transfer and

friction loss but torque curve is also affected by the volumetric efficiency, which can

be thought further as the effects of speed and valve timing since volumetric efficiency

definitely depends on speed and valve timing. Figures 4-13 to 4-16 show that the

trend of torque output can be roughly predicted from the volumetric efficiency curve.

58


50

60

70

80

90

100

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Volu

met

ric

effic

ienc

y (%

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-13 Effect of Speed on the Volumetric Efficiency at WOT


1111.5

1212.5

1313.5

1414.5

15

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Torq

ue (k

g m

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-14 Effect of Speed on the Torque Output at WOT

59

Camshaft No.2; 25% load. at Injection timing 8ºBTDC

50

55

60

65

70

75

80

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Volu

met

ric

effic

ienc

y (%

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-15 Effect of Speed on the Volumetric Efficiency at 25% Load


99.510

10.511

11.512

12.513

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Torq

ue (k

g m

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-16 Effect of Speed on the Torque Output at 25% Load

At low load, shown in figure 4-17, the brake thermal efficiency of high speed is

the lowest even though the power output is the highest because the fuel consumption

of high speed is much more than that of the lower ones. As the load increases, the

value of brake thermal efficiency also increases. However, the effect of speed seems

to be less at high load as shown in figure 4-18. Figure 4-19 is plotted from data of

camshaft no.3. Late intake valve closure does not seem to be appropriate to low-speed

high-load condition.

60


20

21

22

23

24

25

26

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Brak

e Th

erm

al E

ffici

ency

(%

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-17 Effect of Speed on the Brake Thermal Efficiency at 25% Load


20

21

22

23

24

25

26

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Brak

e Th

erm

al E

ffici

ency

(%

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-18 Effect of Speed on the Brake Thermal Efficiency at WOT

61


2020.5

2121.5

2222.5

2323.5

24

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Brak

e Th

erm

al E

ffici

ency

(%

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-19 Effect of Speed on the Brake Thermal Efficiency at WOT

The effects of speeds allies with the effects of intake valve timings. Therefore,

the volumetric efficiency will be discussed in section 4.3, “Effects of Intake Valve

Timings”. However, there are some more interesting points if the graphs are

compared.

From data of camshaft no.1 at 25% load, shown in figure 4-20, it can be roughly

predicted that the speed, which gives the highest possible volumetric efficiency might

be lower than 1500 rpm or just a little higher than 1500 rpm. For the simplicity, the

speed of 1500 rpm is assumed to give highest possible volumetric efficiency at 25%

load. The speed of 2000 rpm and 2500 rpm provide lower volumetric efficiency

because the air-fuel mixture moves faster, the throttle loss also increases.

As the load increases to WOT, shown in figure 4-21, the speed of 1500 rpm

does not have enough momentum to fill the cylinder. Accordingly, the air-fuel

mixture is pushed back to the intake manifold and has a lower volumetric efficiency.

While the speed of 2500 rpm operates under high throttle loss as already mentioned.

62


60

65

70

75

80

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Volu

met

ric

effic

ienc

y (%

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-20 Effect of Speed on the Volumetric Efficiency at 25% Load


80

85

90

95

100

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Volu

met

ric

effic

ienc

y (%

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-21 Volumetric Efficiency at WOT for Camshaft no.1

The air-fuel flow rate is highest at 2500 rpm and lowest at 1500 rpm. The

amount of THC emission should be highest at 2500 rpm and lowest at 1500 rpm.

Nevertheless, figure 4-22 demonstrates that the result is in the opposite way.

The exhaust temperature under the speed of 1500 rpm, 2000 rpm and 2500 rpm

are in the range of 590-660 ºC, 640-715 ºC and 680-760ºC respectively. From this

data, it can be believed that, this might involve with the combustion time. As the

speed increases, the combustion cannot complete within the power stroke. Therefore,

the combustion might continue in exhaust stroke and exhaust pipe, which causes the

63

higher exhaust temperature. Because of not having enough combustion time, the

combustion temperature in the combustion chamber is also lowered. NOx, which

primarily depends on the combustion temperature, reduces as speed increases, as

shown in figure 4-23. On the other hand, it might be said that there is less time for

NOx formation in high-speed condition. Maher A.R. and Sadiq Al-Baghdadi [23] did

an experiment on hydrogen fuel. They found that NOx emission decreases as the

speed increases if the equivalent ratio is over 0.8.


0

200

400

600

800

1000

15 18 21 24 27 30 33 36 39 42 45 48 51 54


THC

(ppm

)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-22 Effect of Speed on THC Emission


0

500

1000

1500

2000

2500

3000

15 18 21 24 27 30 33 36 39 42 45 48 51 54


NO

x (pp

m)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-23 Effect of Speed on NOx Emission

64

Figure 4-24 shows that engine speed almost has no effect on the NOx emission

in case of low effective compression ratio and low combustion temperature.


0

500

1000

1500

2000

2500

3000

15 18 21 24 27 30 33 36 39 42 45 48 51 54


NO

x (pp

m)

1500 rpm.2000 rpm.2500 rpm.

FIGURE 4-24 NOx at 25% Load for Camshaft No.3

4.3 Effects of Intake Valve Timings

Figures 4-25 to 4-28 are the test result at 2000 rpm, WOT (Wide Open Throttle)

at the injection timing of 40ºBTDC. These graphs show the trend of brake power,

torque, SFC and brake thermal efficiency according to the change in ignition timing.

The lowest SFC and highest brake thermal efficiency occur at MBT ignition timing.

In every tested speed and injection timing, the output power and torque from camshaft

no.1 is higher than camshaft no.2. In addition, the result from camshaft no.2 is higher

than camshaft no.3.

65

2000 rpm. WOT at Injection timing 40ºBTDC

20

25

30

35

40

45

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Pow

er (P

s) Camshaft No.1Camshaft No.2Camshaft No.3

FIGURE 4-25 Effect of Intake Valve Timing on the Power Output


89

10111213141516

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Torq

ue (k

g m

)

Camshaft No.1Camshaft No.2Camshaft No.3

FIGURE 4-26 Effect of Intake Valve Timing on the Torque Output

66


6

7

8

9

15 18 21 24 27 30 33 36 39 42 45 48 51 54


SFC

(10

-5 g/

J)


FIGURE 4-27 Effect of Intake Valve Timing on the SFC


20

21

22

23

24

25

26

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Brak

e Th

erm

al E

ffici

ency

(%

)


FIGURE 4-28 Effect of Intake Valve Timing on the Brake Thermal Efficiency

Figure 4-29 is plotted from the engine speed versus volumetric efficiency. As

mentioned in chapter 3, this experiment was undergoing the engine speed of 1500,

2000 and 2500 rpm. Therefore, the data was not in detail enough to find the best

engine speed for each valve timing. Nevertheless, if this data is brought to the theory

in section 2.3.3, the graph can be roughly fitted by curve lines and shows the

approximate highest volumetric efficiency for each valve timing.

The solid lines are plotted from actual data while the dashed lines are rough

plots. This figure shows that the highest volumetric efficiency for camshaft no.1, 2

67

and 3 are approximately at 2100, 2200 and 2300 rpm respectively (in the yellow area).

Comparing figure 4-29 to figure 2-16, they give out the same logic that the later

intake valve closure has lower volumetric efficiency in low speed. However, it will

take advantage in high speed. Camshaft no.1 is preferable up to the speed around

3000 rpm. Then, camshaft no.2 will have higher volumetric efficiency. Finally,

camshaft no.3 will become the best in volumetric efficiency point of view starting in

some speed very high.

This is to emphasize again that, figure 4-29 wants to indicate only the effect on

volumetric efficiency. As the speed increases, the friction loss also increases.

Therefore, the higher volumetric efficiency does not mean higher brake thermal

efficiency. For example, volumetric efficiency of camshaft no.2 at 3500 rpm is

certainly higher than that of camshaft no.1. However, it is not possible to conclude

that brake thermal efficiency of camshaft no.2 is higher.

FIGURE 4-29 Volumetric Efficiency versus Speed

From the result of volumetric efficiency curves in figure 4-29, comparing

among the same engine speed, the higher volumetric efficiency means more intake

air-fuel mixture is sucked into the cylinder. This makes the obvious result in THC

exhaust emission. In addition, the higher effective compression ratio increases the

combustion chamber temperature, which results to the higher NOx emission as shown

in figure 4-30 and 4-31. This point must be emphasized again that the higher

68

volumetric efficiency does not mean the higher air-fuel flow quantity if the engine

speeds are different.


0

200

400

600

800

1000

15 18 21 24 27 30 33 36 39 42 45 48 51 54


THC

(ppm

)


FIGURE 4-30 Effect of Intake Valve Timing on the THC Emission


0

500

1000

1500

2000

2500

3000

15 18 21 24 27 30 33 36 39 42 45 48 51 54


NO

x (p

pm)


FIGURE 4-31 Effect of Intake Valve Timing on the NOx

This experiment cannot clearly indicate the effect of valve timing on the CO

emission as shown in figure 4-32. There is no significant change in CO emission

according to the change in the intake valve timing or ignition timing.

69


0

0.5

1

1.5

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO (%

) Camshaft No.1Camshaft No.2Camshaft No.3

FIGURE 4-32 Effect of Intake Valve Timing on the CO

CO2 is another output that has some uncertain trend. In the speed of 1500 rpm,

combustion from camshaft no.1 seems to give high product of, figure 4-33.

While speed increases to 2000 rpm and 2500 rpm, the result has a little change to

figure 4-34.

1500 rpm. 25% load at Injection timing 40ºBTDC

10.110.210.310.410.510.610.710.810.9

11

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO2

(%) Camshaft No.1

Camshaft No.2Camshaft No.3

FIGURE 4-33 Effect of Intake Valve Timing on the CO2 at 1500 rpm.

70


10.310.410.510.610.710.810.9

1111.1

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO2

(%) Camshaft No.1


FIGURE 4-34 Effect of Intake Valve Timing on the CO2 at 2000 rpm.

The most important issue of Miller Cycle is to investigate the brake thermal

efficiency. Chin Wu and group [9] simulated the Miller Cycle comparing to Otto

Cycle based on thermodynamic method and strongly recommended that Miller Cycle

should operate with supercharger. Because their simulation showed that, the Miller

cycle without supercharger processed lower mass than Otto cycle without

supercharger. The pressure and temperature at the end of compression process were

lower. On the other hand, Lee J.H. [21] researched on a natural gas diesel engine

based on the effects of Miller Cycle, equivalent ratio and injection timing. He proved

that Miller cycle without supercharging could increase brake thermal efficiency up to

3% and reduce specific fuel consumption up to 8%.

This investigation found that late intake valve closure reduces output power,

torque and brake thermal efficiency except the test condition at 25% load 2500 rpm.

At this condition, the injection timing was test in three different points. The results are

plotted versus ignition timing in figures 4-35 to 4-37.

Camshaft no.3, LIVC at 77ºABDC, increases the brake thermal efficiency from

camshaft no.2, LIVC at 51ºABDC, up to 0.42% in the injection timing of 40ºBTDC

as shown in figure 4-35.

71


18

19

20

21

22

23

24

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Brak

e Th

erm

al E

ffici

ency

(%

)


FIGURE 4-35 Brake Thermal Efficiency versus Ignition Timing

2500 rpm. 25% Load at Injection Timing 40ºBTDC

After the injection timing was changed to 8ºBTDC and 103.5ºATDC,

figures 4-36 and 4-37 present an increment of brake thermal efficiency of camshaft

no.2 and 3 over the original camshaft (camshaft no.1). Figure 4-36 shows that the

maximum brake thermal efficiencies of camshaft no.1 and 2 are 22.55024% and

23.63199% respectively. That is increased by 1.08175%. While figure 4-37 shows an

increment of 0.36191%. The data also presents the benefit in term of SFC, which are

4.58% and 1.58% for the injection timing of 8ºBTDC and 103.5ºATDC respectively.

Enumerating again, the only change from figure 4-35 to figures 4-36 and 4-37 is

the injection timing. This is evidence that Miller Cycle can better brake thermal

efficiency if it works with appropriate injection timing.

72


18

19

20

21

22

23

24

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Brak

e Th

erm

al E

ffici

ency

(%

)



2500 rpm. 25% Load at Injection Timing 8ºBTDC

2500 rpm. 25% load at Injection timing 103.5ºBTDC

18

19

20

21

22

23

24

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Brak

e Th

erm

al E

ffici

ency

(%

)



2500 rpm. 25% Load at Injection Timing 103.5ºATDC

4.4 Effects of Injection Timings

Injection timing is another important issue. However, this research found that

injection timing has small effect on the output performance as shown in figures 4-38

and 4-39. The reason might be that the fuel is gaseous, which has less difficulty in

mixing up with air. The other reason could be that the injection timings in this

experiment are not too late. However, late injection timing was tested at the intake

73

valve closing time. The results were not collected because the engine conditions were

severe. The knocking sound occurred with high exhaust emissions. This might be

because the intake valve close timing occurs in the beginning of compression stroke.

The mixture is pumped out from the cylinder or the mixture is sucked in with very

high throttle loss due to very small valve lift. The momentum of mixture could be

very low. It can be believed that the mixture in the combustion chamber is not

homogeneous. Rich mixture located near the intake valve while other locations are

lean mixture. Moreover, the intake valve close timing is very near to the ignition

timing. There is less time for the air-fuel mixing process. Thus, the experiment of this

injection timing did not go on. Nevertheless, this was a very useful experience in

analyzing the effects of injection timing in a natural gas diesel engine.

Camshaft No.2; 2000 rpm. at 50% load

34

35

36

37

38

39

40

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Pow

er (P

s) 40ºBTDC8ºBTDC103.5ºATDC

FIGURE 4-38 Effect of Injection Timing on Power

74


1212.513

13.514

14.515

15 18 21 24 27 30 33 36 39 42 45 48 51 54Ignition timing (ºBTDC)

Torq

ue (k

g m

)

40ºBTDC

8ºBTDC

103.5ºATDC

FIGURE 4-39 Effect of Injection Timing on Torque

Volumetric Efficiency, shown in figure 4-40, seems to be independent from the

injection timing. Volumetric efficiency is calculated from the sucked air mass. If the

sucked medium is pure air, the volumetric efficiency can be higher than air-fuel

mixture suction. However, this experiment did not use fuel direction injection.

Therefore, changing injection timing does not affect the volumetric efficiency but the

ability of mixing with fresh air can affect some exhaust emissions.


60

70

80

90

100

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Volu

met

ric

effic

ienc

y (%

)

40ºBTDC8ºBTDC103.5ºATDC

FIGURE 4-40 Effect of Injection Timing on Volumetric Efficiency

75

Yusoff Ali et al. [2] investigated on compressed natural gas direct injection in a

spark ignition engine. They confirmed that direct injection timing has a very close

interrelation with valve timing. Therefore, the setting of direct injection timing

depends on the both timing of intake and exhaust valve. This can increase volumetric

efficiency, power, and air-fuel mixing ability and reduce emissions.

Figures 4-41 and 4-42 are to review the discussion in the previous topic about

the influence of injection timing. According to the whole data, it is not possible to

find the exact influence on brake thermal efficiency and specific fuel consumption.

There could be some errors from the experiment. Firstly, the fuel is gaseous. The

density is extremely sensitive to the temperature. While the calculation assumed the

constant density. Secondly, SFC represents in the unit of 10-5 g/J. Therefore, a very

small error seems to be magnified by this small scale.


7

7.5

8

8.5

9

15 18 21 24 27 30 33 36 39 42 45 48 51 54


SFC

(10-5

g/J

)


FIGURE 4-41 Effect of Injection Timing on SFC

76

CamshaftNo.2; 2000 rpm. at 50% load

20

21

22

23

24

25

26

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Brak

e Th

erm

al E

ffici

ency

(%

)


FIGURE 4-42 Effect of Injection Timing on Brake Thermal Efficiency

Figures 4-43 and 4-44 show the effect of injection timing on exhaust THC and

NOx emissions. Comparing the effects of the injection timing in camshaft no.1 and 2,

there is no difference. But camshaft no.3 can indicate a small effect that injection

timing of 103.5°ATDC gives slightly higher in THC and NOx emissions.


0

200

400

600

800

1000

15 18 21 24 27 30 33 36 39 42 45 48 51 54


THC

(ppm

)


FIGURE 4-43 Effect of Injection Timing on THC

77


0

500

1000

1500

2000

2500

15 18 21 24 27 30 33 36 39 42 45 48 51 54


NO

x (p

pm)


FIGURE 4-44 Effect of Injection Timing on NOx

A very interesting point in the injection timing is the effect on knocking. This

experiment found a knocking in camshaft no.3. Figures 4-45 and 4-46 show the data

in the speed of 1500 rpm and 2000 rpm respectively. At 1500 rpm, the 103.5° ATDC-

injection timing seems to provide more advance ignition timing while the 40° BTDC-

injection timing can provide advance ignition timing up to 42 degrees. And the 8°

BTDC-injection timing can provide advance ignition timing up to 45 degrees.

The speed of 2000 rpm, the 8° BTDC-injection timing can provide advance

ignition timing up to 48 degrees. But the 40° BTDC-injection timing can provide

advance ignition timing up to 51 degrees. From the result, it is not possible to

conclude that the late injection timing can or cannot provide more advance ignition

timing. Nevertheless, it shows that, for each valve timing and engine speed including

load, there is particularly appropriate injection timing.

78


0500

10001500

2000250030003500

15 18 21 24 27 30 33 36 39 42 45 48 51 54


THC

(ppm

)


FIGURE 4-45 Knocking at 1500 rpm in Camshaft No.3


200

400

600

800

1000

1200

15 18 21 24 27 30 33 36 39 42 45 48 51 54


THC

(ppm

)


FIGURE 4-46 Knocking at 2000 rpm in Camshaft No.3

The data of CO and CO2 also show the frustrating result, shown in figures 4-47

to 4-50. For camshaft no.1 at 1500 rpm 25% load, the 103.5°-ATDC-injection timing

gives out the highest CO2. When the load changes to 50%, it shows the lowest. In the

speed of 2000 rpm, there is no difference among three injection timings. The 103.5°-

ATDC-injection timing gives out the lowest CO2 again in every load of 2500 rpm

while 8°-BTDC-injection timing gives the highest. Camshaft no.2 gives almost the

same trend as camshaft no.1.

79


10.4

10.5

10.6

10.7

10.8

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO

2


FIGURE 4-47 CO2 at 1500 rpm 25% Load for Camshaft No.2


10.4

10.5

10.6

10.7

10.8

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO2


FIGURE 4-48 CO2 at 1500 rpm 50% Load for Camshaft No.2

80


0

0.2

0.4

0.6

0.8

1

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO (%

) 40ºBTDC8ºBTDC103.5ºATDC

FIGURE 4-49 CO at 1500 rpm 25% Load for Camshaft No.2


0

0.2

0.4

0.6

0.8

1

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO (%

) 40ºBTDC8ºBTDC103.5ºATDC

FIGURE 4-50 CO at 1500 rpm 50% Load for Camshaft No.2

Camshaft no.3 at 1500 rpm, The 103.5°-ATDC-injection timing gives out the

highest CO2 while the 40°-BTDC-injection timing gives the lowest. At 2000 rpm,

the 103.5°-ATDC-injection timing still gives out the highest CO2 while the

8°-BTDC-injection timing becomes the lowest. At 2500 rpm, everything seems to be

opposite to 2000 rpm. The 103.5°-ATDC-injection timing becomes the lowest CO2

while the 8°-BTDC-injection timing becomes the highest. The result of CO mainly

81

opposite to the CO2 because if CO reacts with O2 and becomes CO2, the amount of

CO will decrease and increase amount of CO2.

The result cannot bring to the clear answer of the effect of injection timing.

Because this experiment proved that, the injection timing must operate synchronously

with each particular valve timing, engine speed and load. Therefore, the effects of

injection timing must be researched in a very well prepared methodology to give a

useful result.

4.5 Effects of Ignition Timings

The ignition timing with gaseous fuel operation is perhaps the most important

adjustment that can be made to accomplish best engine performance. Ignition timing

affects nearly all the major operating parameters that include specific fuel

consumption, power output, efficiency and tendency to knock.

Referring to figures 4-30 and 4-31, they show the effect of ignition timing on

the THC and NOx emissions. The retard ignition timing causes lower THC emission

but higher exhaust gas temperature because the combustion process continues in the

exhaust stroke and in the exhaust pipe. While early ignition timing makes the peak

pressure moves close to TDC as discussed in figure 2-18. This rises up the cylinder

pressure and temperature, which mainly effects to the NOx emission.

Knocking occurred in this experiment when the ignition timing was too early.

This occurred in camshaft no.3 at ignition timing 45º BTDC, the speed of 1500 rpm

and 25% load. The fuel cannot well react with O2 to become CO2. Therefore, the CO2

curve drops while THC and O2 curves rapidly rise up. This is undesired situation in

the engine operation. This situation is presented in figures 4-51 to 4-53.

82


10.110.210.310.4

10.510.610.710.8

15 18 21 24 27 30 33 36 39 42 45 48 51 54


CO2

(%)

Camshaft No.3

FIGURE 4-51 CO2 Concentration according to the Knocking

1500 rpm. 25% load at Injection timing 40 BTDC

0500

10001500

2000250030003500

15 18 21 24 27 30 33 36 39 42 45 48 51 54

Ignition timing

THC

(ppm

)

Camshaft No.3

FIGURE 4-52 THC Concentration according to the Knocking

83

1500 rpm. 25% load at Injection timing 40 BTDC

0

0.2

0.4

0.6

0.8

1

1.2

15 18 21 24 27 30 33 36 39 42 45 48 51 54

Ignition timing

O2 (

%)

Camshaft No.3

FIGURE 4-53 O2 Concentration according to the Knocking

MBT timing obviously depends on valve timing and speed, referring to the

results. MBT timing also slightly depends on load. Nevertheless, there was not any

evidence that injection timing affected the change in MBT timing.

The following tables show the MBT timing in degree BTDC.

TABLE 4-1 MBT Timing at 25% Load

25% load 1500 rpm. 2000 rpm. 2500 rpm.

Camshaft No.1 24º- 27º 27º- 30º 30º- 33º


Camshaft No.3 33º- 39º 33º- 36º 51º

TABLE 4-2 MBT Timing at 50% Load

50% load 1500 rpm. 2000 rpm. 2500 rpm.

Camshaft No.1 24º 27º 30º



84

TABLE 4-3 MBT Timing at WOT

WOT load 1500 rpm. 2000 rpm. 2500 rpm.


Camshaft No.2 24º 27º- 33º 33º

Camshaft No.3 33º 33º- 36º 42º- 45º

As speed increases, the spark must be advanced to maintain optimum timing

because the period of the combustion process in crank angle degree increases.

Comparing among three camshafts, later intake valve closure needs more advanced

ignition timing. Lower load, especially at low speed, also desires more advanced

ignition timing.

CO and O2 emissions are less than one percent in the exhaust gas. The variation

of CO concentration in the exhaust is minimal because CO emission levels are hardly

affected by spark advance variation (Heywood, 1998) [27]. The highest level of

hydrocarbon emission (about 900 ppm), much lower than THC level for gasoline

engines under normal operating conditions (1000 ppm – 3000 ppm), was observed.

The MBT ignition timing brings high power output, high brake thermal

efficiency and low SFC. This ignition timing is desirable. There is another point of

view if the exhaust emissions are considered. Look in figures 4-54 to 4-59, these

figure are presenting the results of every camshaft at WOT in the speed of 2000 rpm.

Camshaft no.1 has MBT of 14.879 kg-m at ignition timing of 24º BTDC.

Camshaft no.2 has 14.2165 kg-m at 30º BTDC. In addition, camshaft no.3 has 11.825

kg-m at 33º and 36º BTDC.

85


89

10111213141516

15 18 21 24 27 30 33 36 39 42 45 48 51 54

Ignition timi

Torq

ue (k

g m

)


ng (ºBTDC)

MBT timing for camshaft no.1



FIGURE 4-54 MBT at 2000 rpm and WOT versus Torque


20

25

30

35

40

45

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Pow

er (P

s) Camshaft No.1Camshaft No.2Camshaft No.3




FIGURE 4-55 MBT at 2000 rpm and WOT versus Power

86


6

7

8

9

15 18 21 24 27 30 33 36 39 42 45 48 51 54

Ignition timi ºBTD

SFC

(10-5

g/J)


ng ( C)




FIGURE 4-56 MBT at 2000 rpm and WOT versus SFC


20

21

22

23

24

25

26

15 18 21 24 27 30 33 36 39 42 45 48 51 54


Brak

e Th

erm

al E

ffici

ency

(%

)





FIGURE 4-57 MBT at 2000 rpm and WOT versus Brake Thermal Efficiency

87


0

200

400

600

800

1000

15 18 21 24 27 30 33 36 39 42 45 48 51 54

Igni

Camshaft No.1

tion tim

THC

(ppm

)


ing (ºBTDC)




FIGURE 4-58 MBT at 2000 rpm and WOT versus THC


0

500

1000

1500

2000

2500

3000

15 18 21 24 27 30 33 36 39 42 45 48 51 54


NO

x (p

pm)





FIGURE 4-59 MBT at 2000 rpm and WOT versus NOx

This section is to show the benefit of choosing a little retard ignition timing

from MBT timing based on the data in figures 4-54 to 4-59. Tables 4-4 to 4-6 show

the output of using MBT ignition timing comparing to 3-degree and 6-degree retard.

88

TABLE 4-4 Comparison between MBT Ignition Timing and Retard Ignition Timing

for Camshaft no.1

Camshaft no.1 24º BTDC (MBT timing)

21º BTDC

Difference

percentage

(%)

18º BTDC

Difference

percentage

(%)

Torque (kg-m)

Power (Ps)

ηth

SFC(E-05) (g/J)

THC (ppm)

NOx (ppm)

14.879

41.5605

24.98928

7.342624832

775

1845

14.7685

41.2565

24.8314

7.38931024

743

1639

- 0.74266

- 0.73146

- 0.63179

0.63663

- 4.12903

-11.16531

14.6305

40.872

24.48444

7.494021837

714

1446

- 1.67014

- 1.65662

- 2.02022

2.06189

- 7.87097

-21.62602


for Camshaft no.2


27º BTDC

Difference

percentage

(%)

24º BTDC

Difference

percentage

(%)

Torque (kg-m)

Power (Ps)

ηth

SFC(E-05) (g/J)

THC (ppm)

NOx (ppm)

14.2165

39.71

24.21422

7.57765369

738

1759

14.205

39.67

24.15543

7.59609634

703

1613

- 0.08089

- 0.10073

- 0.24279

0.24338

- 4.74255

-8.30017

14.1865

39.6185

24.15842

7.595154482

658

1504

- 0.21102

- 0.23042

- 0.23044

0.23095

- 10.84011

-14.49687

89


for Camshaft no.3


30º BTDC

Difference

percentage

(%)

27º BTDC

Difference

percentage

(%)

Torque (kg-m)

Power (Ps)

ηth

SFC(E-05) (g/J)

THC (ppm)

NOx (ppm)

11.825

33.023

23.41653

7.83578774

606

1308

11.8055

32.9315

23.36546

7.85291219

582

1174

- 0.16490

- 0.27708

- 0.21809

0.21854

- 4.0

-10.24464

11.7315

32.7145

23.25002

7.8919031

537

1081

- 0.7907

- 0.9342

- 0.71108

0.71614

- 11.38614

-17.35474

Tables 4-4 shows that the best ignition timing for camshaft no.1 at WOT and

2000 rpm should be 21º BTDC because brake thermal efficiency reduces around 0.6%

but the THC and NOx emissions reduce 4.12% and 11.165% respectively, while the

ignition timing of 18º BTDC is not recommended because the brake thermal

efficiency reduces up to 2%.

Table 4-5 recommends the ignition timing of 24º BTDC because the loss in the

efficiency is very similar to the ignition timing of 27º BTDC, while it can reduce THC

and NOx emissions 10.84% and 14.49%

For camshaft no.3, table 4-6 shows a big difference between two ignition

timings. If the ignition timing of 30º BTDC is chosen, brake thermal efficiency

reduces 0.277% while THC and NOx emissions reduce 4.00% and 10.24%

respectively. Otherwise, the brake thermal efficiency reduces 0.71% while THC and

NOx emissions reduce up to 11.39% and 17.35% respectively. However, The levels of

emissions from camshaft no.3 are relatively low. Therefore, the MBT timing (33º

BTDC) or 30º BTDC should be appropriate.

CHAPTER 5

CONCLUSIONS AND RECOMMENDATIONS

This study is one of researches on the alternative fuel aimed to improve the

pollution. This experiment is supported by Keimyung University and EROOM

Company, Republic of South Korea.

The present investigation focuses on the effects of intake valve timing and

injection timing in a natural gas diesel engine, which are defined into five parts

because the experiment was undergoing with four main parameters including the

effects of ignition timing.

Enumerating the purpose of the experiment again, which are:

1. Study of a modification of a diesel engine to a natural gas dedicated diesel

engine.

2. Developing an experimental set-up to continue research on natural gas

diesel engine at the Keimyung University and EROOM Company.

3. Investigation the effects of intake valve timing, speed, load and injection

timing on power, torque, brake thermal efficiency, volumetric efficiency, SFC, CO,

CO2, THC and NOx are discussed with the consideration based on other data such as

airflow rate, fuel flow rate, exhaust gas temperature etc.

5.1 Conclusions

This study provides results the of spark ignition natural gas diesel engine (2.2

liters, 4-stroke-4-cylinder Daedong 4A220A-S1 engine). Even though, the useful data

are in the MBT timing. The data were intended to show versus the ignition timing

between the range 15ºBTDC and 54ºBTDC. This way of presentation can illustrate a

primary overview, which leads to better understanding. Then, the MBT ignition-

timing region becomes easier to be utilized.

Note that this conclusion is based on the LIVC and the speed between 1500 and

2500 rpm. In addition, the injection timings are focusing on only 40ºBTDC, 8ºBTDC

and 103.5ºATDC. The result can be different in case of being beyond this scope.

92

The following conclusions have been reached:

1. Higher load comes with higher power, torque, volumetric efficiency, SFC,

THC, NOx and CO while brake thermal efficiency and CO2 are lower.

2. Engine speed affects the heat loss, friction loss and volumetric efficiency,

which affect output torque. Engine speed limits the combustion time, which raises the

exhaust gas temperature around 50ºC for every increment of 500 rpm.

3. Camshaft no.1, 2 and 3 can obtain the maximum volumetric efficiency of

approximately 94%, 88% and 78% at the speed of 2100 rpm, 2200 rpm and 2300 rpm

respectively, if they are operated at WOT.

4. At 25% load and the speed of 2500 rpm, camshaft no.2 can increase the

brake thermal efficiency 1.08% and reduce brake specific fuel consumption up to

4.58% comparing to the original camshaft (camshaft no.1). This condition must be

operated only with the injection timing of 103.5ºATDC.

5. For gaseous indirect injection system, the injection timing has less influence

on the engine performance than load, speed, valve timing and ignition timing.

However, appropriate injection timing can improve and/or control the efficiencies,

emissions and knock margin.

6. MBT ignition timing is not always the best choice. The ignition timing of

camshaft no.1 and 2 should be retarded around 3ºCA (Crank Angle) and 6ºCA

respectively. Since the THC and NOx emissions can decrease up to 10.84% and

14.5% respectively while camshaft no.3 should go with MBT timing to maintain high

brake thermal efficiency.

5.2 Recommendations for Future Works

This experiment is relatively rough but it shows many effects. It shows quite

clear effects of loads. Intake valve timings and speeds relate to each other. On the

other hand, the effects of injection timings seem to have the least effect among the

parameters. It does not show the overview trend but it shows that injection timing has

a specific effect in each valve timing, speed and/or load.

The future works can focus on camshaft no.1 and compare with other valve

timings around 35º ABDC, which can be 25º, 30º, 40º and 45º ABDC. The load 25%,

50% and 100% should be enough for the investigation. Speed can move closer to get

93

more detail. Especially, the injection timing should be very detailed. The experiment

should go over the entire period of valve timing. The fuel should be injected from

around 40º BTDC (before intake valve opens.) until the intake valve closure, with a

small interval. The combustion analyzer should be brought to the data collection.

Other parameters can be added to further researches. Different compression

ratios, air-fuel ratio included lean mixture or valve lift can be very interesting. The

Early Intake Valve Closure (EIVC) is another interested thing to compare with Late

Intake Valve Closure (LIVC) based on the same effective compression ratio.

This experiment tested at the same throttle position and worked on the output

data. Another way to modify a natural gas diesel engine is an engine operated with

constant load condition. An engine is tested with a specific fixed load condition (for

example 50 Nm) at each operating speed such as 1500 rpm to 3000 rpm with 500

intervals. After the engine is retrofitted, the same method is going again. This can

compare the fuel flow rate, efficiencies and emissions, which can show another view

about economics. Moreover, it may lead to show the possibility that natural gas diesel

engine can give out more power than the unmodified engine.

Biogas is another very interesting renewable energy. Many applications use

biogas and diesel (dual fuel) in compress-ignition engines for electricity generation.

Most of the engines are old, which have indirect injection system and are difficult to

control the injection timing. This is a big problem to improve the efficiencies in

engineering vision. Therefore, retrofitting from diesel engine to spark-ignition engine

with close loop control is the way that engineers should do for better efficiencies and

emissions.

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ratio and engine speed on the performance and emission characteristics of a

spark ignition engine using hydrogen as a fuel.” Renewable Energy Vol. 29

(2004) : 2245-2260.

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HarperCollinsPublishers, c1973.

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Liquid-Phase LPG Injection Engine for a Heavy-Duty Vehicle.” JSME

International Journal. Vol. 47 (2004) : 410-415.

29. J.B. Heywood, Internal Combustion Engine Fundamentals. 1st ed. Singapore .

McGraw-Hill, c1988.

APPENDIX A

NATURAL GAS PROPERTY

100

NATURAL GAS PROPERTY

NGP Version 4.54 PROGRAM OUTPUT

(HA YOUNG CHEOL, R&D CENTER, KOGAS)

Standard gas.NGP

INPUT VALUES ARE AS FOLLOWS

Ref. Temperature for Volume , Tb : 0 C

Ref. Pressure for Volume , Pb : 101.325 kPa (abs)

Flow Temperature for Volume , Tf : 0 C

Flow Pressure for Volume , Pb : 101.325 kPa (abs)

Ref.Temperature for Heating Value , Th : 15 C

Ref Pressure for Heating Value , Ph : 101.325 kPa (abs)

Conversion Factor of kcal to kJ : 4.1868

Normalizing the composition of NG : YES

------------- NG mol% ------------- ----------- NG mol % normalized ----------

CH4[%] : 90.09 90.09000

C2H6 : 6.04 6.04000

C3H8 : 2.54 2.54000

IC4H10 : 0.54 0.54000

NC4H10 : 0.58 0.58000

IC5H12 : 0.02 0.02000

N2 : 0.19 0.19000

101

CALCULATED VALUES ARE AS FOLLOWS

ACCORDING TO VARIOUS LITERATURES AND JOURNAL, FOLLOWING

VALUES ARE ACCURATELY COMPUTED

Specific Gravity : 0.6268385 (Dim.less) ISO 6976-95

Ref. Compressibility : 0.9968333 (Dim.less) ISO 6976-95

Ref. Density : 0.8104538 (kg/m^3) ISO 6976-95

Flow Density : 0.810422 (kg/m^3) AGA 8-92

Gross Heating Value : 1185.4752 (Btu/ft^3) ISO 6976-95

Gross Heating Value : 44169.5633 (kJ/m^3) ISO 6976-95

Gross Heating Value : 10549.7190 (kcal/m^3) ISO 6976-95

Gross Heating Value : 54499.7946 (kJ/kg) ISO 6976-95

Inferior Heating Value : 9532.3384 (kcal/m^3) ISO 6976-95

Inferior Heating Value : 49244,0134 (kJ/kg) ISO 6976-95

Wobbe index : 55788.5774 (kJ/m^3) ISO 6976-95

Wobbe index : 13324.8728 (kcal/m^3) ISO 6976-95

Ref. Viscosity : 0.9973E-05 (Pa.s) API TECH.

Flow Viscosity : 0.9973E-05 (Pa.s) DATA BOOK

PETROLEUM

REFINING

Ref. Isentropic Expo. : 1.29905 (Dim.less) FLUID PHASE

F low Isentropic Expo. : 1.29905 (Dim.less) EQUILIBRIA 6

Speciflc Heat, Cp : 2.06610 (J/g-K)

Speed of sound : 401.148 (m/s)

Methane & Octane Number : 72.7/121.9

APPENDIX B

EXPERIMENTAL CALCULATION

104

EXPERIMENTAL CALCULATION

This investigation defined the effects on power, torque, CO, CO2, THC, NOx,

O2, Specific Fuel Consumption (SFC), volumetric efficiency and brake thermal

efficiency. Power and torque can be gotten from the dynamometer controller. CO,

CO2, THC, NOx and O2 are from the exhaust gas analyzer. The values of SFC,

volumetric efficiency and brake thermal efficiency must be calculated from other data

as follow.

Gas flow meter, as shown in figure 3-16, gives the natural gas flow rate in term

of NCMH (Normal Cubic Meter per Hour). The volume flow rate must be:

3600)/( 3 NCMHsmV =

•

Eq.B-1

Multiplication Eq.B-1 to ref. density in Appendix A gives mass flow rate.

Eq.B-2 0.8104538*)/(••

= Vskgm

Therefore, the calculation of SFC is:

)(

)/()/(

kWPowerBrake

skgmJgSFC

•

= Eq.B-3

Whereas, 1 PS = 0.7355104 kW

Dynamometer controller also collects the data of atmospheric pressure (mmHg).

Eq.B-4 is to convert the unit to kPa.

325.101*760

)(Pr)(Pr mmHgessurekPaessure = Eq.B-4

105

The assumption of ideal gas provides the calculation of air density as follow:

)).(15.273(*287.0)(Pr

CTempAirIntakekPaessureDensityAir o+

= Eq.B-5

Air mass flow rate (kg/s) is the product of air density, Eq.B-5, and volume flow

rate (m3/s) which is given by the laminar flow element, as shown in figure 3-15.

Eq.B-6 )/(*)/()/( 33 smFlowAirmkgDensityAirskgFlowAir =

Engine specification, shown in table 3-1, is brought to the volumetric efficiency

calculation.

%100*2925.1*

60.*

4***4

)/(*2. 2 rpmstrokeboreskgFlowAirEffVol

π= Eq.B-7

Where, 1.2925 is the air density at standard condition.

%100*)/(*)/(

)(. 33 mkJHVsmFlowFuelkWOuputPowerEffThermal = Eq.B-8

Where, HV (Heating Value) equals 44169.5633, as mentioned in Appendix A.

107

BIOGRAPHY

Name : Mr. Chedthawut Poompiaptpong

Thesis Title : Effects of Intake Valve Timing and Injection Timing in a Natural Gas

Dedicated Diesel Engine

Major Field : Automotive Engineering

Biography

Mr. Chedthawut Poompipatpong graduated his Bachelor Degree of Mechanical

Engineering from Department of Mechanical Engineering, Faculty of Engineering,

King Mongkut’s Institute of Technology North Bangkok in April 2005. He continued

to the Master of Science in Automotive Engineering, The Sirindhorn International

Thai-German Graduate School of Engineering (TGGS), King Mongkut’s Institute of

Technology North Bangkok in the academic year of 2005. He attended the third and

forth semesters of academic program, industrial internship and master thesis, at

Power-Train Laboratory, Department of Mechanical and Automotive Engineering,

Keimyung University, Republic of South Korea. He was also supported by Professor

Dr. Choi Gyeung Ho, Keimyung University and EROOM Company throughout this

master thesis.

He can be reached at 73/17 Soi Rompho Chaiyapruck Rd. Talingchan Bangkok

10170 Thailand.

Documents

Effects of Intake Valve Timing and Injection Timing