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EFFECTS OF INTAKE VALVE TIMING AND INJECTION TIMING
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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
4-5 Comparison between MBT Ignition Timing and Retard Ignition Timing
for Camshaft no.2 88
4-6 Comparison between MBT Ignition Timing and Retard Ignition Timing
for Camshaft no.3 89
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
LIST OF FIGURES (CONTINUED)
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
4-36 Brake Thermal Efficiency versus Ignition Timing
2500 rpm. 25% Load at Injection Timing 8ºBTDC 72
x
LIST OF FIGURES (CONTINUED)
Figure Page
4-37 Brake Thermal Efficiency versus Ignition Timing
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.
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−=
−= 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.
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.
Camshaft no.2, intake valves start opening at 8 degree BTDC during the exhaust
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.
Camshaft no.3, intake valves start opening at 8 degree BTDC during the exhaust
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.
Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC
89
1011
12131415
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
Torq
ue (k
g m
)
25%50%100%
FIGURE 4-2 Effect of Load on Torque
Camshaft No.2; 1500 rpm. at Injection timing 40ºBTDC
22232425
26272829
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 4-3 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.
Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC
7
7.5
8
8.5
9
9.5
10
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
SFC
(10-5
g/J
)
25%50%100%
FIGURE 4-4 Effect of Load on SFC
Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC
19
20
21
22
23
24
25
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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.
Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC
50
60
70
80
90
100
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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.
Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC
0
200
400
600
800
1000
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
THC
(ppm
)
25%50%100%
FIGURE 4-7 Effect of Load on THC
55
Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC
0
500
1000
1500
2000
2500
3000
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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 .
Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC
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 (ºBTDC)
CO (%
) 25%50%100%
FIGURE 4-9 Effect of Load on CO
56
Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC
00.10.20.3
0.40.50.60.7
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
O2 (
%) 25%
50%100%
FIGURE 4-10 Effect of Load on O2
Camshaft No.2; 2500 rpm. at Injection timing 40ºBTDC
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
Ignition timing (ºBTDC)
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
Camshaft No.2; WOT at Injection timing 40ºBTDC
50
60
70
80
90
100
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
Volu
met
ric
effic
ienc
y (%
)
1500 rpm.2000 rpm.2500 rpm.
FIGURE 4-13 Effect of Speed on the Volumetric Efficiency at WOT
Camshaft No.2; WOT at Injection timing 40ºBTDC
1111.5
1212.5
1313.5
1414.5
15
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
Ignition timing (ºBTDC)
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
Camshaft No.2; 25% load. at Injection timing 8ºBTDC
99.510
10.511
11.512
12.513
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
Camshaft No.1; 25% load. at Injection timing 40ºBTDC
20
21
22
23
24
25
26
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
Camshaft No.1; WOT at Injection timing 40ºBTDC
20
21
22
23
24
25
26
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
Camshaft No.3; WOT at Injection timing 40ºBTDC
2020.5
2121.5
2222.5
2323.5
24
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
Camshaft No.1; 25% load. at Injection timing 40ºBTDC
60
65
70
75
80
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
Camshaft No.1; WOT at Injection timing 40ºBTDC
80
85
90
95
100
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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.
Camshaft No.1; 25% load. at Injection timing 40ºBTDC
0
200
400
600
800
1000
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
THC
(ppm
)
1500 rpm.2000 rpm.2500 rpm.
FIGURE 4-22 Effect of Speed on THC Emission
Camshaft No.1; 25% load. at Injection timing 40ºBTDC
0
500
1000
1500
2000
2500
3000
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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.
Camshaft No.3; 25% load. at Injection timing 40ºBTDC
0
500
1000
1500
2000
2500
3000
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
Ignition timing (ºBTDC)
Pow
er (P
s) Camshaft No.1Camshaft No.2Camshaft No.3
FIGURE 4-25 Effect of Intake Valve Timing on the Power Output
2000 rpm. WOT at Injection timing 40ºBTDC
89
10111213141516
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
2000 rpm. WOT at Injection timing 40ºBTDC
6
7
8
9
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
SFC
(10
-5 g/
J)
Camshaft No.1Camshaft No.2Camshaft No.3
FIGURE 4-27 Effect of Intake Valve Timing on the SFC
2000 rpm. WOT at Injection timing 40ºBTDC
20
21
22
23
24
25
26
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
Brak
e Th
erm
al E
ffici
ency
(%
)
Camshaft No.1Camshaft No.2Camshaft No.3
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.
2000 rpm. WOT at Injection timing 40ºBTDC
0
200
400
600
800
1000
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
THC
(ppm
)
Camshaft No.1Camshaft No.2Camshaft No.3
FIGURE 4-30 Effect of Intake Valve Timing on the THC Emission
2000 rpm. WOT at Injection timing 40ºBTDC
0
500
1000
1500
2000
2500
3000
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
NO
x (p
pm)
Camshaft No.1Camshaft No.2Camshaft No.3
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
2000 rpm. WOT at Injection timing 40ºBTDC
0
0.5
1
1.5
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
Ignition timing (ºBTDC)
CO2
(%) Camshaft No.1
Camshaft No.2Camshaft No.3
FIGURE 4-33 Effect of Intake Valve Timing on the CO2 at 1500 rpm.
70
2000 rpm. WOT at Injection timing 40ºBTDC
10.310.410.510.610.710.810.9
1111.1
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
CO2
(%) Camshaft No.1
Camshaft No.2Camshaft No.3
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
2500 rpm. 25% load at Injection timing 40ºBTDC
18
19
20
21
22
23
24
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
Brak
e Th
erm
al E
ffici
ency
(%
)
Camshaft No.1Camshaft No.2Camshaft No.3
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
2500 rpm. 25% load at Injection timing 8ºBTDC
18
19
20
21
22
23
24
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
Brak
e Th
erm
al E
ffici
ency
(%
)
Camshaft No.1Camshaft No.2Camshaft No.3
FIGURE 4-36 Brake Thermal Efficiency versus Ignition Timing
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
Ignition timing (ºBTDC)
Brak
e Th
erm
al E
ffici
ency
(%
)
Camshaft No.1Camshaft No.2Camshaft No.3
FIGURE 4-37 Brake Thermal Efficiency versus Ignition Timing
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
Ignition timing (ºBTDC)
Pow
er (P
s) 40ºBTDC8ºBTDC103.5ºATDC
FIGURE 4-38 Effect of Injection Timing on Power
74
Camshaft No.2; 2000 rpm. at 50% load
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.
Camshaft No.2; 2000 rpm. at 50% load
60
70
80
90
100
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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.
Camshaft No.2; 2000 rpm. at 50% load
7
7.5
8
8.5
9
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
SFC
(10-5
g/J
)
40ºBTDC8ºBTDC103.5ºATDC
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
Ignition timing (ºBTDC)
Brak
e Th
erm
al E
ffici
ency
(%
)
40ºBTDC8ºBTDC103.5ºATDC
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.
Camshaft No.3; 2000 rpm. at 50% load
0
200
400
600
800
1000
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
THC
(ppm
)
40ºBTDC8ºBTDC103.5ºATDC
FIGURE 4-43 Effect of Injection Timing on THC
77
Camshaft No.3; 2000 rpm. at 50% load
0
500
1000
1500
2000
2500
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
NO
x (p
pm)
40ºBTDC8ºBTDC103.5ºATDC
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
Camshaft No.3; 1500 rpm. at 25% load
0500
10001500
2000250030003500
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
THC
(ppm
)
40ºBTDC8ºBTDC103.5ºATDC
FIGURE 4-45 Knocking at 1500 rpm in Camshaft No.3
Camshaft No.3; 2000 rpm. at 25% load
200
400
600
800
1000
1200
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
THC
(ppm
)
40ºBTDC8ºBTDC103.5ºATDC
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
Camshaft No.2; 1500 rpm. at 25% load
10.4
10.5
10.6
10.7
10.8
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
CO
2
40ºBTDC8ºBTDC103.5ºATDC
FIGURE 4-47 CO2 at 1500 rpm 25% Load for Camshaft No.2
Camshaft No.2; 1500 rpm. at 50% load
10.4
10.5
10.6
10.7
10.8
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
CO2
40ºBTDC8ºBTDC103.5ºATDC
FIGURE 4-48 CO2 at 1500 rpm 50% Load for Camshaft No.2
80
Camshaft No.2; 1500 rpm. at 25% load
0
0.2
0.4
0.6
0.8
1
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
CO (%
) 40ºBTDC8ºBTDC103.5ºATDC
FIGURE 4-49 CO at 1500 rpm 25% Load for Camshaft No.2
Camshaft No.2; 1500 rpm. at 50% load
0
0.2
0.4
0.6
0.8
1
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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
1500 rpm. 25% load at Injection timing 40ºBTDC
10.110.210.310.4
10.510.610.710.8
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
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.2 24º- 30º 30º- 33º 33º- 42º
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
Camshaft No.2 24º- 27º 27º- 30º 33º- 36º
Camshaft No.3 30º- 33º 36º- 39º 45º- 48º
84
TABLE 4-3 MBT Timing at WOT
WOT load 1500 rpm. 2000 rpm. 2500 rpm.
Camshaft No.1 21º- 24º 24º- 27º 30º- 33º
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
2000 rpm. WOT at Injection timing 40ºBTDC
89
10111213141516
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timi
Torq
ue (k
g m
)
Camshaft No.1Camshaft No.2Camshaft No.3
ng (ºBTDC)
MBT timing for camshaft no.1
MBT timing for camshaft no.2
MBT timing for camshaft no.3
FIGURE 4-54 MBT at 2000 rpm and WOT versus Torque
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
Ignition timing (ºBTDC)
Pow
er (P
s) Camshaft No.1Camshaft No.2Camshaft No.3
MBT timing for camshaft no.1
MBT timing for camshaft no.2
MBT timing for camshaft no.3
FIGURE 4-55 MBT at 2000 rpm and WOT versus Power
86
2000 rpm. WOT at Injection timing 40ºBTDC
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)
Camshaft No.1Camshaft No.2Camshaft No.3
ng ( C)
MBT timing for camshaft no.1
MBT timing for camshaft no.2
MBT timing for camshaft no.3
FIGURE 4-56 MBT at 2000 rpm and WOT versus SFC
2000 rpm. WOT at Injection timing 40ºBTDC
20
21
22
23
24
25
26
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
Brak
e Th
erm
al E
ffici
ency
(%
)
Camshaft No.1Camshaft No.2Camshaft No.3
MBT timing for camshaft no.1
MBT timing for camshaft no.2
MBT timing for camshaft no.3
FIGURE 4-57 MBT at 2000 rpm and WOT versus Brake Thermal Efficiency
87
2000 rpm. WOT at Injection timing 40ºBTDC
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
)
Camshaft No.2Camshaft No.3
ing (ºBTDC)
MBT timing for camshaft no.1
MBT timing for camshaft no.2
MBT timing for camshaft no.3
FIGURE 4-58 MBT at 2000 rpm and WOT versus THC
2000 rpm. WOT at Injection timing 40ºBTDC
0
500
1000
1500
2000
2500
3000
15 18 21 24 27 30 33 36 39 42 45 48 51 54
Ignition timing (ºBTDC)
NO
x (p
pm)
Camshaft No.1Camshaft No.2Camshaft No.3
MBT timing for camshaft no.1
MBT timing for camshaft no.2
MBT timing for camshaft no.3
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
TABLE 4-5 Comparison between MBT Ignition Timing and Retard Ignition Timing
for Camshaft no.2
Camshaft no.2 30º BTDC (MBT timing)
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
TABLE 4-6 Comparison between MBT Ignition Timing and Retard Ignition Timing
for Camshaft no.3
Camshaft no.3 33º BTDC (MBT timing)
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.
REFERENCES
1. PTT Public Company Limited. Natural Gas Road Map. [Online] 2005. [cited
15 Dec. 2006]. Available from : URL : http://www/pttplc.com/th/default.asp
2. Yusoff, A., Zailani, M. and Muthana, I. K. Valve Timing and Ignition Issues in
fuel system for Compressed Natural Gas Direct Injection (CNGDI).
Faculty of Engineering, Universiti Kebangsaan Malaysia.
3. M. A. Kalam, et al. Power Improvement of a Modified Natural Gas Engine.
Department of Mechanical Engineering, University of Malaya.
4. Effect of advanced injection timing on the performance of natural gas in diesel
engines. India . Sadhana Vol. 25 (Feb. 2000) : 11-20.
5. Michael K. A. et al. “First and Second Law Analyses of a Natural-Aspirated,
Miller Cycle, SI Engine with Late Intake Valve Closure.” SAE International
Paper 980889. (1998) : 1-16.
6. Yorihiro F., et al. “Development of High Efficiency Miller Cycle Gas Engine.”
Mitsubishi Heavy Industry, Ltd. Technical Review. Vol. 38 No.3
(Oct. 2001) : 146-150.
7. M. D. Basset, et al. “A simple Two-State Late Intake Valve Closing Mechanism.”
Proc. Instn. Mech. Engrs. Vol. 211 (1997) : 237-241.
8. S. Shiga, et al. “Effect of Over-Expansion Cycle in a Spark-Ignition Engine using
Late-Closing of Intake Valve and Its Thermodynamic Consideration of the
Mechanism.” International Journal of Automotive Technology. Vol. 2 No.1
(2001) : 1-7.
9. Chih Wu, et al. “Performance Analysis and Optimization of a Supercharged
Miller Cycle Otto Engine.” Applied Thermal Engineering. 23 (2003) :
511-521.
10. Gyeung H. C., et al. “An experimental and numerical study of a miller cycle for
gas engine converted from a diesel engine.” ASME/IEEE Joint Rail
Conference & Internal Combustion Engine Spring Technical Conference.
(March 2007) : 1-6.
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11. Y. Wang. et al. “Experimental investigation of applying Miller cycle to reduce
NOx emission from diesel engine.” Proc. IMechE. Vol.219 (2005) :
631-638.
12. G.H. Abd Alla, et al. “Effects of injection timing on the performance of a dual
fuel engine.” Energy Conversion&Management. Vol. 43 (2002) : 269-277.
13. Takagaki S. “The effects of compression ratio on nitric oxide and hydrocarbon
emissions from a spark-ignition natural gas fuelled engine.”
SAE paper 970506. (Feb. 1997) .
14. Koichi H. et al. “A study of the improvement effect of Miller-cycle on mean
effective pressure limit for high-pressure supercharged gasoline engines.”
JSAE. 18 (1997) : 101-106.
15. Jerald A. C. “The Use of a Three-Zone Combustion Model to Determine Nitric
Oxide Emissions from a Homogeneous-Charge, Spark-Ignited Engine.”
2003 Spring Technical Conference. (11-14 May 2003) : 1-12.
16. J. A. Caton. “Effects of the compression ratio on nitric oxide emissions for a
spark ignition engine : results from a thermodynamic cycle simulation.”
Int. J. Engine Res.. Vol. 4 No. 4 (2003) : 249-268.
17. Tsukida et al. “Production Miller-Cycle Natural Gas Engine.”
Inter- Tech Energy Progress, Inc.. (1999) : 1-9.
18. A. Al-Sarki. et al. “Efficiency of a Miller Engine.” Applied Energy.
(2005) : 1-9.
19. Akira T. et al. “Mitsubishi Lean-Burn Gas Engine with World’s Highest Thermal
Efficiency.” Mitsubishi Heavy Industry, Ltd. Technical Review.” Vol. 40
No.4 (Aug. 2003) : 1-6.
20. Mohamed, M. B. Investigation of the Performance of a Spark Ignition Engine
with Gaseous Fuels. Master Thesis, Faculty of Engineering, Dalhousie
University, 1998.
21. Lee J. H. A Study of the Thermal Efficiency on the Industrial Engine with Miller
Cycle. Master Thesis, Department of Automotive Engineering, Graduate
School, Keimyung University, 2006.
97
22. Yunas A.C. and Michael A.B. Thermodynamics. 4th ed. Singapore :
McGraw-Hill, c2002.
23. Willard, W. Engineering Fundamentals of the Internal Combustion Engine.
2nd ed. USA . Prentice Hall, c2004.
24. Richard, S. Introduction to Internal Combustion Engines. 3rd ed. Great
Britain . SAE, c1999.
25. Andrew J. K., et al. “Effects of Vehicle Speed and Engine Load on Motor
Vehicle Emissions.” Environmental Science&Technology. Vol. 37 (2003) :
3739-3746.
26. Maher A.R. and Sadiq Al-Baghdadi. “Effect of compression ratio, equivalence
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.
27. F. O. Edward, Internal Combustion Engine and Air Pollution. 1st ed. USA .
HarperCollinsPublishers, c1973.
28. Yonggyu LEE, et al. “Effects of Injection Timing on Mixture Distribution in a
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.
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