experimental investigation on reduction of emission in hcci engine

experimental investigation on reduction of emission in hcci engine

EXPERIMENTAL INVESTIGATION ON

REDUCTION OF EMISSION IN HCCI ENGINE

BY PCCI MODE WITH DIFFERENT FUELS

A THESIS

Submitted by

S. MOHANAMURUGAN

for the award of the degree of

DOCTOR OF PHILOSOPHY

DEPARTMENT OF MECHANICAL ENGINEERING

DR. M.G.R EDUCATIONAL AND RESEARCH INSTITUTE

(Declared u/s 3 of the UGC Act 1956)

CHENNAI – 600 095

APRIL 2010

ii

BONAFIDE CERTIFICATE

Certified that this thesis titled

“EXPERIMENTAL

INVESTIGATION ON REDUCTION OF EMISSION IN HCCI

ENGINE BY PCCI MODE WITH DIFFERENT FUELS” is the bonafide

work of Mr. S. MOHANAMURUGAN who carried out the research under my supervision. Certified further that to the best of my knowledge the work reported herein does not form part of any other thesis or dissertation on the basis of which a degree or award was conferred on an earlier occasion on this or any other candidate.

Place:

Date :

Signature of the Supervisor

Dr.S. SENDILVELAN

Principal

Aksheyaa College of Engineering

Kancheepuram District, India

iii

DECLARATION

I declare that the thesis entitled

“EXPERIMENTAL

INVESTIGATION ON REDUCTION OF EMISSION IN HCCI

ENGINE BY PCCI MODE WITH DIFFERENT FUELS” submitted by

me for the degree of Doctor of Philosophy is the record of work carried out by me during the period from November 2005 to April 2010 under the guidance of Dr. S. SENDILVELAN and has not formed the basis for the award of any degree, diploma, associate-ship, fellowship, titles in this or any other university or other similar institution of higher learning.

Signature of the candidate

iv

ABSTRACT

The internal combustion engine is the key to the modern society.

Without the transportation performed by the millions of vehicles on road and at sea we would not have reached today’s living standard. There are two types of internal combustion engines, namely, the Spark Ignition (SI), and the

Compression Ignition (CI). Petrol and diesel are at present the principal fuels for SI and CI engines, respectively. These fuels are on the verge of getting extinct, and during combustion, these fuels release substantial amount of pollutants into the atmosphere and create environmental related problems.

The internal combustion (IC) engine is known to be one of the major sources of air pollution in the environment. The fuel oxidation process in the engine generates not only useful power, but also a considerable amount of pollutant emissions including Carbon Dioxide (CO

2

), Carbon Monoxide (CO),

Unburned Hydrocarbon (HC), Nitrogen Oxides (NOx), and Particulate Matter

(PM). CO

2

is mainly responsible for the global warming issue as it creates a reflective layer in the atmosphere that reflects heat from the earth back to the earth surface, increasing the earth’s average temperature over time. CO is a very dangerous substance, since it reduces the oxygen-carrying capacity of the blood stream. At low concentrations, CO inhalation can cause dizziness and nausea, while at higher concentrations it can be deadly. Unburned hydrocarbon emission, a result of an incomplete combustion process, is a common source of respiratory problems. Particulate emissions or soot also

v cause some respiratory problems. Both unburned hydrocarbon and soot emissions have been linked to diseases like cancer. The high flame temperature generated during the combustion process is responsible for NOx formation, which causes various health problems in addition to contributing to acid rain and global warming issues. The advent of stringent emission norms and depletion of fossil fuel resources led engineers to work out new combustion technologies to substantially reduce harmful emission and improve the overall efficiency of an IC Engine. The factors to be considered while designing a new combustion process are, higher compression ratio, lean homogeneous air fuel mixture, complete and instantaneous combustion, which lead to Homogeneous Charge Compression Ignition (HCCI). HCCI is a clean and efficient combustion process. In this research work an attempt is made to experimentally analyze the performance and emission characteristics of the HCCI compression process in a Premixed Charge Compression Ignition

(PCCI) mode assisted with Pilot Injection (PI) as the combustion initiator.

Several experiments were conducted in a modified single cylinder watercooled diesel engine, employing a conceptual system known as Transient

State Fuel Induction (TSFI) with different fuels such as diesel, petrol, and biodiesel. In the present research setup, it is observed that there is a reduction in the emission level of CO and HC, with the same power as obtained from a conventional diesel engine.

vi

ACKNOWLEDGEMENT

I express my deep sense of gratitude and indebtedness to my supervisor Dr. S. Sendilvelan, Principal, Aksheyaa College of Engineering,

Kancheepuram District, India, for his constant encouragement, help and valuable guidance in carrying out the research work.

I am grateful to my Doctoral Committee members,

Dr. N. Padmamabhan,

Professor and Head, Department of Chemical

Engineering, Dr.M.G.R Educational and Research Institute, Chennai, and

Dr. K. Balagurunathan, Dean-Academic, St. Peter’s Engineering College,

Chennai, for providing the necessary support, review and guidance.

I sincerely thank Mr. M. Ganesan, Professor and Head, Department of

Mechanical Engineering, Dr.M.G.R Educational and Research Institute,

Chennai, Dr. P. Aravindan, Dean (IIPC & Research), Dr.M.G.R Educational and Research Institute, Chennai, and Principal, Velammal Engineering

College, Chennai, for their kind support and timely help.

Finally, I wish to thank all those who helped me directly or indirectly in completing this research work.

S. MOHANAMURUGAN

vii

TABLE OF CONTENTS

CHAPTER NO.

ABSTRACT

LIST OF TABLES

TITLE PAGE NO.

LIST OF FIGURES

LIST OF SYMBOLS AND ABBREVIATIONS

iv

xi

xii xv

1 INTRODUCTION

1.1 SPARK IGNITION ENGINE

1.2 COMPRESSION IGNITION ENGINE

1.3 HOMOGENEOUS CHARGE COMPRESSION

IGNITION ENGINE

1.3.1

1.3.2

HCCI Combustion Chemistry

Ignition Resistance of Diesel Fuel

1

3

4

1.3.3

1.3.4

1.3.5

1.3.6

1.3.7

Injection System

Premixed HCCI

Early Direct Injection HCCI

Late Direct Injection HCCI

Variable Valve Timing

1.4 BIO-DIESEL

1.4.1 Bio-diesel in Diesel Engine

1.5 ORGANIZATION OF THE THESIS

12

12

13

14

15

5

10

11

15

16

17

2 LITERATURE REVIEW

2.1 INTRODUCTION

19

19

viii

CHAPTER NO. TITLE

2.2 HOMOGENEOUS CHARGE

2.3

2.4

COMPRESSION IGNITION

IC ENGINE EMISSIONS

PAGE NO.

NON-ROAD DIESEL ENGINE EMISSION

20

23

2.5

STANDARDS

SETTING THE STANDARD

2.5.1 Stage I/II Standards

2.5.2 Stage III/IV Standards

2.6 ENGINE PERFORMANCE

2.7 EXHAUST EMISSIONS

2.8 CLOSURE

2.9 MOTIVATION AND OBJECTIVE OF

THIS WORK

2.10 SCOPE OF THE STUDY

67

67

33

36

38

66

30

32

33

3 EXPERIMENTATION

3.1 INTRODUCTION

3.2 EXPERIMENTAL APPARATUS

3.3 SAMPLE FUELS

3.4 INJECTOR DETAILS

3.5 INJECTION TIMING

3.6 BACKGROUND OF THE ENGINE

3.6.1 General Characteristics of DI Single

3.6.2

Cylinder Research Engine

Functioning Mode

3.7 EXPERIMENTAL PROCEDURE

3.7.1 Mode 1 - Diesel Conventional

3.7.2 Mode 2 - Bio-diesel Conventional

73

73

75

76

68

68

68

76

76

77

78

79

ix

CHAPTER NO. TITLE

3.7.3 Mode 3 - Diesel - Diesel

3.7.4

PCCI-DI Combustion

Mode 4 - Bio-diesel- Bio-diesel

PCCI-DI Combustion

PAGE NO.

79

79

4

3.7.5 Mode 5 - Diesel - Bio-diesel

PCCI-DI Combustion

3.8 EXHAUST GAS ANALYSER

3.8.1 Details of the Exhaust Gas Analyzer

79

80

81

RESULTS AND DISCUSSION

4.1 INTRODUCTION

4.2 EFFECT OF THE LOAD ON EXHAUST

GAS TEMPERATURE

4.3 EFFECT OF LOAD ON HYDROCARBON

IN EXHAUST

4.4 EFFECT OF LOAD ON NOx EMISSION

IN EXHAUST GAS

4.5 EFFECT OF LOAD ON SPECIFIC FUEL

84

84

85

90

93

CONSUMPTION

4.6 EFFECT OF LOAD ON BRAKE

THERMAL EFFICIENCY

97

99

4.7 EFFECT OF LOAD ON CARBON DIOXIDE 103

4.8 EFFECT OF LOAD ON CARBON

MONOXIDE EMISSION 106

5 CONCLUSIONS

5.1 FINDINGS FROM THE RESEARCH

5.2 CONTRIBUTIONS

109

109

110

x

CHAPTER NO. TITLE

5.3 LIMITATIONS

5.4 SCOPE FOR FUTURE RESEARCH

5.5 CONCLUSION

REFERENCES

LIST OF PUBLICATIONS

VITAE

PAGE NO.

111

111

112

xi

LIST OF TABLES

TABLE NO.

2.1

2.2

TITLE

Stage I/II Emission Standards for Non Road

Diesel Engines

Stage III A Standards for Non Road Engines

PAGE NO.

33

34

2.3

2.4

3.1

3.2

3.3

4.1

Stage III B Standards for Non Road Engines

Stage IV Standards for Non Road Engines

Specifications of the experimental setup

Properties of diesel, petrol and bio-diesel

Specifications of Exhaust Gas Analyzer

The variation of the Exhaust gas temperature with an

34

34

69

73

81

4.2

4.3

4.4 increase of load

The variation of Exhaust gas temperature with an increase of load

Variation of HC emission for different loads with different fuels as input

Variation of HC (ppm) emission with different loads with different fuels as input

86

88

91

92

xii

LIST OF FIGURES

TITLE FIGURE NO.

3.1

3.2

3.3

Schematic diagram of the experimental setup

Valve timing diagram

Photographic view of the experimental setup

PAGE NO.

70

71

72

3.4

3.5

Injector assembly and components of the primary fuel injector

Photographic view of the Primary fuel injector

74

74

75 3.6

3.7

Photographic View of the Pilot Fuel Injector

Photographic view of the exhaust gas analyzer showing a sample result

3.8

4.1

4.2

Photographic view of the exhaust gas analyzer

Comparison of the conventional and HCCI methods

Comparison between different fuels in the

80

81

85

4.3

4.4

4.5

4.6

4.7

HCCI mode

Variation of the exhaust gas temperature with different fuels as input

Variation of the exhaust gas temperature with load with different fuels as input

Comparison of hydrocarbon emission between the conventional and HCCI methods

Comparison of hydrocarbon emission between

87

89

89

90 different fuels in the HCCI mode 92

Variation of hydro carbon with different fuels as input 93

xiii

FIGURE NO.

4.8

TITLE

Comparison of NOx emission between the

4.9 conventional and HCCI methods

Comparison of NOx emission between different fuels in the HCCI mode

PAGE NO.

94

95

4.10 Variation of NOx Emission with different fuels as input

4.11 Comparison of NOx emission between different

4.12

4.13 loads with different HCCI modes

Comparison of specific fuel consumption between the conventional and HCCI methods

Variation of specific fuel consumption with different fuels as input

96

96

97

98

4.14

4.15

4.16

Comparison of specific fuel consumption with different loads in different HCCI modes

Comparison of brake thermal efficiency between the conventional and HCCI methods

Comparison of brake thermal efficiency between different fuels in the HCCI mode

4.17

4.18

Brake thermal efficiency under different load conditions for all modes of operations

Variation of Brake Thermal Efficiency with

4.19

4.21 different fuels as input

Comparison of Brake Thermal Efficiency under different loads in different HCCI modes

4.20 Comparison CO

2 between the conventional and

HCCI methods

Comparison of CO

2

emission between different fuels in the HCCI mode

99

100

100

101

102

103

103

104

xiv

FIGURE NO. TITLE PAGE NO.

4.22 Variation of CO

2

emission with different fuels as input

4.23 Comparison of Brake Thermal Efficiency under

105

105

4.24

4.25

4.26 different loads in different HCCI modes

Comparison of CO emission between the conventional and HCCI methods

Comparison of CO emission between different fuels in the HCCI mode

Variation of CO emission with different fuels as input

106

107

107

LIST OF SYMBOLS AND ABBREVIATIONS

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

Z c

°

°C

P i

N g hp

Symbols

bhp-hr

CO

2

CO cm

H

2

O

2

K kg kg/h kW kWh lit m f

MJ/kg

Mpa ml/min

Brake horse power hours

Carbon Dioxide

Carbon Monoxide

Centimeter

Combustion efficiency

Degree

Degree Centigrade

Effect of injection pressure

Engine speed

Gram

Horsepower

Hydrogen peroxide

Kelvin

Kilogram

Kilogram/ hour

Kilowatts

Kilowatts hour

Liter

Mass fuel consumption

Mega Joule/Kilogram

Mega Pascal

Milli liter per minute xv

CFD

CI

CNG

CSO

DBE

DI

DME

BDC

BSEC

BSFC

BSU

BTDC

CAI mm min

NOx

%

% H rpm s

SO

2

HC

VA

-

-

Abbreviations

ANN -

ARB

ATDC

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

Millimeter

Minute

Nitrogen Oxides

Percentage

Percentage of Humidity

Revolution per minute

Second

Sulphur di-oxide

Unburned Hydrocarbon

Volt Ampere

Artificial Neural Networks

Air Resources Board

After Top Dead Center

Bottom Dead Center

Brake Specific Energy Consumption

Brake Specific Fuel Consumption

Bosch Smoke Units

Before Top Dead Center

Controlled Auto-Ignition

Computational Fluid Dynamics

Compression Ignition

Compressed Natural Gas

Cotton Seed Oil

Diesel, Bio-diesel and Ethanol mixture

Direct Injection

Dimethyl Ether xvi

EGR

EPA

EU

PM

PPM

RON

SI

SOC

SOP

NRTC

NSPS

NVO

OH

PCCI

PFI

HCCI

HDD

HHV

IC engine

IMEP

JBD

JSR

LTC

MK

MTBE

NMHC

NRSC

TDC

THC

TxLED

UHC

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

-

Exhaust Gas Recirculation

Environmental Protection Agency

European Union

Homogenous Charge Compression Ignition

Heavy-Duty Diesel Engine

Higher Heating Value

Internal Combustion engine

Indicate Mean Effective Pressure

Jatropha based Bio-Diesel

Jet-Stirred Reactor

Low Temperature Combustion

Modulated Kinetics

Methyl-tert-Butyl Ether

Non-Methane Hydrocarbon

Non-Road Steady Cycle

Non-Road Transient Cycle

New Source Performance Standard

Negative Valve Overlap

Oxygen and Hydrogen

Premixed Charge Compression Ignition

Port Fuel Injection

Particulate Matter

Parts Per Million

Research Octane Number

Spark Ignition

Start of Combustion

Standard Operating Procedures

Top Dead Center

Total Hydro Carbons

Texas Low Emission Diesel

Unburned Hydro Carbons xvii

UV

VCR

VVT

-

-

-

Ultraviolet

Variable Compression Ratio

Variable Valve Timing xviii

xix

CHAPTER 1

INTRODUCTION

In today’s world, the Internal Combustion (IC) engine is the key to the entire transportation sector. Without the transportation performed by the millions of vehicles on road and at sea, we would not have reached modern living standards. The Spark Ignition (SI) and the Compression Ignition (CI) are the two types of IC engines. Petrol and diesel are at present the principal fuels used for SI and CI engines respectively. These fuels are on the verge of getting extinct and during combustion these fuels release a substantial amount of pollutants into the atmosphere and create environment related problems.

The IC engine is known as one of the major sources of air pollutants in the environment. The fuel oxidation process in the engine generates not only useful power, but also a considerable amount of pollutant emissions including

Carbon Dioxide (CO

2

), Carbon Monoxide (CO), Unburned Hydrocarbon

(HC), Nitrogen Oxides (NOx), and Particulate Matter (PM). Reducing the exhaust emissions and increasing the fuel economy of IC engines are of global importance.

CO

2

is mainly responsible for the global warming issue, as it creates a reflective layer in the atmosphere that reflects heat from the earth back to the earth’s surface, increasing the earth’s average temperature over time. CO is a very dangerous substance since it reduces the oxygen-carrying capacity of the blood stream. At low concentrations, CO inhalation can cause dizziness and nausea, while at higher concentrations, it can be deadly (Shre et al 1998). Unburned hydrocarbon emission, a result of an incomplete

xx combustion process, is a common source of respiratory problems. Particulate emissions or soot also cause some respiratory problems. Both unburned hydrocarbon and soot emissions have been linked to diseases like cancer

(Shre et al 1998). The high flame temperature generated during the combustion process is responsible for NOx formation which causes various health problems and also contributes to acid rain and global warming issues.

The development of efficient IC engines with ultra low emissions is necessitated by strict regulations on exhaust gas composition and fuel economy. Increasing concern over the potential global warming effects of major greenhouse gases from current fossil fuels, coupled with a rapidly growing vehicle fleet around the world has intensified the uptake of alternative fuels and become an important area of research (Yap et al 2005).

Bio diesel is capable of solving the problem of fuel supply in a decentralized fashion and simultaneously reducing environment related problems. Jatropha oil has been recognized as a major source to augment declining fuel resources which could be used in IC engines. Although the burning of bio-diesel produces CO

2

emissions similar to those from ordinary fossil fuels, the plant feedstock used in the production absorbs CO

2

from the atmosphere when it grows. Plants absorb CO

2

through a process known as photo-synthesis which allows them to store energy from sunlight in the form of sugars and starches. After the biomass is converted into bio-diesel and burnt as fuel, the energy and carbon are released again. Some of that energy can be used to power an engine while the CO

2

is released back into the atmosphere.

Parallel to this interest in alternative fuels, there has also been an increased attention in Homogenous Charge Compression Ignition (HCCI) technology, incorporating the advantages of both spark ignition and compression ignition which is a potential candidate for future ultra low-

xxi emission engine strategies. HCCI engines are being actively developed because they have the potential to be highly efficient and to produce low emissions. HCCI engines can have efficiencies close to those of diesel engines, with low levels of emissions of NOx and PM. In addition, HCCI engines have been shown to operate with a range of fuels, e.g. natural gas, gasoline and bio-ethanol (Christensen et al 1997).

1.1 SPARK IGNITION ENGINE

In an SI engine, fuel is introduced into the intake system upstream of the cylinder to produce a homogeneous air-fuel mixture. The mixture is inducted into the cylinder during the intake valve open period and compressed by the piston to high pressure and temperature. At a certain time close to the top dead center, the mixture is ignited by an electrical discharge across the spark plug electrode. The flame then propagates across the cylinder volume to completely oxidize the fuel inside the cylinder. The combustion process generates high pressure which then pushes the piston downward and rotates the crankshaft. The generated power can be used to move a vehicle or other applications.

Load control in an SI engine is provided by means of a throttle which adjusts the opening of the intake port, thus controlling the mass flow rate of the air-fuel mixture entering the cylinder. Throttling creates a pressure drop in the intake system which reduces the peak pressure attainable in the cylinder, thus limiting the amount of power produced by the engine. The advantage of the SI engine is that the ignition timing is easily controlled for optimum performance by varying the spark timing. The SI engine also has a high power density under naturally aspirated conditions due to the stoichiometric nature of the mixture. In addition, the homogeneous mixture produces only a negligible amount of particulate emission or soot.

xxii

In order to ensure the flame propagation inside the cylinder and to facilitate a catalytic clean-up of the exhaust, the air-fuel ratio of the mixture must be maintained at, or close to, stoichiometric conditions. Flame propagation cannot occur if the mixture is too lean or too rich. Stoichiometric air-fuel mixtures burn at very high temperature, producing significant NOx emissions. Part of the homogeneous air-fuel mixture also gets trapped in the relatively cold crevice volume and does not burn. This unburned fraction leaves the engine as hydrocarbon emissions. The main challenge in the development of an SI engine is controlling the amount of NOx, CO and HC emissions to meet the strict regulations. The maximum compression ratio of an SI engine is limited by knocking constraints. If the air-fuel mixture is compressed to a certain pressure and temperature, it can auto-ignite prior to the spark timing, creating a high pressure fluctuation inside the cylinder. The auto-ignition process creates unpleasant pinging noise and can cause severe mechanical damage to the engine. To prevent potential damage from this auto-ignition behavior, modern engines are equipped with a knock sensor that retards the ignition timing appropriately when the knock is detected. This method may be very useful in preventing engine damage, and retarding the spark timing, but it also reduces the engine efficiency.

1.2 COMPRESSION IGNITION ENGINE

In a CI or diesel engine, air is inducted into the cylinder and compressed by the piston during the compression stroke. Fuel is injected directly into the cylinder when the piston is close to the Top Dead Center

(TDC), creating a locally rich or stoichiometric region that can auto-ignite and initiate the combustion process. Further, oxygen and fuel mixing occurs in the cylinder as the fuel oxidation process continues until all the fuel is consumed or the cylinder temperature is reduced to a point where the reactions can no longer take place.

xxiii

In a diesel engine, load control is provided by adjusting the amount of fuel injected into the cylinder in each cycle. The elimination of the throttle in the intake system increases the efficiency of diesel engine compared to the

SI engine. In addition, diesel engines typically have higher compression ratios to achieve auto-ignition, and are not limited by knock compared to SI engines; a naturally aspirated diesel engine has a lower power density because of the leaner overall air-fuel ratio and lower maximum engine speed. The locally rich or stoichiometric region in the cylinder also burns at very high temperature and produces a significant amount of NOx emissions. The presence of locally rich regions is the source of soot formation. The simultaneous reduction of NOx and soot emissions becomes the main challenge for the development of a diesel engine to meet the strict emissions regulations.

1.3 HOMOGENEOUS CHARGE COMPRESSION IGNITION

ENGINE

Engine emissions remain a critical issue affecting the design and operation of IC engines while energy conservation is becoming increasingly important. One approach to favorably address these issues is to achieve a homogeneous charge combustion at lower peak temperatures with a high compression ratio which is the ultimate goal of Homogeneous Charge

Compression Ignition (HCCI) engines.

During the last few years, the diesel market has increased rapidly, covering 49% of the total vehicles sold in Europe at the end of 2005. This increase is probably linked to the good fuel economy provided by the diesel engine. However, other factors can also explain this progress, such as the increase of power since the turbocharger has been generally available on the diesel engine, which usually works in parallel with a high pressure common rail that allows noise-reduction as well.

xxiv

Nevertheless, diesel vehicles were probably higher polluters than gasoline vehicles in the past. Fortunately, several improvements have been made to reduce the tailpipe emissions from these engines. One of these is the recent particulate trap, which allows the particulates to drop to nearly zero; the older NOx catalyst is still used in a vehicle exhaust gas system even today. These reductions in tailpipe emissions have been efficient enough to meet the Euro IV regulations: nevertheless they are, for the moment, not adapted to meet the new 2010 Euro-V regulations. As a result, some investigations have been performed over the last few years, and one of them, dealing with the HCCI theory, revealed that the usual diesel trade-off limit could be overtaken; this emerges to be a new research area.

However, the difficulty lies in controlling the operation of an HCCI engine over a wide range of speeds and loads. This is probably the most difficult hurdle faced by HCCI engines. HCCI ignition is determined by the charge mixture composition, its time temperature history, and to a lesser extent, pressure. Several potential control methods have been proposed to control HCCI combustion: varying the amount of Exhaust Gas Recirculation

(EGR), using a Variable Compression Ratio (VCR), and using Variable Valve

Timing (VVT) to change the effective compression ratio and/or the amount of hot exhaust gases retained in the cylinder. VCR and VVT technologies are particularly attractive because their time response could be made sufficiently fast to handle rapid transients (i.e., accelerations/decelerations).

Fuel injection is one of the key parameters to achieve the HCCI mode; therefore, its contribution to the diesel engine needs to be studied. The combustion processes that take place inside a diesel engine are essentially dependent on the way in which the fuel is injected into the combustion chamber. The most important criteria are the timing and duration of the injection, the degree of atomization and the distribution of the fuel inside the

xxv combustion chamber, the timing of ignition, the mass of fuel injected relative to the crankshaft rotation, and the total amount of fuel injected relative to engine load. For a diesel engine and its fuel injection-system to the function properly, all these factors must be carefully balanced, relying on an efficient injector-diesel pump couple.

Older engines make use of a mechanical fuel pump and valve assembly which is driven by the engine crankshaft, usually via the timing belt or chain. These engines use simple injectors, which are basically very precise spring-loaded valves, that open and close at a specific fuel pressure. The

HCCI experimentation makes use of the latest diesel technology of having a separate fuel pump, which supplies fuel constantly at high pressure to each injector.

Each injector then has a solenoid which is operated by an electronic control unit that enables more accurate control of injector opening times, depending on other control conditions such as engine speed and loading, resulting in better engine performance and fuel economy. This design is also mechanically simpler than the combined pump and valve design, making it generally more reliable, and less noisy than its mechanical counterpart. Both mechanical and electronic injection systems can be used in either direct or indirect injection configurations.

The HCCI is considered as an alternative to the SI and CI engines due to its capability of producing very low NOx and operating with higher efficiency, combining the advantages of both the engines. The HCCI concept involves premixing of fuel and air prior to its induction into the cylinder (as is done in the present SI engine) then igniting the fuel-air mixture through the compression process (as is done in current diesel engines). The combustion occurring in an HCCI engine is fundamentally different from that in an SI or diesel engine. In HCCI, the heat release occurs as a global auto-ignition

xxvi process, as opposed to the turbulent flame propagation or mixing controlled combustion used in current engines. The short combustion duration is closer to the ideal constant volume combustion of the otto cycle, thus maximizing the engine efficiency. The advantage of this global auto-ignition is that the temperatures within the cylinder are uniformly low, yielding very low emissions of oxides of nitrogen (NOx, the chief precursors to photochemical smog). The inherent features of HCCI combustion allow the design of engines with efficiency comparable to, or potentially higher than that of diesel engines.

Since there is no flame propagation, the air-fuel mixture can be made very lean. The lean mixture burns at a lower temperature compared to a stoichiometric mixture, resulting in significantly less NOx emissions. Since the mixture is homogeneous, soot emissions are not a concern. The main limitation of the HCCI is the narrow operating window which results from the lack of direct ignition timing control. In an SI engine, the start of ignition is easily controlled by adjusting the spark timing, and in a diesel engine, ignition timing is controlled by varying the fuel injection timing. In the HCCI, the start of ignition is mainly determined by the chemical kinetics reaction of the air-fuel mixture. In the present research work, various methods have been proposed and investigated for HCCI ignition timing control, including intake charge preheating, internal or external EGR, and a stratified fuel injection system.

In a low compression ratio engine, it is necessary to preheat the airfuel mixture to a higher temperature prior to entering the cylinder to ensure proper HCCI combustion. The preheating process reduces the mass flow rate of the air-fuel mixture into the cylinder, further reducing the engine power density.

xxvii

To meet the demand for economy, energy conservation, less environmental harmful exhaust emissions, especially carcinogenic NOx, and responsibility for a greenhouse effect of CO

2

, the HCCI combustion concept is ideal. This concept of a new engine that achieves higher efficiency with lower fuel consumption and generates less NOx emissions has recently been proposed. Many institutes have already studied HCCI, but only a few of them performed experiments using bio diesel as a potential alternative fuel.

HCCI is a strategy that has shown the possibility of both lower emissions and lower fuel consumption than SI combustion. However, HCCI combustion can be sensitive to changes in fuel composition. SI engines use a homogeneous air-fuel mixture that is compressed and subsequently ignited by an electric spark. CI engines compress the air charge to a higher level than SI engines and then the fuel is injected into the air, which is hot enough after compression to ignite the fuel. This results in a highly unhomogeneous mixture. It has been shown that an engine can be run using a combination of the SI and CI strategies, by utilizing a homogeneous mixture, but relying on the compression to ignite the mixture. This approach is called the HCCI.

Many different strategies have been formed from this basic idea and they have different names such as Controlled Auto-Ignition (CAI), Low Temperature

Combustion (LTC), Premixed-Charge Compression Ignition (PCCI), etc. The

HCCI operates much better and has no flame front, which results in low in-cylinder temperatures, and hence, low NOx formation. The load in an

HCCI engine is controlled by the amount of fuel, allowing unthrottled operation. This reduces the pumping losses and decreases the fuel consumption.

Even though HCCI combustion can provide emissions and fuel consumption benefits compared to SI combustion, it is still important to investigate the effect of fuel composition on emissions and fuel consumption.

xxviii

1.3.1 HCCI Combustion Chemistry

The HCCI auto-ignition combustion process is chemically controlled. The homogeneous mixture is compressed until ignition occurs simultaneously at multiple spots across the combustion chamber. The influence of turbulence is limited as the mixture is already homogeneous, so turbulence will not alter mixture composition, and there are no flame fronts or diffusion combustion to affect the process. HCCI combustion can be divided into two phases: a low temperature and a high temperature reaction phase.

The low temperature reactions start approximately at 700 K and can be described by dividing the process into the following stages: initiation, chain propagation, degenerate branching and chain terminating. During the initiation steps, radicals are formed that react further with fuel molecules to form more radicals during the chain propagating stage. During the branching step, molecules that were formed during the previous step split up further into two radicals. During the chain propagation step, hydrogen peroxide (H

2

O

2

) is formed, which dissociates at temperatures of around 1100 K into two oxygen and hydrogen (OH) radicals. This dissociation is considered to be an important step, as the reaction of the OH radicals with fuel molecules initiates the main combustion process.

The low temperature combustion process is both temperature and pressure dependent, so altering the ambient conditions during compression affects the low temperature reactions, and thus the start of the high temperature combustion. The low temperature reactions are clearly distinguished. The high temperature reactions start at a temperature below

1100 K, at which H

2

O

2 decomposition starts, but it should be noted that the temperature curve represents the average gas temperature, rather than the local temperature where the H

2

O

2 decomposition occurs.

xxix

1.3.2 Ignition Resistance of Diesel Fuel

The low resistance to the self-ignition property of diesel fuel is desirable for a conventional diesel operation, but not for an HCCI operation.

When the fuel ignites too early, the combustion will be very rapid, noisy, and potentially damaging to the engine components. For diesel fuel, the charge should, therefore, be cooler than the normal to avoid the fuel ignition too early. An effective way to reduce the temperature during compression is to reduce the compression ratio, which is typically high for diesel engines.

Cooled external EGR can also be used. The presence of EGR lowers the ratio of specific heat, and thus the temperature during compression to some degree, but the main effect of EGR is that it lowers the rate of combustion, and is therefore effective in suppressing the knocking of the combustion.

The problem with ignition occurring too early is that it makes the diesel fuel (actually diesel HCCI) operation difficult to achieve. Several researchers have therefore experimented with the use of alternative fuels, such as alcohols or Compressed Natural Gas (CNG), in diesel engines used for

HCCI operation. This approach shares similarities with the attempts to develop HCCI engines based on SI engines. However, the use of diesel fuel is attractive, as it offers the potential to combine HCCI combustion with diesel combustion using a single fuel and combined fuel system.

The ignition quality of diesel fuel is defined by its cetane number.

While the octane number used in conjunction with petrol increases with its resistance to ignition, the cetane number increases with the ignitability of the fuel. Fuels with higher cetane numbers require less activation energy, which in practice means lower temperatures, in order to ignite (Rui 2002).

1.3.3 Injection System

xxx

Depending on the fuelling technique, investigations and the development of diesel-fuelled HCCI may be divided into three main categories; firstly, the premixed HCCI, in which the fuel is dispersed into the air prior to intake; secondly, the early direct injection HCCI, in which fuel is injected to the in-cylinder sufficiently in advance of TDC, so that mixing can be accomplished prior to auto-ignition, and finally, the late direct-injection

HCCI, in which fuel is injected later than for conventional diesel combustion, with high swirl levels for rapid mixing and high levels of EGR to delay ignition that allows the fuel and air to be sufficiently mixed prior to autoignition (Christensen 1997).

1.3.4 Premixed HCCI

Early and recent probed investigations into diesel-fuel HCCI, in which the fuel is injected into the intake air, upstream of the intake valve, similar to a conventional Port Fuel-Injected (PFI) SI engine to obtain a premixed charge. Later research involved an intake air upstream of the fuel injection point, and allowed preheating, a compression ratio varying from 7.5 to 17.1 and the application of EGR; the following key issues were solved that led to dramatically decreased NOx and soot emissions.

First, very premature ignition and knocking occur, if normal diesel compression ratios are used. Ideally, the HCCI mode requires a compression ratio of 10 to 11.2.

Second, when intake temperatures are reduced below about

130°C, smoke emission increases significantly. In the HCCI mode, intake temperatures of 90°C to 130°C are required.

Third, the HCCI mode makes very high emissions of unburned HC.

xxxi

Premixed diesel-fuelled HCCI has also been accomplished when combined with a normal diesel injection. A Gasoline Direct Injection (GDI) injector with 5 Mpa injection pressure was used to introduce diesel fuel into the intake manifold. However, this led to severe knock problems at high loads. Hence, a mixture of 50% Methyl-tert-Butyl Ether (MTBE) and 50% diesel fuel was used to delay the auto-ignition sufficiently for combustion to be controlled by the near-TDC Direct Injection (DI).

Premixed diesel-fuel has also been tested by using a pintle-nozzle injector (12 Mpa injection pressure) mounted on an intake manifold. In the present research work, the intake temperature was maintained at 20°C.

However, the use of diesel fuel led to a dramatic increase of HC emissions and the dilution of the lubricating oil. Nevertheless, by replacing the diesel fuel with a light naphtha (distillation characteristics similar to those of gasoline), the latter problems were virtually eliminated.

The results of these investigations using premixed fuelling show the strong potential of HCCI to substantially reduce NOx and soot emissions of diesel-fuelled engines. Alternative fuel-delivery and mixing techniques are most likely to be required to overcome the difficulties associated with using diesel fuel for HCCI.

1.3.5 Early Direct Injection HCCI

A direct fuel injection technique, well before TDC, compared to premixing in the intake offers three potential advantages:

By injecting the fuel part way up the compression stroke, the higher in cylinder temperatures and densities can help vaporize the diesel fuel and promote mixing. This allows

xxxii cooler intake temperatures, reducing the tendency for early ignition.

With a carefully designed injector, the possibility exists to minimize fuel wall wetting that can cause combustion inefficiency and oil dilution.

In principle, only one fuelling system is required for both

HCCI and conventional-diesel operations; however, not all the

HCCI Direct Injection (DI) fuelling systems reported in the literature are compatible with conventional diesel combustion.

The main disadvantage of early-DI for HCCI is that less time is available for fuel/air mixing the NOx and PM emissions can be significant if mixing is not sufficiently complete. It is also easy to produce wall-wetting due to over-penetration of the fuel. Finally, controlling combustion phasing is still a critical issue for early-DI HCCI, since the injection timing does not provide an effective means of directly controlling combustion phasing as it does in conventional diesel combustion.

1.3.6 Late Direct Injection HCCI

Nissan Motor Company (2002) successfully achieved a diesel fuelled HCCI, limited to half loads and three quarters speed with a late injection DI-HCCI technique known as Modulated Kinetics (MK). Ignition delay was retarded to 3° After Top Dead Center (ATDC) with the use of high levels of EGR cooling, and the swirl ratio increased to 12. In their experiment, the compression ratio of 16 and the piston bowl of 56 mm contributed to the improvements. Hence, NOx was reduced significantly and PM emissions were improved.

xxxiii

1.3.7 Variable Valve Timing

By the Variable Valve Timing (VVT) method, the exhaust valve is closed earlier during the exhaust stroke compared to a common diesel engine.

Hence, hot residual gases are retained in the cylinder and re-inducted from the previous combustion cycle. Variable Valve Actuation (VVA) allows control over the compression ratio and the exhaust gas percentage. However, VVA is complicated and the required components are expensive.

1.4 BIO-DIESEL

Bio-diesel is made by chemically reacting vegetable oil or animal fat (or combinations of oils and fats) with alcohol (usually nearly pure methanol or denatured ethanol) and a catalyst (sodium hydroxide, or lye). The oil is chemically acidic; the alcohol is chemically a base. This chemical reaction breaks the fat molecules in the oils into an ester, which is the biodiesel fuel, and glycerol. This reaction is called trans-esterification. Since the bio-diesel is less dense than the glycerol, it floats on top of the glycerol and may be pumped off, or the glycerol can be drained off the bottom. The fuel can then be filtered and used in heating or lighting applications. Some researchers use it in diesel engines without further processing, but others recommend removing impurities (soap, un-reacted alcohol, and sodium hydroxide) by a washing process. It is well known that the diesel fuel turns into a waxy gel at low temperatures and cannot be pumped until it is warmed up. The temperature at which the fuel no longer pours is called the pour point or gel point. Bio-diesel has a higher pour point than petroleum diesel (biodiesel gels at a higher temperature). Some oil feed-stocks, such as coconut oil or animal fats, result in bio-diesel that gel at relatively high temperatures, where as bio-diesel made from canola or rapeseed oil have a lower pour point.

Additives can lower the pour point in cold weather, or bio-diesel can be mixed with petroleum diesel to lower the pour point. Bio-diesel should be

xxxiv stored at temperatures above the freezing point, and temperature controlled heaters can be installed on tanks and fuel lines in diesel vehicles.

1.4.1

Bio-Diesel in Diesel Engine

Diesel engines are the most efficient prime movers. From the point of view of protecting the global environment, and concerns for long-term energy security, it becomes necessary to develop alternative fuels with properties comparable to petroleum-based fuels. Unlike the rest of the world,

India’s demand for diesel fuels is roughly six times that of gasoline; hence, seeking an alternative to mineral diesel is a natural choice need (Barnwal et al

2005). The rapid depletion of petroleum reserves and rising oil prices has led to the search for alternative fuels. Non edible oils are promising fuels for agricultural applications. Vegetable oils have properties comparable to those of diesel and can be used to run CI engines with little or no modifications.

Usage of bio-diesel will allow a balance to be sought between agriculture, economic development and the environment. Jatropha curcas is a non-edible oil being singled out for large-scale plantation on waste lands. The jatropha curcas plant can thrive under adverse conditions. It is a drought resistant, perennial plant, living up to fifty years and has the capability to grow on marginal soils. It requires very little irrigation and grows in all types of soils.

The production of jatropha seeds is about 0.8 kg per square meter per year.

The oil content of jatropha seed ranges from 30% to 40% by weight and the kernel itself ranges from 45% to 60%. Fresh jatropha oil is slow drying, odorless and colorless, but it turns yellow after aging (Sarin et al 2007). In

Madagascar, Cape Verde and Benin, jatropha oil was used as a mineral diesel substitute during the Second World War. Forson et al (2004) used jatropha oil and diesel blends in CI engines and found its performance and emission characteristics similar to those of mineral diesel at low concentration of jatropha oil in blends. Pramanik et al (2003) tried to reduce the viscosity of

xxxv jatropha oil by heating it and blending it with mineral diesel. Additives are abundantly manufactured and mixed with IC engine fuels to achieve the proper performance of fuel in an engine. Additives act like catalysts so that they aid combustion, control emission and fuel quality during distribution and storage, and reduce refiner’s operating cost. Now in India, Multi Functional

Additives (MFAs) are sold in the retail market for better mileage of the vehicles and keeping the engine components clean, for better performance and to decrease pollution. For a long time, the industry has been using various types of chemical additives which are corrosive, toxic and non- ecofriendly.

Use of MFAs for diesel will lead to better fuel conservation and low emission.

In the present research work, the design and fabrication of valve arrangements, secondary injector and heating setup have been developed for analyzing the emission characteristics of IC engines in the HCCI mode operation with premixed conditions. The exhaust gas emissions of NOx, CO,

CO

2

, HC, smoke and O

2

have been analyzed under different load conditions with different fuel combinations. The performance analysis also has been carried out to study the effect of different fuel combinations in the HCCI mode. Fuels such as diesel, petrol and bio diesel (100%) and combinations of any of the two fuels which are mentioned above in different mixing ratios have been studied, in order to identify the optimum mixing ratio for different fuel combinations.

1.5 ORGANIZATION OF THE THESIS

In the present thesis, Chapter 1 describes the importance of IC engines, HCCI engines, bio fuels and their emission characteristics. The literature on the effect of the operating parameters on HCCI engine with different fuels, and the general scenario of IC engines are reviewed in Chapter

2. The objectives and scope of the study are also discussed in this chapter.

The description of the experimental setup, properties of the samples and the

xxxvi experimental procedure are detailed in Chapter 3. The experimental results obtained are analyzed and the reasons for the different results are discussed in

Chapter 4. The conclusions derived from this work and directions for future investigations are presented in Chapter 5.

xxxvii

CHAPTER 2

LITERATURE REVIEW

2.1 INTRODUCTION

The Homogeneous Charge Compression Ignition (HCCI) engine is a promising concept for future automobile engines and stationary power plants. The main features of HCCI are breathing premixed air/fuel mixture, as in conventional Spark Ignition (SI) engines, and ignition without a spark plug, as in conventional Compression Ignition (CI) engines. Ultra lean burn combustion is achieved through homogeneous mixture formation and compression ignition, enabling a combustion temperature much lower than that of conventional SI and CI engines. Owing to this lean mixture low temperature combustion, Nitrogen Oxide (NOx) emissions are reduced dramatically and fuel economy is improved. However, greater amounts of

Hydrocarbon (HC) and Carbon monoxide (CO) emissions are released relative to the conventional SI and CI engines. The oxidation reactions of HC and CO emissions during the expansion stroke are reduced due to the lower combustion temperature. Most previous studies in this field have been carried out in conjunction with conventional gasoline and diesel fuels. However, the early combustion of gasoline HCCI limits its operating range. The early combustion leads to high combustion pressure and knock. Another important issue in relation to the HCCI engine is the combustion phase control. Hot residual gas supplies heat to the combustion chamber and promotes HCCI combustion. This hot residual gas can be controlled by a Variable Valve

Timing (VVT) device. Moreover, the VVT device can improve volumetric

xxxviii efficiency by varying the intake valve’s open and close timing

(Mohanamurugan et al 2009). An exhaustive literature on the effect of the operating parameters the HCCI engine with different fuels and the general scenarios of IC engines are reviewed in detail in this chapter. The objectives and scope of the study are also discussed.

2.2 HOMOGENEOUS CHARGE COMPRESSION IGNITION

Onishi et al (1979) used schlieren photography to visualize the combustion process in an engine operating in spark-ignited and HCCI modes.

The schlieren imaging showed very distinct flame propagation during the spark-ignited operation and no apparent flame propagation during the HCCI operation.

Ishibashi et al (1998) studied the emission characteristics of two stroke spark-ignited engines and HCCI motor cycle engines. The conclusions from their studies are that Hydrocarbon (HC) and Carbon monoxide emissions (CO) from the HCCI engine are still very high compared with the current automotive emission standards. An improved version of the engine has been recently evaluated, which shows improvements in fuel economy and emissions.

Thring (1989) studied a four-stroke, Diesel- fueled, HCCI operation with control, by varying the intake temperature and EGR fraction over a range of equivalence ratios. The diesel fueled HCCI engine achieved lower indicated specific fuel consumption than a conventional diesel engine; although the energy needed for intake preheating was not accounted for in this assessment.

Law et al (2001) studied the combustion in an IC engine under controlled conditions by the fully variable valve train method and concluded

xxxix that, variable valve actuation could effectively vary the engine compression ratio and be used to retain the necessary residual gas to control timing.

Residual gases provide a readily available source of heat that can be used to adjust the temperature time history of the inducted mixture. Inexpensive, durable valve systems that allow for precision timing adjustments that would be needed, are not currently available.

Stanglmaier et al (2001) studied the efficiency of the HCCI operation of a dual-fuel natural gas engine. The outcome from their studies is that blending two fuels with different auto ignition characteristics controls the ignition timing. This kind of system works effectively to control timing over a small range, but would likely need to be coupled with another type of system

(EGR or intake preheating) to obtain a wider load range control. Also, these systems are somewhat impractical for production engines because of the system and infrastructure requirements imposed by two fuels.

Heywood et al (1988) explained the fundamentals of combustion in an IC engine. The important points from his studies are, particulate matter or soot emission reduction is a main challenge in diesel engine development.

The source of soot formation in a diesel engine is the diffusion-limited combustion process and the presence of locally fuel-rich regions within the cylinder. HCCI combustion is found to produce very low soot emissions due to the homogeneous nature of the mixture. However, soot emissions can be a problem once mixture heterogeneity is introduced into the cylinder.

Christensen et al (1997) found that NOx emissions in the HCCI operation depend on the fuel characteristics. Fuels with a higher octane number require higher temperature for auto-ignition, thus resulting in higher

NOx emissions. During HCCI combustion, the cylinder peak temperature is relatively low compared to that of SI or diesel combustion, which leads to a significant reduction of NOx emissions.

xl

Richter et al (1999) employed the absorption spectroscopy technique to study the effect of different fuels on HCCI combustion. The results showed similar absorption spectra when using methanol, ethanol, or isooctane as the fuel. Distinct absorption features were observed at 310 nm and 284 nm, which correspond to Oxygen and Hydrogen (OH) radical absorption. When the engine was fueled with a mixture of isooctane and n-heptane, the spectra showed a significant absorption in the Ultra Violet

(UV) region from about 20

0

Before Top Dead Center (BTDC) until the start of the main heat release, which was attributed to the presence of the cool flame.

Christensen et al (1997) suggested that CO emissions are highly dependent on combustion temperature, with higher intake temperature yielding lower CO emissions. This condition usually occurs close to the rich limit. At low load operation with lower intake temperature, a high concentration of CO can be generated. The author also noted that CO emissions increase with an increasing compression ratio. The main reason for this behavior is the lack of time for complete fuel oxidation. With a higher compression ratio, the cylinder temperature decreases faster during the exhaust stroke and there is less time for complete oxidation of CO to CO

2

.

Dae and Chang (2006) investigated partial HCCI (homogeneous charge compression ignition) combustion as a control mechanism for HCCI combustion. The premixed fuels used in this experiment are gasoline, diesel and n-heptane. The results show that with diesel premixed fuel, a simultaneous decrease of NOx and soot can be obtained by increasing the premixed ratio. However, when the inlet charge is heated for the improved vaporization of diesel fuel, higher inlet temperature limits the operational range of HCCI combustion due to severe knocking and high NOx emission at high premixed ratios. Gasoline premixing shows the most significant effects

xli in the reductions of NOx and soot emissions, compared to other kinds of premixed fuels.

An engine with HCCI combustion offers a number of benefits over conventional SI and CI, such as much lower NOx emission, negligible cycleto-cycle variation, higher combustion efficiency at part load than its SI counterpart, and low soot emissions. Unlike the conventional SI and CI engines, where the combustion is directly controlled by the engine management system, the combustion in HCCI engines is controlled by its chemical kinetics only. Trapping hot internal EGR appears to be a potential technology for a practical application of HCCI (Rui and Nesa 2002).

The HCCI engine has the potential to combine the best of the SI and CI Engines. With a high octane number fuel, the engine operates with a high compression ratio, and lean mixtures giving an equivalent fuel consumption of a CI engine or better. Due to premixed charge without rich or stoichiometric zones, the production of soot and NOx can be avoided (Bengt

2002).

Lei et al (2006) studied the effects of internal and cooled external

EGR on the combustion and emission performance of diesel fuel HCCI. The use of fuel injection before the TDC of an exhaust stroke and the negative valve overlap (NVO) to form the homogeneous mixture achieves low NOx and smoke emissions of HCCI.

2.3 IC ENGINE EMISSIONS

Dae (2007) studied the combustion and emission characteristics of a partial HCCI engine with a two stage injection system. The effect of the premixed ratio and timings of the first and main injection on the combustion characteristics and exhaust emissions in a DI diesel engine were discussed.

xlii

The results showed that two-stage injection was very effective in reducing

NOx emissions from a DI diesel engine. Additionally, at optimized injection timing, the ignition of the premixed fuel can be controlled by the main injection without premature auto ignition.

The simultaneous reduction of nitrous oxides and soot emissions has been analyzed by Molina (2006) by the highly premixed combustion concept in diesel engines. Retarding the injection process until the expansion stroke avoided the instabilities in the auto-ignition. The effect of the start of injection, the injection pressure and the boost pressure, has also been analyzed. The sufficiently extended premixed combustion phase showed the reduction of both nitrogen oxides and soot emissions.

Francisco (2009) studied the amount of each hydrocarbon species present in the exhaust gases of a diesel engine operated with different biodiesel blends. The levels of reactive and non-reactive hydrocarbons present in diesel engine exhaust gases powered by different bio-diesel fuel blends, were also analyzed. The conclusions from this study are, the use of oxygenated fuel blends shows a reduction in the Engine-Out emissions of total hydrocarbons.

But the potential of the hydrocarbon emissions is more dependent on the compositions of these hydrocarbons in the engine-out, than on the quantity; a large percent of hydrocarbons existing in the exhaust, when bio-diesel blends are used, are partially burned hydrocarbons, and are interesting as they have the maximum reactivity; but with the use of pure bio-diesel and diesel, most of the hydrocarbons are from unburned fuel and they have lesser reactivity.

The best composition in the fuel, for the control of the hydrocarbon emissions reactivity, needs to be a fuel with high-saturated fatty acid content.

The effect of narrow fuel spray angle injection and dual injection strategy on the exhaust emissions of a common rail diesel engine was investigated by Myung (2007). The investigation showed that a dual injection

xliii strategy consisting of an early timing for the first injection for HCCI combustion and a late timing for the second injection was effective to reduce the NOx emissions while it suppresses the deterioration of the combustion efficiency caused by the HCCI combustion.

Dae (2006) studied the improved emission characteristics of an

HCCI engine with various premixed fuels. Partial HCCI combustion was used as a control mechanism. The premixed fuel is supplied via a port fuel injection system located in the intake port of a DI diesel engine. The results show that with diesel premixed fuel, a simultaneous decrease of NOx and soot can be obtained by increasing the premixed ratio. However, when the inlet charge is heated for the improved vaporization of diesel fuel, the higher inlet temperature limits the operational range of HCCI combustion due to severe knocking and high NOx emission at high premixed ratios. Gasoline premixing shows the most significant effects in the reductions of NOx and soot emissions, compared to other kinds of premixed fuels.

A new concept of combustion at low flame temperature, based on the use of a highly diluted charge, is investigated by Claudio (2006). Its advantages and drawbacks on diesel engine application are critically discussed. It was found that the flame temperature limit, corresponding to soot and NOx free conditions, can be achieved at a low load by the accurate control of the EGR and injection parameters. Tests have demonstrated that near zero emission limits can be reached in a diesel engine at low load without strong fuel economy penalties. The use of a low temperature superdiluted combustion mode is limited to a low load range, indicating that further major research activity is needed to reach reliable practical applications.

A compound combustion technology of "Premixed Combustion" and "Lean Diffusion Combustion" for the purpose of realizing the concept of

HCCI combustion in a DI diesel engine, is analyzed by Wanhua Su (2004).

xliv

A flash mixing technology is developed from the development of the socalled BUMP combustion chamber. The combustion of fuel in the main injection occurs at a much higher equivalence air/fuel ratio than the one that occurs in a conventional diesel engine, which is called lean diffusion combustion.

Miguel (2009) experimentally studied the performance of a modified diesel engine operating in the HCCI combustion mode versus the original diesel combustion mode. The experimental results for the modified diesel engine in the HCCI combustion mode fueled with commercial diesel fuel were compared to those of the diesel engine mode. An experimental installation, in conjunction with systematic tests to determine the optimum crank angle of fuel injection, has been used to measure the evolution of the cylinder pressure and to get an estimate of the heat release rate from a singlezone numerical model. From these, the angle of start of the combustion has been obtained. The performance and emissions of HC, CO and the huge reduction of NOx and smoke emissions of the engine are presented. These results have allowed a deeper analysis of the effects of external EGR on the

HCCI operation mode, on some engine design parameters and also on NOx emission reduction.

An experimental study of HCCI-DI combustion and emissions in a diesel engine with dual fuel has been carried out by Junjun (2008). An experimental study of port injected n-heptane HCCI in combination with incylinder diesel fuel direct injection (DI) was conducted on a single cylinder diesel engine. It was found that the NOx emissions decreased dramatically with partial premixing and it exhibited a descending trend as a function of r

p

increase when the r

p

was lower than 0.3, but it showed a great tendency to increase when the premixed ratio was higher. The inherent trade-off of NOx and soot was not obvious since the soot emissions remained at the same level

xlv of the prototype diesel engine under the condition of lower r

p

. Moreover, the influence of r

p

on the CO and Unburned Hydro Carbon (UHC) emissions was assessed. The results also revealed that the HCCI-DI combustion could effectively improve the indicated thermal efficiency of the diesel engine at low to medium loads.

Wimmer (2006) studied the potential of the HCCI concepts for DI diesel engines. Several alternative combustion processes are being studied as options for reducing the emissions of nitrogen oxide (NOx) and soot of direct injection (DI) diesel engines. Such processes are characterised by adequate combustion control that helps to avoid areas of high local flame temperatures, which lead to NOx formation, and fuel rich combustion areas which promote soot formation. In most cases, this will be achieved by a complete or partial charge homogenisation prior to combustion. Four processes of alternative combustion control have been applied to a single-cylinder research engine and compared to conventional diesel combustion. By means of adequate measurement techniques and engine cycle simulation, the combustion processes were subjected to an exact thermodynamic analysis in terms of mixture formation, combustion and wall heat transfer, and assessed for their potential relation to emissions, efficiency and load limits.

Homogeneous charge compression ignition (HCCI) combustion of diesel fuel with external mixture formation was studied by Ganesh (2010).

A fuel vapouriser was used to achieve excellent HCCI combustion in a single cylinder air-cooled direct injection diesel engine. No modifications were made to the combustion system. In this study, a vaporized diesel fuel was mixed with air to form a homogeneous mixture and inducted into the cylinder during the intake stroke. To control the early ignition of the diesel vapour–air mixture, the cooled (30°C) EGR technique was adopted. Experiments were conducted with diesel vapour induction without EGR and diesel vapour

xlvi induction with 10%, 20% and 30% EGR, and the results are compared with those of conventional diesel fuel operation (DI at 23° BTDC and 200 bar injection pressure).

Zeynep (2007) studied the development of the HCCI engine as compared to the diesel engine; the study shows that, diesel engines produce high levels of NOx and particulate matter, and spark ignition engines have lower efficiency. HCCI appears as a new combustion technique that could combine in itself the advantages of both engines significant soot and NOx reduction, and low consumption. Various names have been attributed to this particular mode of combustion such as Controlled Auto-Ignition (CAI), PCCI etc.

Marcello (2007) explained that low NOx and particulate matter emissions at part-load operations are achievable by Homogeneous Charge

Compression Ignition and HCCI combustion can be obtained on conventional diesel engines by premixing the charge in the intake manifold with a dedicated fuel atomiser. His study focused on experimental and modeling activities oriented to understanding and controlling diesel HCCI combustion with external mixture formation. Results obtained from different engines showed that stable diesel HCCI combustion could be achieved over a range of operating conditions, and confirmed its benefits in terms of NOx and soot reduction.

The effect of the compression ratio on exhaust emissions from a

PCCI diesel engine was studied by Laguitton (2007) with the objective to establish how engine-out NOx emissions can be reduced to the estimated levels required by the next emissions target ‘Euro 6’ and thus be able to apply the findings to the original 4-cylinder engine and minimize the requirement for currently immature NOx after-treatment. It has been proposed that further reduction in the compression ratio beyond current levels would be beneficial

xlvii to engine-out emissions and specific power, and could be facilitated by developments in cold-start technology. The results of a study using this single-cylinder facility to evaluate the effect of reducing the compression ratio from 18.4 to 16.0 are presented. It was found that, although there was a small

CO and HC penalty, either reducing the compression ratio or retarding the injection timing greatly reduced NOx and soot emissions when both premixed and diffusion–combustion phases were present. This effect was less significant when the combustion was solely premixed.

Mingfa (2009) studied the progress and recent trends in homogeneous charge compression ignition engine. The observations from this study are that HCCI combustion has been drawing considerable attention due to high efficiency and lower Nitrogen Oxide (NOx) and Particulate Matter

(PM) emissions. Massive research throughout the world has led to great progress in the control of HCCI combustion. The key to diesel-fuelled HCCI combustion control is mixture preparation, while EGR is the main path to achieve gasoline-fuelled HCCI combustion. Specific strategies for dieselfuelled, gasoline-fuelled and other alternative fuelled HCCI combustion are also discussed by Mingfa (2009).

Gilles (2006) studied the combustion in different engines and explained that the necessary reduction of tail pipe pollutant emissions such as nitrogen oxides and soot, must not threaten this contribution to reducing the global warming effect by unacceptable after-treatment costs. By achieving a low temperature lean combustion, HCCI allows the generation of a very small amount of particles and NOx, and therefore may avoid the need of a NOx trap for the main vehicle applications. Its success depends on the ability to control noise and hydrocarbon emissions at part load as well as to allow high power and torque densities.

xlviii

Tiegang fang (2009) studied bio-diesel combustion in an optical

HSDI diesel engine under low load premixed combustion conditions. The experimental results indicated that the heat release rate was dominated by a premixed combustion pattern and the heat release rate peak became smaller with injection timing retardation. The ignition and heat release rate peak occurred later with increasing bio-diesel content. Fuel impingement on the wall was observed for all test conditions. The liquid penetration became longer and the fuel impingement was stronger with the increase in bio-diesel content. Early and late injection timings result in lower flame luminosity due to improved mixing with longer ignition delay. For all the injection timings, lower soot luminosity was seen for the bio-diesel blends than for pure diesel fuel. Furthermore, NOx emissions were drastically reduced in the premixed combustion mode with retarded ATDC injection strategies.

Tests were conducted with two commercially available bioadditives and the results confirmed that pollution can be controlled by reducing CO and HC emissions and conserving fuel by high thermal efficiency (Raghunadham and Deshpande 2004). Ethylene glycol mono-alkyl ethers as oxygenated fuel additives had been studied for performance parameters such as brake specific fuel consumption, brake thermal efficiency and emission levels. A significant reduction in particulate emission is observed with fuel additives.

2.4 NON ROAD DIESEL ENGINE EMISSION STANDARDS

On road diesel engines are employed in literally thousands of applications. The smallest engines may be just a few horsepower (hp). Non road diesel engines are used in farming, construction and industrial applications. Large engines are used to drive the farm equipment that cultivates and harvests crops; and the construction equipment that clears the land for buildings and roads. Smaller engines are used in countless other

xlix applications. Non-Road Diesel (NRD) engines currently account for about

44% of diesel PM emissions and about 12% of NOx emissions from mobile sources nationwide and in some urban areas the percentage is greater. Not surprisingly, environmentalists have gained support for tough new emissions standards for NRD engines, similar to those already set for over-the-highway diesel engines. Regulatory agencies around the world are enforcing stringent rules aimed at reducing the sulphur content of non-road, locomotive, and marine diesel fuel and encouraging to build clean NRD engines. The companies most likely to succeed in this regulatory climate will be those that can meet these new emission standards at competitive prices.

Ayhan (2008) studied bio-fuel sources, bio-fuel policy, bio-fuel economy and global bio-fuel projections. The information received from their studies indicates that, bio-fuels include bio-ethanol, bio-methanol, vegetable oils, bio-diesel, biogas, bio-synthetic gas (bio-syngas), bio-oil, bio-char,

Fischer-Tropsch liquids and bio-hydrogen. Bio-fuels are easily available from common biomass sources; they represent a carbon dioxide-cycle in combustion; bio-fuels have considerable environmentally friendly potential.

The usage of bio fuels gives many benefits to the environment, economy and consumers, and they are biodegradable and so contribute to sustainability. The

Higher Heating Value (HHV) of rapeseed cake oil was found as 36.4 MJ/ kg.

Ayhan (2008) concluded that, this bio-oil can be used in engines and turbines in practice. The predictions say that the modernized biomass energy contribution by 2050 will be about one half of the total energy demand in developing countries during that period. The European Union (EU) has set the goal of obtaining 5.75% of their transportation fuel needs from bio-fuels by

2010 in all member states. The recent commitment by the USA government indicated to increase its bio-energy three-fold within 10 years. The Kyoto

Protocol cannot be achieved without establishing a large role for biofuel in the global energy economy by 2050.

l

Ayhan (2007) studied global bio-fuel scenarios, the various methods of producing bio-fuels like bio-ethanol, bio-methanol, bio-diesel and bio-oil from biomass. The conclusions from his studies are that biomass appears to be an attractive feedstock for three main reasons. First, a renewable resource could be sustainably developed in the future. Second, it appears to have formidably positive environmental properties resulting in no net releases of carbon dioxide (CO

2

) and very low sulfur content. The usage of bio-fuels in the transportation sector is increasing continuously due to its easy availability; it represents a CO

2

cycle in its combustion, has considerable environmentally friendly potential, and biodegradability of its nature contributes to sustainability. Bio-fuels include energy security reasons, environmental concerns, foreign exchange savings, and socioeconomic issues related to the rural sector.

2.5 SETTING THE STANDARD

A new proposal from the U.S. Environmental Protection Agency

(EPA) describes Tier 4 emission standards for non-road diesel engines. These new non-road diesel engine standards are similar to the standards of and in harmony with the non-road diesel standards of the European Union (EU) and other parts of the world. Today’s non-road engines already produce 70% less

NOx emissions and 90 % less PM emissions compared to CI engines. For engine makers striving to meet stage I, II and III emission standards, the stage IV standards aim to reduce emissions another 80 to 90 %. Stage III and stage IV non-road emission standards for select midsize diesel engines. These new emission standards are just the latest in a series of events aimed at motivating diesel engine makers into building cleaner diesel engines. They come as no surprise to diesel engine makers; yet they provide exact targets that need to be met, so engine makers can plan accordingly.

li

2.5.1 Stage I/II Standards

Stage I and II emissions shall not exceed the amount as shown in

Table 2.1. Stage I emissions are engine-out limits and shall be achieved before any exhaust after-treatment device.

Table 2.1 Stage I/II Emission Standards for Non Road Diesel Engines

Stage I/II Emission Standards for Non Road Diesel Engines

Categories **

C

Net Power kW

P < 75

Date*

CO HC NOx PM g/kWh

Stage I

1999.04 6.5 1.3 9.2 0.85

D P < 37

Stage II

2001.01 5.5 1.5 8.0

* Stage II also applies to constant speed engines effective 2007.01

** Diesel Net 2007

0.8

A sell-off period of up to two years is allowed for engines produced prior to the respective market placement date. Since the sell-off period between zero and two years is determined by each Member State, the exact timeframe of the regulations may be different in different countries.

2.5.2 Stage III/IV Standards

Stage III standards which are further divided into two sub-stages:

Stage III A and Stage III B, and Stage IV standards for non-road diesel engines are listed. These limit values apply to all non-road diesel engines of the indicated power range for use in applications other than propulsion of locomotives, railcars and inland waterway vessels.

lii

The implementation dates listed in the following Tables 2.2, 2.3 and 2.4 refer to the market placement dates. For all engine categories, a selloff period of two years is allowed for engines produced prior to the respective market placement date. The dates for the new type approvals are, with some exceptions, one year ahead of the respective market placement date.

Table 2.2 Stage III A Standards for Non Road Engines

Stage III A Standards for Non Road Engines

Net Power

Categories **

Date†

CO NOx+HC PM kW g/kWh

K P < 37 2007.01 5.5 7.5 0.6

† dates for constant speed engines are: 2011.01 for categories K; 2012.01

** Diesel Net 2007

Table 2.3 Stage III B Standards for Non Road Engines

Stage III B Standards for Non Road Engines

Categories **

P

† NOx+HC

** Diesel Net 2007

Net Power kW

P < 56

Date

2013.01

CO HC NOx PM

5.0

g/kWh

4.7†

0.025

Table 2.4 Stage IV Standards for Non Road Engines

Stage IV Standards for Non Road Engines

Net Power

Categories ** Date

CO HC NOx PM kW g/kWh

R P < 130 2014.10 5.0 0.19 0.4 0.025

** Diesel Net 2007

liii

Stage III B standards introduce the PM limit of 0.025 g/kWh, representing about 90% emission reduction relative to Stage II. To meet this limit value, it is anticipated that engines will have to be equipped with particulate filters. Stage IV also introduces a very stringent NOx limit of

0.4 g/kWh, which is expected to require NOx after-treatment.

To represent emissions during real conditions, a new transient test procedure termed as ‘Non-Road Transient Cycle (NRTC)’ was developed in cooperation with the US Environmental Protection Agency (EPA). The

NRTC is run twice with a cold and a hot start. The final emission results are weighted averages of 10% for the cold start and 90% for the hot start run. The new test will be used in parallel with the prior steady-state schedule, ISO

8178 C1, referred to as the Non-Road Steady Cycle (NRSC).

The NRTC (transient) shall be used for the measurement of particulate emissions for Stage III B and IV for all engines but constant speed engines. By the choice of the manufacturer,

NRTC can be used also for Stage III A and for gaseous pollutants in Stage III B and IV.

The NRSC (steady-state) shall be used for stages I, II and III

A and for constant speed engines, as well as for Stage III B and IV for gaseous pollutants.

Understanding the overall effect of using bio-diesel as a fuel is a complex science. This complexity owes to the fact that biodiesel can be produced from almost any plant oil or animal fat.

Each of these feedstocks has different characteristics that can affect production cost, engine performance, and exhaust emissions. The objective of this section is to provide a review of previous research associated with bio-diesel cost,

liv composition, performance, and emissions. A review of the regulations that pertain to non-road engines and alternative fuels is also included.

The primary factors affecting the economics of bio-diesel include the purchase price and the quality of feedstock (Piazza

2007). Raw materials for fuel production, such as soybean oil, cottonseed oil, renderings, and waste oil each carry a purchase price based on the feedstock quantity and geographic availability, competition with other uses of the feedstock, and product quality (Capareda 2007). High quality feed stocks tend to require little pre-treatment, but they can have a high purchase price. Low quality oil can be purchased at a lower price, but usually require a greater deal of pre-treatment.

Pretreatment processes include refining, degumming, neutralizing, drying, bleaching, and dewaxing. Low quality oils also tend to be variable in free fatty acid composition; using low quality fuels can have a negative effect on the end product cold flow properties (Capareda 2007). Feedstock quality is dependent upon the amount of phosphatides, free fatty acids, waxes, insoluble impurities, and water present in the feedstock (Piazza 2007).

2.6 ENGINE PERFORMANCE

According to Piazza (2007), raw material costs have the largest effect on the cost of bio-diesel, and processing costs are significant. Since the energy content of bio-diesel is approximately eight percent lower than that of petroleum diesel, it is expected that, in certain situations, engines fueled with bio-diesel will not produce the same power that is produced when using

lv petroleum diesel. At full load conditions with a Wide-Open Throttle (WOT), or at intermediate loads with equal fuel consumption or accelerator position, the output power should reduce with respect to the energy content (Lapuerta et al 2007). WOT is equivalent to the accelerator being fully pressed.

Contrary to the expectations, researchers have reported varying results. Some researchers have shown a smaller decrease in power than expected when using bio-diesel, while some have reported power loss in the same scope as reduced energy content, and others have shown an increase in the rated power and torque. Some have also reported no significant difference in the output power and torque (Lapuerta et al 2007).

Cetinkaya et al (2005) observed that the reduction of torque was only 3% to 5% when comparing waste oil bio-diesel to petroleum diesel in a

75 kW four-cylinder common rail engine. Lin et al (2006) found that the power at full load when using pure palm oil bio-diesel was only 3.5% less than that of petroleum diesel in a 2.84 lit. naturally aspirated engine. When using a 70% tall oil bio-diesel blend, Altiparmak et al (2007) measured a

6.1% increase in the maximum torque. Usta (2005) observed increases in torque and power when fueling an indirect injection diesel engine with tobacco seed oil bio-diesel blends.

Yucesu and Ilkilic (2006) observed that the heating value of

Cottonseed Oil (CSO) bio-diesel was only 5% less than the heating value of petroleum diesel. They observed power and torque reductions of 3% to 5% when using pure CSO bio-diesel. Murillo et al (2007) also observed power loss similar to the percent reduction in the heating value when using cooking oil bio-diesel.

Romig and Spataru (1996) observed no significant difference in the rated power when using rapeseed and soybean oil bio-diesel blends in a

6-cylinder engine. Shaheed and Swain (1999) also observed no significant

lvi differences when using CSO bio-diesel at several speeds in a single cylinder

2.75 kW engine.

Brake-specific fuel consumption is the ratio between the mass of fuel consumed and the brake effective power produced by an engine. Brake-

Specific Fuel Consumption (BSFC) is inversely proportional to thermal efficiency, Lapuerta et al (2007). Graboski et al (1996) found a good correlation between fuel oxygen content, which is higher for bio-diesel, and

BSFC when using soybean oil the bio-diesel. According to Rakopoulos et al

(2004), the increase in BSFC is attributed to oxygen enrichment from fuel, and not from intake air. Most researchers have reported an increase in BSFC when using bio-diesel and bio-diesel blends. Turrio-Baldassarri et al (2004),

Last et al (1995), Alam et al (2004), Canakci and Van Gerpen (2001) reported increases in BSFC when using bio-diesel and bio-diesel blends, compared to petroleum diesel fuel. These increases tended to be in line with the loss of the heating value in the fuel blends.

According to Lapuerta et al (2007), “thermal efficiency is the ratio between power output and energy introduced through fuel injection”. Most researchers observed no significant change in thermal efficiency when using bio-diesel (Lapuerta et al 2007, Canakci 2005 and Monyem et al 2001).

2.7 EXHAUST EMISSIONS

According to the national bio-diesel board (2006), bio-diesel is a clean burning alternative fuel produced from domestic, renewable resources, such as plant oils or animal fats. While bio-diesel contains no petroleum, it can be blended with petroleum diesel to create a fuel suitable for use in diesel engines. It is important to understand the relationship between bio-diesel blends and exhaust emissions. Pure bio-diesel is essentially free of sulfur compared to petroleum diesel. Bio-diesel blends, consequently, contain less

lvii sulfur than petroleum diesel. Since bio-diesel blends have less sulfur than petroleum diesel, using bio-diesel blends should decrease the emission of SO

2 and sulfate particulate matter.

Munoz et al (2004) found that the concentration of Carbon monoxide (CO) in the exhaust decreased, except at high speed and load, while hydrocarbon emissions Total Hydro Carbons (THC) reduced at low loads, and

NOx emissions depended on the speed and load of the engine when petroleum diesel was replaced with bio-diesel mixtures.

Graboski and McCormick (1998) found that, for bio-diesel, NOx emissions increase when large, two-stroke engines were tested at full load.

Schumacher et al (2001) found that, as the percent mixture of bio-diesel increased, emissions of THC, CO, and PM decreased, while emissions of

NOx increased. Neat bio-diesel exceeded the 1991-1994 nitrogen oxide emission standards. Using bio-diesel blends yielded a positive CO

2

balance.

According to Schumacher et al (2001), as the percentage of bio-diesel in the blend increased, with no timing changes and no addition of alkylates, THC,

CO, and PM exhaust emissions decreased, while NOx increased.

Canakci and Van (2001) found that CO, THC, and NOx emissions significantly increased with pure bio-diesel made from yellow grease and soybean oil in comparison. When using B20, there was not a significant change in the emission of HC and NOx; the decrease in the levels of CO was

“borderline” significant. No significant changes were observed for CO

2 emissions when using bio-diesel. Canakci and Van (2003) found statistically significant reductions in CO and hydrocarbon emissions when using biodiesel from sunflower oil and yellow grease.

Nitrogen oxide emissions are of specific importance due to the fact that NOx is a contributor to ozone. Since it has been found that NOx

lviii emissions tend to increase with bioiesel blends, measures have been taken in an attempt to decrease NOx emissions when using bio-diesel blends.

Sometimes these efforts result in increased emissions of other pollutants.

According to Schumacher et al (2001), NOx emissions can be reduced by retarding the injection timing or by substituting 20 % of the petroleum diesel in the B20 blend with heavy alkylate. Replacing DF with heavy alkylate also reduced CO and PM concentrations, while THC concentrations were not affected. Retarding the injection timing increased the CO concentrations.

Munoz et al (2004) found that, when the injection timing was moved up by 3°, THC emissions generally increased, while NOx emissions were observed to decline slightly at medium engine speeds at certain loads.

Jha et al (2006) blended ethanol with bioiesel blends in an effort to reduce

NOx emissions. Jha et al observed an increase in NOx emissions when using a diesel, bio-diesel, ethanol (DBE) mixture in an old engine, and using DBE in new engines reduced NOx emissions. CO emissions increased with an increase in ethanol in the fuel blends. Kass et al (2006) observed a decrease in

NOx and PM emissions when using a bio-diesel emulsified with 10% water, by mass. Using the emulsified blend along with Exhaust Gas Recirculation

(EGR), lowered both NOx and PM emissions. Last et al (1995) incorporated retarding injection timing with EGR and observed a 30% decrease in NOx emissions.

Federal standards for new non-road diesel engines were originally adopted in 1994 for engines rated greater than 37 kW, to be phased out from

1996 to 2000. In 1996, a Statement of Principles was signed between engine makers EPA, and the California Air Resources Board (ARB), and in 1998, the

EPA finalized the rule reflecting the Standard Operating Procedures (SOP).

Tier 1 standards for engines less than 37 kW and more stringent Tier 2 and 3 standards for all non-road equipment were scheduled to be phased out from

lix

2000 to 2008. In 2004, the EPA finalized Tier 4 emission standards, which are to be phased out through 2015 (Diesel Net 2007). EPA regulations for mobile non-road diesel engines may apply to those who manufacture and import diesel engines intended to be used in non-road vehicles or movable equipment. Producers and importers of vehicles and machinery that use these engines may also be affected. Those who convert non-road vehicles and equipment to use alternative fuels, and those that produce and distribute nonroad diesel fuel also fall under the umbrella of these regulations (EPA 2004).

According to the final rule for the control of emissions of air pollution from non-road diesel engines and fuel EPA 2004, vehicles and fuels are generally treated as a system, so standards shall be promulgated in tandem. This will achieve the greatest emission reductions while maintaining cost-effectiveness. Standards for non-road diesel engines and fuels have been constructed using on-highway diesel engine standards as a model. In an effort to extend advanced emission controls, engine standards and emission test procedures have been set out, along with sulfur control requirements for diesel included in the engine standards will be not-to-exceed the requirements. Diesel engine manufacturers will have to ensure that their nonroad engines meet the standards and specifications laid forth by the EPA, and the engines must not exceed the NTE standards, which are typically 1.25 or

1.5 times the NSPS. In an effort to lower SO

2

emissions, new standards for sulfur content in non-road diesel fuel were established. Starting in 2007, nonroad diesel fuel shall have a maximum sulfur concentration of 500 ppm; starting in 2010, the maximum sulfur concentration was lowered to 15 ppm.

Federal engine and fuel standards affect the emissions of NOx, SO

2

, PM 2.5,

CO, HC, and air toxins (EPA 2004, DieselNet 2007).

In 2006, the EPA adopted the Compression Ignition Internal

Combustion Engine performance standards for stationary diesel engines

lx

(EPA, 2006). The pollutants regulated by this rule include NOx, PM, CO, and

Nonmethane Hydrocarbon (NMHC). Sulfur oxides shall be reduced with the use of low sulfur non-road fuel, and smoke emissions will be reduced.

Sources affected by the NSPS (New Source Performance Standard) include stationary, non-road diesel engines manufactured or reconstructed after 2005.

Engines are considered stationary if they are not mobile and remain at one location for a year (EPA 2006, DieselNet 2007).

The Texas Commission for Environmental Quality (TCEQ) has attempted to ban the sales of B20 in Texas in order to prevent the increase of

NOx levels in several parts of the state. As of December 2006, the ban has been delayed (Lacey, 2007). TCEQ began a program called the Texas Low

Emission Diesel (TxLED) program in order to help control NOx emissions.

Currently, no bio-diesel is TxLED certified, which signifies that the fuel has been formulated to decrease NOx emissions. The EPA is now developing a report regarding NOx emissions when using bio-diesel. Despite some research claiming that NOx increases are negligible, an EPA study in 2002 found that using a 20% bio-diesel blend increased NOx emissions above the levels allowed by TxLED (Lacey, 2007). If TCEQ decides to uphold the ban on sales of bio-diesel blends throughout Texas or in the 110 Texas counties surrounding the critical air quality areas, Texas bio-diesel producers will likely have to ship their product out of the state or begin building bio-diesel plants out of the state (Lacey 2007, Stillman 2006).

Nitz (2008) observed two important trends concerning the second injection. First, whether the second injection from the Direct Injection (DI) had little effect on combustion and performance. Second, the timing of the second injection had some effect on combustion phasing, but it was less consistent than varying the timing of the first injection. It was suggested that for control purposes, the first injection be fixed and only the timing of the

lxi second injection be varied. The best performance was found when the timing was set near the intake valve open timing.

Laguitton (2007) presented a description of the development of a single cylinder test facility based on the production of a 4-cylinder DI Diesel engine and designed to allow the study of the emission characteristics over a very wide range of operating conditions. The objective was to establish how engine-out NOx emissions can be reduced to the estimated levels required by the next emissions target ‘Euro 6’ and thus be able to apply the findings to the original 4-cylinder engine and minimise the requirement for currently immature NOx after-treatment. It has been proposed that further reduction in the compression ratio beyond the current levels would be beneficial to engineout emissions and specific power, and could be facilitated by developments in cold start technology. The results of a study using this single cylinder facility to evaluate the effect of reducing the compression ratio from 18.4 to 16.0 are presented. It was found that, although there was a small CO and HC penalty, either reducing the compression ratio or retarding the injection timing greatly reduced NOx and soot emissions, when both premixed and diffusioncombustion phases were present. This effect was less significant when the combustion was solely premixed.

Payri1 (2006) discussed the combustion and exhaust emissions in a heavy-duty diesel engine with an increased premixed combustion phase by means of injection retarding. A an experimental study has been conducted in a diesel engine operating at light load with standard injection and combustion systems, retarding the start of injection with the aim of promoting the first phase of combustion in premixed conditions. A detailed study of the combustion characteristics has been carried out by means of a combustion diagnosis model, and the basic phenomena behind the changes in engine performance and pollutant emissions have been studied. Retarded fuel

lxii injection has produced very low levels in Nitrogen Oxides (NOx) and soot emissions, with the inconvenience of higher Carbon monoxide (CO) and unburned or partially burned Hydrocarbon (HC) emissions and a significant penalty in fuel efficiency. At the light engine load operation point studied here, the combination of retarded fuel injection and the introduction of

Exhaust Gas Recirculation (EGR) have proven to be very efficient in achieving soot and NOx levels below the future emission regulations, at the cost of increased fuel consumption and high CO and HC emissions.

Mingfa (2009) discussed HCCI combustion that has been drawing considerable attention due to its high efficiency and lower nitrogen oxide

(NOx) and Particulate Matter (PM) emissions. However, there are still tough challenges in the successful operation of HCCI engines, such as controlling the combustion phasing, extending the operating range, and high unburned hydrocarbon and CO emissions. Massive research throughout the world has led to great progress in the control of HCCI combustion. The first thing paid attention to, is that a great deal of fundamental theoretical research has been carried out. First, numerical simulation has become a good observational and powerful tool to investigate HCCI and to develop control strategies for HCCI because of its greater flexibility and lower cost compared with engine experiments. Five types (of models) applied to HCCI engine modelings are discussed by Mingfa (2009). Second, HCCI can be applied to a variety of fuel types. Combustion phasing and operation range can be controlled by the modification of fuel characteristics. Third, it has been realized that advanced control strategies of fuel/air mixture are more important than a simple homogeneous charge in the process of controlling HCCI combustion processes. The stratification strategy has the potential to extend the HCCI operation range to higher loads, and Low Temperature Combustion (LTC) diluted by Exhaust Gas Recirculation (EGR) has the potential to extend the operation range to high loads; even to full loads, for diesel engines. Fourth,

lxiii optical diagnostics has been applied widely to reveal in-cylinder combustion processes. In addition, the key to diesel-fuelled HCCI combustion control is mixture preparation, while EGR is the main path to achieve gasoline-fuelled

HCCI combustion. Specific strategies for diesel-fuelled, gasoline-fuelled and other alternative fuelled HCCI combustion are also discussed by Mingfa

(2009).

Miguel (2009) presented a predictive model of an engine running in the HCCI combustion mode with diesel fuel, and validated it with the experimental results. The predictive model behaves as a transfer function with four independent input variables; engine speed, air mass flow, fuel consumption, and constant pressure specific heat. Combustion pressure and

NOx emissions are given as results. The model is based on several sub models that include intake flow characterization, the start of combustion from chemical kinetics, and a new heat release law applied in a unique volume.

This new heat release rate for the HCCI combustion mode allows reproducing the combustion chamber pressure for any load condition, including EGR. The predictive model has been developed in a Matlab environment from 269 observations that cover the full operation range of the engine in the HCCI combustion mode. Research is underway to estimate other relevant engine performance parameters, such as brake mean effective pressure, fuel consumption, or engine torque.

Ming and Maozhao (2006) discussed a chemical kinetics model of iso-octane oxidation for HCCI engines. The necessity of developing a practical iso-octane mechanism for HCCI engines is presented after various different experiments and currently available mechanisms for iso-octane oxidation were reviewed and the performance of these mechanisms applied to experiments relevant to HCCI engines was analyzed. A skeletal mechanism including 38 species and 69 reactions is developed, which could predict

lxiv satisfactorily the ignition timing, burn rate and the emissions of HC, CO and

NOx for HCCI multi-dimensional modeling. Comparisons with various experiment data including shock tube, rapid compression machine, jet stirred reactor and HCCI engine, indicate the good performance of this mechanism over wide ranges of temperature, pressure and equivalence ratio, especially at high pressure and lean equivalence ratio conditions.

Amit (2006) conducted a computational study of an HCCI engine with direct injection during gas exchanges. He presented a new Probability

Density Function (PDF) based computational model to simulate an HCCI engine with DI during gas exchange. This Stochastic Reactor Model (SRM) accounts for the engine breathing process in addition to the closed-volume

HCCI engine operation. A weighted-particle Monte Carlo method is used to solve the resulting PDF transport equation. While simulating the gas exchange, it is necessary to add a large number of stochastic particles to the ensemble due to the intake air and EGR streams as well as fuel injection, resulting in increased computational expense. Therefore, in this work he applied a down-sampling technique to reduce the number of stochastic particles, while conserving the statistical properties of the ensemble. In this method some of the most important statistical moments (e.g., concentration of the main chemical species and enthalpy) are conserved exactly, while other moments were conserved in a statistical sense. A detailed analysis demonstrates that the statistical error associated with the down-sampling algorithm was more sensitive to the number of particles than to the number of conserved species for the given operating conditions. For a full-cycle simulation this down-sampling procedure was observed to reduce the computational time by a factor of as compared to the simulation without this strategy, while still maintaining the error within an acceptable limit.

Following the detailed numerical investigation, the model, intended for volatile fuels only, was applied to simulate a two-stroke, naturally aspirated

lxv

HCCI engine fueled with isooctane. The in-cylinder pressure and CO emissions predicted by the model agree reasonably well with the measured profiles. In addition, the new model was applied to estimate the influence of engine operating parameters such as the relative air–fuel ratio and early direct injection timing on HCCI combustion and emissions. The qualitative trends observed in the parametric variation study matched well with the experimental data in literature.

Zhi (2006) carried out a computational study of the direct injection gasoline HCCI engine with secondary injection. The detailed intake, spray, combustion and pollution formation processes of a compression ignition engine with high-octane fuel were studied by coupling the multi-dimensional

Computational Fluid Dynamic (CFD) code with detailed chemical kinetics.

An extended hydrocarbon oxidation reaction mechanism used for high-octane fuel was constructed and a modeling strategy of 3D-CFD/chemistry coupling for engine simulation was introduced to meet the requirements of the execution time acceptable to simulate the whole engine physicochemical process including intake, compression, spray and combustion process. The improved 3D CFD/chemistry model was validated using the experimental data from the HCCI engine with direct injection. Then, the CFD/chemistry model had been employed to simulate the intake, spray, combustion and pollution formation process of the gasoline direct injection HCCI engine with two-stage injection strategy. The models account for intake flow structure, spray atomization, droplet evaporation and gas phase chemistry in complex multi-dimensional geometries. The calculated results by Zhi (2006) showed that the periphery of a fuel-rich zone formed by the second injection ignited first, then the fuel-rich zone ignited and worked as an initiation to ignite the surrounding lean mixture zone formed by the first injection. The two-zone

HCCI leads to sequential combustion; this made ignition timing and combustion rate controllable. In addition, the HCCI load range can be

lxvi extended. However, the periphery of the fuel-rich zone leads to fierce burning, which results in slightly high NOx emissions.

Xing-Cai (2005) made a fundamental study of the control of the

HCCI combustion and emissions by the fuel design concept combined with controllable EGR. Xing-Cai (2005) investigated the basic combustion parameters including the start of the ignition timing, burn duration, cycle-tocycle variation, and Carbon monoxide (CO), Unburned Hydrocarbon (UHC), and Nitric Oxide (NOx) emissions of HCCI engines fueled with Primary

Reference Fuels (PRFs) and their mixtures. Two primary reference fuels, nheptane and iso-octane, and their blends RON25, RON50, RON75, and

RON90 were evaluated. The experimental results showed that, in the firststage of combustion, the start of ignition retards, the maximum heat release rate decreases, and the pressure rises and the temperature rising during the first-stage combustion decrease with the increase of the Research Octane

Number (RON). Furthermore, the cumulative heat release in the first-stage of combustion was strongly dependent on the concentration of n-heptane in the mixture. The start of ignition of the second-stage combustion was linear with the start of ignition of the first-stage. The combustion duration of the secondstage combustion decreases with the increase of the equivalence ration and the decrease of the octane number. The cycle-to-cycle variation improved with the decrease of the octane number.

Xing-Cai (2005) conducted a fundamental study of the control of the HCCI combustion and emissions by the fuel design concept combined with controllable EGR. The effect of the operating conditions and EGR on

HCCI combustion, the effects of the octane number of primary reference fuels and equivalence ratio on the combustion characteristics of a single-cylinder

HCCI engine was studied. In this part, the influence of the EGR rate, intake charge temperature, coolant temperature, and engine speed on the HCCI

lxvii combustion characteristics and its emissions were evaluated. The experimental results indicate that the ignition timing of the first-stage combustion and second-stage combustion retard, and the combustion duration prolongs with the introduction of the cooled EGR. At the same time, the

HCCI combustion using high cetane number fuels can tolerate a higher EGR rate, but only 45% of the EGR rate for RON75 at 1800 rpm. Furthermore, there was a moderate effect of the EGR rate on CO and UHC emissions for

HCCI combustion engines fueled with n-heptane and RON25, but a distinct effect on emissions for higher octane number fuels. Moreover, the combustion phase advances, and the combustion duration shortens with the increase of the intake charge temperature and the coolant out temperature, and the decrease of the engine speed. At last, it can be found that the intake charge temperature has the most sensitive influence on the HCCI combustion characteristics.

Magnus (2006) did an investigation into the lowest acceptable combustion temperatures for hydrocarbon fuels in HCCI engines. The combustion temperatures required for complete combustion in HCCI engines have been investigated computationally and experimentally. Fuels from several different hydrocarbon classes are covered including: iso-octane, n-heptane, toluene, and methylcyclohexane. Over a wide range of conditions, it was found that the temperature requirements are well described by the peak cycle temperature. For operation with a compression ratio of 18 at an engine speed of 1200 rpm, 1500 K is needed for the CO-to-CO

2

reactions to go to completion before the combustion is quenched by piston expansion. CO oxidation does not go to completion for operation with peak temperature below 1500 K since the OH level becomes too low. The required peak combustion temperature was found to be independent of the fuel-type and auto ignition characteristics. This occurs because the final CO oxidation process is independent of the original fuel molecule structure. Additionally, the combustion phasing relative to TDC surprisingly has no influence on the

lxviii required peak combustion temperature. This happens because the time spent close to the peak temperature does not change much with combustion phasing for an operation that has slow, but complete, CO oxidation. The onset of CO emissions is more gradual for the experiment compared to the single-zone model. This can be attributed to an in-cylinder thermal distribution caused by crevices and a thermal boundary layer.

Yap (2005) performed an investigation into a propane HCCI engine operation with residual gas trapping. Propane is available commercially for use in conventional internal combustion engines as an alternative fuel for gasoline. However, its application in the developing HCCI engines requires various approaches such as high compression ratios and/or inlet charge heating to achieve auto ignition. The approach documented here utilizes the trapping of internal residual gas (as used before in gasoline controlled auto ignition engines), to lower the thermal requirements for the auto ignition process. Yap (2005) carried out the experiment with a moderate engine compression ratio in the achievable engine load range that was controlled by the degree of internal trapping of exhaust gas supplemented by inlet charge heating. Increasing the compression ratio decreased the inlet temperature requirements; however, it also resulted in higher pressure rise rates. Varying the inlet valve timing affects the combustion phasing which can help to decrease the maximum pressure rise rates. NOx emissions were characteristically low due to the nature of the homogeneous combustion.

Magnus (2005) compared the late-cycle autoignition stability for single- and two-stage ignition fuels in HCCI engines. The characteristics of autoignition after TDC for both single- and two-stage ignition fuels have been investigated in an HCCI engine. The single stage ignition fuel was iso-octane and the two-stage ignition fuel was PRF80 (80% iso-octane and 20% nheptane). The results showed that both the heat-release rate and the pressure-

lxix rise rate decrease as the combustion is retarded later into the early expansion stroke. This is an advantage for high-load HCCI operation. However, for both fuel-types, cycle-to-cycle variations of the ignition and combustion phasing increase with combustion-phasing retard. Also, the cycle-to-cycle variations are higher for iso-octane compared to PRF80. These observations can be explained by considering the magnitude of random temperature fluctuation and the temperature-rise rate just prior to thermal run-away. Furthermore, too much combustion phasing retard leads to the appearance of partial-burn or misfire cycles, but the responses of the two fuels are quite different. The different behaviors can be explained by considering the thermal and chemical state of the residual exhaust gases that are recycled from one cycle to the next.

The data indicated that a partial burn cycle with iso-octane produces residuals that increase the reactivity of the following cycle. However, for the already more reactive PRF80 fuel, the partial-burn products present in the residuals do not increase the reactivity enough to overcome the retarding effect of cool residual gases.

Seref (2004) examined the combustion characteristics and phasing strategies of a natural gas HCCI engine; controlling the auto ignition timing over a wide range of speeds and loads was challenging. Overcoming this challenge in practical HCCI engines requires an improved understanding of the in-cylinder processes and how these processes can be favorably altered by various control techniques. In his study, a zero dimensional thermodynamic model that contains a simple heat release sub-model and an auto ignition model were used in a predictive fashion to better understand the in-cylinder processes and the efficiency potential of a natural gas engine in the HCCI mode. The model was also used for parametric studies to evaluate HCCI control strategies that can be tested on the research engine. The results indicated that if the initial conditions of the mixture are known precisely at intake valve closing, the auto ignition timing is controllable. A thermal

lxx efficiency close to 0.45 was possible with an IMEP range from 4 to 5 bars for the described engine, also.

Anthony (2004) attempted an HCCI combustion in which the effects of NOx in EGR experiments were studied for the oxidation of mixtures of n-heptane and iso-octane and of n-heptane and toluene in a jetstirred reactor (JSR) under dilute conditions, at 10 atm. The effect of the addition of variable initial NOx concentration was also studied. A detailed kinetic model was prepared to rationalize the results. Experiments were also performed using an HCCI engine to characterize the effect of EGR rates

(from 0% to 50%) with NOx addition (from 0 to 500 ppm) on ignition delays at low and high temperatures for an equivalence ratio of 0.3 and a constant intake temperature of 350 K. Two surrogate automotive fuels (n-heptane/isooctane, n-heptane/ toluene) were used and compared to the pure n-heptane case. Zero-dimensional single zone modeling was also performed using the detailed kinetic scheme and compared to the experimental results in terms of cool and principal flames ignition delays, phasing time and also the importance of the cool flame combustion heat release in comparison to the main one.

Dae (2005) improved the emission characteristics of an HCCI engine by various premixed fuels and cooled EGR. His work investigated partial HCCI combustion as a control mechanism. The premixed fuel is supplied via a port fuel injection system located in the intake port of a DI diesel engine. Cooled EGR was introduced for the suppression of advanced auto ignition of the premixed fuel. The premixed fuels used in this experiment are gasoline, diesel, and n-heptane. The results showed that with a diesel premixed fuel, a simultaneous decrease of NOx and soot can be obtained by increasing the premixed ratio. However, when the inlet charge is heated for the improved vaporization of diesel fuel, the higher inlet temperature limits

lxxi the operational range of the HCCI combustion due to severe knocking and high NOx emission at high premixed ratios. Gasoline premixing shows the most significant effects in the reduction of NOx and soot emissions, compared to other kinds of premixed fuels.

Panao (2005) interpreted the influence of fuel spray impact on mixture preparation for HCCI combustion with port-fuel injection. His work addressed the influence of fuel spray impact on fuel/air mixture for combustion in port-fuel injection engines. The experiments included time resolved measurements of the surface temperature synchronized with measurements of droplet dynamics at impact and were conducted to quantify the effects of the interactions between successive injections on the mixture preparation for combustion in Homogeneous Charge Compression Ignition

(HCCI) engines. The analysis showed that, during engine warm up, the heat transfer over the entire valve surface occurs within the vaporization-nucleateboiling regime and the local instantaneous surface temperature correlates with the dynamics of the droplets impacting at the same point. A functional relation was found for the heat transfer coefficient, which also described other experiments reported in the literature. Similarity does not hold after the engine warms up because the heat transfer and droplet vaporization at the surface were dominated by multiple interactions between the droplets arisen from diverse heat transfer regimes. However, results evidence the existence of a critical surface temperature which sets a transition between the overall heat transfer regimes dominated by local nucleate boiling at lower temperatures and by local intermittent transition regimes at higher temperatures. The heat transfer within the overall nucleate boiling regime was shown to be due to a thin film boiling mechanism leading to the breakdown of the liquid-film at a nearly constant surface temperature, regardless of the injection frequency or any other spray conditions. While at low frequencies this regime was not limited neither by the delivery of the liquid to the surface, nor by the removal

lxxii of the vapour from the surface, at higher frequencies it was triggered by enhanced vaporization induced by piercing and mixing the liquid film. The results further evidenced the important role of spray impingement for mixture preparation as required for HCCI.

Yap (2005) studied the natural gas HCCI engine operation with exhaust gas fuel reforming. Natural gas has a high auto-ignition temperature, requiring high compression ratios and/or intake charge heating to achieve

HCCI engine operation. It was shown that hydrogen in the form of reformed gas helped in lowering the intake temperature required for stable HCCI operation. It has been shown that the addition of hydrogen advances the start of combustion in the cylinder. This was the result of a lowering of the minimum intake temperature required for auto-ignition to occur during the compression stroke, resulting in advanced combustion for the same intake temperatures. Yap (2005) documents experimental results using closed loop exhaust gas fuel reforming for the production of hydrogen. When this reformed gas is introduced into the engine, a decrease in the intake air temperature requirement was observed for a range of engine loads. Thus, for a given intake temperature, lower engine loads can be achieved. This would translate to an extension of the HCCI lower load boundary for a given intake temperature.

Lei (2005) studied a low emission HCCI engine using combined internal and external EGR. The study focuses on the effects of internal and cooled external EGR on the combustion and emission performance of diesel fuel HCCI. The use of the fuel injection before the top center of an exhaust stroke and the Negative Valve Overlap (NVO) to form the homogeneous mixture achieves low NOx and smoke emissions. Internal and external EGR are combined to control the combustion. Internal Exhaust Gas Recirculation

(IEGR) helped to form a homogeneous mixture and reduce smoke emission

lxxiii further, through it lowered the high load limits of HCCI. Cooled external

EGR can delay the Start of Combustion (SOC) effectively, which is very useful for high cetane fuel (diesel) HCCI because these fuels can be easily self-ignited, making the SOC earlier. External EGR can avoid the knock combustion of HCCI at high load, which means it can expand the high load limit. HCCI maintains low smoke emission at various EGR rates and various loads compared to a conventional diesel engine, because there are no fuel-rich volumes in the cylinder.

Mingfa (2006) studied the controlling strategies of HCCI combustion with a fuel of dimethyl ether and methanol. The controlling strategies of HCCI fueled by Dimethyl ether (DME) and methanol were investigated. The experimental work was carried out on a modified singlecylinder diesel engine, which was fitted with the port injection of DME and methanol dual fuel. The results showed that the EGR rate and DME percentage are two important parameters to control the HCCI combustion process. The ignition timing and combustion duration can be regulated in a suitable range with high indicated thermal efficiency and low emissions by adjusting the DME percentage and EGR rate. EGR cannot extend the maximum Indicate Mean Effective Pressure (IMEP) of the HCCI operation range with dual fuel, but can enlarge the DME percentage range in normal combustion. The combustion efficiency depends largely on the DME percentage, and EGR can improve combustion efficiency. The results also showed that HC emissions depend strongly upon the DME percentage, and

CO emissions have good coherence to the peak mean temperature in the cylinder. In normal combustion, adopting a large DME percentage and high

EGR rate can attain an optimal HCCI combustion.

Ramanan (2006) discussed the effects of non-uniform temperature distribution on the ignition of a lean homogeneous hydrogen–air mixture. To

lxxiv characterize the ignition process in HCCI engines, high fidelity simulations are performed to study the effects of different initial temperature distributions on the auto ignition of a turbulent homogeneous mixture at high pressure. The effects of the initial temperature distribution on the ignition and the subsequent heat release are studied by the comparison of simulations with three initial random temperature fields having different skewness. It was found that the scalar mixing and turbulence have a significant influence on the initial location and further evolution of the ignition kernels. A comparison of the integrated heat release rates showed that the presence of a hot core leads to early ignition and increased duration of burning, while a cold core leads to a dormant end gas, which was consumed by slow combustion. The extent of the flame fronts was quantified by a temperature gradient cut-off, revealing a distinct behavior in the appearance of flame fronts for the three cases. Finally, two distinct ignition regimes, namely the spontaneous propagation and the deflagration regimes, are identified, and a predictive criterion is defined based on the spontaneous propagation speed and deflagration speed at the local mixture conditions. The predictions are found to be consistent with the observed results, suggesting a potential strategy in the modeling of the HCCI combustion process.

Ghassan (2006) conducted a factorial analysis of diesel engine performance using different types of bio-fuels. In his study, several biosource-fuels like fresh and waste vegetable oil and waste animal fat were tested at different injector pressures (120, 140, 190, 210 bar) in a directinjection, naturally aspirated, single-cylinder diesel engine with a design injection pressure of 190 bar. Using the 2k factorial analysis, the effect of injection pressure (P i

) and fuel type on three engine parameters, namely, combustion efficiency (Z c

), mass fuel consumption (m f

), and engine speed (N) was examined. It was found that P i

and fuel type significantly affected both Z c and mf while they had a slight effect on the engine speed. Moreover, with

lxxv diesel and bio-diesels, the Z c

increased to a maximum of 190 bar but declined at a higher P i

value. In contrast, higher Pi had a favorable effect on Z c

over the whole P i

range with all the other more viscous fuels tested. In addition, the mass fuel consumption consistently decreased with an increase in P i

for all the fuels including the baseline diesel fuel, with which the engine consistently attained higher Z c

and higher rpm compared to all the other fuels tested.

Rakopoulos (2007) studied the performance and emissions of a bus engine using blends of diesel fuel with bio-diesel of sunflower or cottonseed oils derived from greek feedstock. An experimental investigation is conducted to evaluate the use of sunflower and cottonseed oil methyl esters (bio-diesels) of greek origin as supplements in the diesel fuel at blend ratios of 10/90 and

20/80, in a fully instrumented, six-cylinder, turbocharged and after-cooled,

Direct Injection (DI), Mercedes-Benz, mini-bus diesel engine installed at the research laboratory. The tests were conducted using each of the above fuel blends, with the engine working at two speeds and three loads. Fuel consumption, exhaust smokiness and exhaust regulated gas emissions such as nitrogen oxides, carbon monoxide and total unburned hydrocarbons are measured. The differences in the measured performance and exhaust emissions from the baseline operation of the engine, i.e., when working with neat diesel fuel and the two bio-diesels are determined and compared. The theoretical aspects of diesel engine combustion with the differing physical and chemical properties of these blends, aid the correct interpretation of the observed engine behavior.

Deepak (2007) discussed the performance evaluation of a vegetable oil fuelled compression ignition engine. The fuel crises, because of a dramatic increase in vehicular population and environmental concerns, have renewed the interest of the scientific community to look for alternative fuels of bioorigin such as vegetable oils. Vegetable oils can be produced from forests,

lxxvi vegetable oilcrops, and oil bearing biomass materials. Non-edible vegetable oils such as linseed oil, mahua oil, rice bran oil, etc. are potentially effective diesel substitutes. Vegetable oils have a high-energy content. The study was carried out to investigate the performance and emission characteristics of linseed oil, mahua oil, rice bran oil and Linseed Oil Methyl Ester (LOME), in a stationary single cylinder, four stroke diesel engine, and to compare it with mineral diesel. The linseed oil, mahua oil, rice bran oil and LOME were blended with diesel in different proportions. The baseline data for diesel fuel was collected. Engine tests were performed using all these blends of linseed, mahua, rice bran, and LOME. Straight vegetable oils posed operational and durability problems when subjected to long-term usage in CI engines. These problems are attributed to the high viscosity, low volatility and polyunsaturated character of vegetable oils. However, these problems were not observed for LOME blends. Hence, the process of transesterification is found to be an effective method of reducing vegetable oil viscosity and eliminating operational and durability problems. An economic analysis was also done in this study, and it is found that the use of vegetable oil and its derivative as diesel fuel substitutes has an almost similar cost as that of mineral diesel.

Rakopoulos (2006) studied the comparative performance and emissions of a direct injection diesel engine using blends of diesel fuel with vegetable oils or bio-diesels of various origins. An extended experimental study was conducted to evaluate and compare the use of various diesel fuel supplements at blend ratios of 10/90 and 20/80, in a standard, fully instrumented, four stroke, DI, Ricardo/Cussons ‘Hydra’ diesel engine located at the research laboratory. More specifically, a high variety of vegetable oils or bio-diesels of various origins are tested as supplements, i.e. cottonseed oil, soybean oil, sunflower oil and their corresponding methyl esters, as well as rapeseed oil methyl ester, palm oil methyl ester, corn oil and olive kernel oil.

lxxvii

A series of tests were conducted using each of the above fuel blends, with the engine working at a speed of 2000 rpm, and with a medium and high load. In each test, the volumetric fuel consumption, exhaust smokiness and exhaust regulated gas emissions such as NOx, CO and THC were measured. From the first measurement, the SFC and brake thermal efficiency are computed. The differences in the measured performance and exhaust emission parameters from the baseline operation of the engine, i.e. when working with neat diesel fuel, are determined and compared. This comparison was extended to the use of vegetable oil and bio-diesel blends. The theoretical aspects of diesel engine combustion, combined with the widely differing physical and chemical properties of these diesel fuel supplements against normal diesel fuel, are used to aid the correct interpretation of the observed engine behavior.

Rakopoulos (2005) studied the multi-zone modeling of diesel engine fuel spray development with vegetable oil, bio-diesel or diesel fuels.

This work presented a model of fuel sprays development in the cylinders of diesel engines that are two-dimensional, multizone, with the issuing jet (from the nozzle) divided into several discrete volumes, called zones, formed along the direction of the fuel injection as well as across it. The model follows each zone, with its own time history, as the spray penetrates the swirling air environment of the combustion chamber before and after wall impingement.

After the jet break up time, a group of droplets was generated in each zone, with the model following their motion during heating, evaporation and mixing with the in-cylinder air. The model was applied to the interesting case of using vegetable oils or their derived bio-diesels as fuels, which recently were considered as promising alternatives to petroleum distillates since they are derived from biological sources. Although there are numerous experimental studies that show the curtailment of the emitted smoke with the possible increase of the emitted NOx against the use of diesel fuel, there is an apparent scarcity of theoretical models scrutinizing the formation mechanisms of

lxxviii combustion generated emissions when using these biologically derived fuels.

Thus, in the work, presented by Rakopoulos (2005) a theoretical detailed model of spray formation was developed that was limited to the related investigation of the physical processes by decoupling it from the chemical effects after combustion initiation. The analysis results showed how the widely differing physical properties of these fuels, against normal diesel fuel, affect greatly the spray formation, and consequently, the combustion mechanism and the related emissions.

Pradeep (2006) used the hot EGR for NOx control in a compression ignition engine fuelled with bio-diesel from jatropha oil; environmental degradation and depleting oil reserves are matters of great concern round the globe. Developing countries like India depend heavily on oil import. Diesel being the main transport fuel in India, finding a suitable alternative to diesel is an urgent need. Jatropha based Bio-Diesel (JBD) is a non-edible, renewable fuel suitable for diesel engines, and received increasing attention in India because of its potential to generate large-scale employment and relatively low environmental degradation. Diesel engines running on JBD are found to emit higher oxides of nitrogen, NOx. Hot EGR, a low cost technique of exhaust gas recirculation, was effectively used by Rakopoulos (2005) to overcome this environmental penalty. Practical problems, faced while using a cooled

EGR system, are avoided with hot EGR. Results indicated higher NOx emissions when a single cylinder diesel engine was fuelled with JBD, without

EGR. NOx emissions were reduced when the engine was operated under hot

EGR levels of 5–25%. However, the EGR level was optimized as 15% based on an adequate reduction in NOx emissions, minimum possible smoke, CO,

HC emissions and reasonable brake thermal efficiency. Smoke emissions of

JBD in the higher load region were lower than those of diesel, irrespective of the EGR levels. However, smoke emission was higher in the lower load region. CO and HC emissions were found to be lower for JBD irrespective of

lxxix the EGR levels. The combustion parameters were found to be comparable for both fuels.

Narayana (2005) conducted parametric studies for improving the performance of a jatropha oil-fuelled compression ignition engine. A single cylinder, constant speed, direct injection diesel engine was operated on neat

Jatropha oil. The injection timing, injector opening pressure, injection rate and air swirl level were changed to study their influence on performance, emissions and combustion. The results have been compared with those of neat diesel operation. The injection timing was varied by changing the position of the fuel injection pump with respect to the cam, and the injection rate was varied by changing the diameter of the plunger of the fuel injection pump. A properly oriented masked inlet valve was employed to enhance the air swirl level. Advancing the injection timing from the base diesel valve and increasing the injector opening pressure, increases the brake thermal efficiency and reduces HC and smoke emissions significantly. Enhancing the swirl has only a small effect on emissions. The ignition delay with jatropha oil is always higher than that of diesel under similar conditions. Improved premixed heat release rates were observed with jatropha oil when the injector opening pressure was enhanced. When the injection timing was retarded with an enhanced injection rate, a significant improvement in the performance and emissions was noticed. In this case emissions with jatropha oil are even lower than those of diesel. At full output, the HC emission level is 532 ppm with jatropha oil as against 798 ppm with diesel respectively. The NOx level and smoke with Jatropha oil are, respectively 1162.5 ppm and 2 BSU while they are 1760 ppm and 2.7 BSU with diesel, respectively.

Erol (2004) studied a diesel engine’s performance and exhaust emissions; he determines, using Artificial Neural-Networks (ANNs), the performance of exhaust emissions from a diesel engine with respect to

lxxx injection pressure, engine speed and throttle position. The design injectionpressure of the diesel engine, for the turbocharger and pre-combustion chamber used, is 150 bar. Experiments have been performed for four pressures, namely 100, 150, 200 and 250 bar with throttle positions of 25, 50,

75 and 100% respectively. Engine torque, power, brake mean effective pressure, specific fuel consumption, fuel flow, and exhaust emissions such as

SO

2

, CO

2

, NOx and smoke level (%N) have been investigated. The back propagation learning algorithm with three different variants, single and two hidden layers, and a logistic sigmoid transfer-function have been used in the network. In order to train the network, the results of these measurements have been used. Injection pressure, engine speed, and throttle position have been used as the input layer; performance values and exhaust emission characteristics have been used as the output layer. It is shown that the R2 values are about 0.9999 for the training data, and 0.999 for the test data; RMS values are smaller than 0.01; and mean % errors are smaller than 8.5 for the test data.

Senthil (2003) conducted an experimental comparison of methods to use methanol and jatropha oil in a compression ignition engine. In this work, various methods of using vegetable oil (Jatropha oil) and methanol such as blending, transesterification and dual fuel operation were studied experimentally. A single cylinder direct injection diesel engine was used for this work. Tests were done at a constant speed of 1500 rpm, and varying power outputs. In dual fuel operation the methanol to jatropha oil ratio was maintained at 3:7 on the volume basis. This was close to the fraction of methanol used to prepare the ester with jatropha oil. Brake thermal efficiency was better in the dual fuel operation and with the methyl ester of jatropha oil as compared to the blend. It increased from 27.4% with neat jatropha oil to a maximum of 29% with the methyl ester and 28.7% in the dual fuel operation.

Smoke was reduced with all methods compared to neat vegetable oil

lxxxi operation. The values of smoke emission are 4.4 Bosch Smoke Units (BSU) with neat jatropha oil, 4:1 BSU with the blend, 4 BSU with methyl ester of jatropha oil and 3:5 BSU in the dual fuel operation. The NOx level was lower with jatropha oil compared to that of diesel. It was further reduced in dual fuel operation and the blend with methanol. Dual fuel operation showed higher

HC and CO emissions than the ester and the blend. Ignition delay was higher with neat jatropha oil. It increased further with the blend and in dual fuel operation. It was reduced with the ester. The peak pressure and the rate of pressure rise were higher with all the methods compared to those of neat jatropha oil operation. Jatropha oil and methyl ester showed higher diffusion combustion compared to standard diesel operation. However, dual fuel operation resulted in higher premixed combustion. On the whole, it is concluded that the transesterification of vegetable oils and methanol induction can significantly enhance the performance of a vegetable oil fuelled diesel engine.

Carraretto (2006) discussed bio-diesel as an alternative fuel. He presented the first results of an investigation carried out on the potentialities of bio-diesel as an alternative fuel based on strategic considerations and field experiences in boilers and diesel engines. The operation of a bio-diesel fuelled boiler had been checked for some months. The engines were bench-tested and then installed in urban buses for normal operation. Distances, fuel consumption and emissions (CO

2

, CO, HC and NOx) were monitored; in addition, wear and tear, oil and air filter’s dirtiness and lubricant degradation have also been checked. Further investigations have also been devoted to assess some environmental aspects of bio-diesel. In particular the benefit of bio-diesel to the total net emission of CO

2

during the whole life cycle has been studied and the net energy requirement has been evaluated. Finally, the global environmental support to the production of bio-diesel has been studied according to the energy analysis.

lxxxii

Mustafa (2005) discussed the performance and exhaust emissions of a bio-diesel; engine in his study, the applicability of Artificial Neural

Networks (ANNs) has been investigated for the performance and exhaustemission values of a diesel engine fueled with bio-diesels from different feed stocks and petroleum diesel fuels. The engine performance and emissions characteristics of two different petroleum diesel-fuels, bio-diesel (from soybean oil and yellow grease), and their 20% blends of diesel fuels were used. The fuels were tested at full load (100%) at 1400-rpm engine speed, where the engine torque was 257.6 Nm. To train the network, the average molecular weight, net heat of combustion, specific gravity, kinematic viscosity, C/ H ratio and cetane number of each fuel are used as the input layer, while the outputs are the brake specific fuel-consumption, exhaust temperature, and exhaust emissions. The back-propagation learning algorithm with three different variants, single layer, and logistic sigmoid transfer function were used in the network. By using weights in the network, formulations have been given for each output. The network has yielded the

R2 values of 0.99 and the mean % errors are lesser than 4.2 for the training data, while the R2 values are about 0.99 and the mean % errors are lesser than

5.5 for the test data. The performance and exhaust emissions from a diesel engine, using bio-diesel blends diesel fuel up to 20%, have been predicted using the ANN mode.

Deepak (2006) performed the experimental investigation of the control of NOx emissions in a bio-diesel-fueled compression ignition engine.

Bio-diesel was an alternative fuel consisting of the alkyl esters of fatty acids from vegetable oils or animal fats. Vegetable oils are produced from numerous oil seed crops (edible and non-edible), e.g., rapeseed oil, linseed oil, rice bran oil, soybean oil, etc. Research has shown that bio-diesel-fueled engines produce less CO, HC, and PM but higher NOx emissions compared to those of mineral diesel fuel. EGR is effective to reduce the NOx from diesel

lxxxiii engines, because it lowers the flame temperature and the oxygen concentration in the combustion chamber. However, EGR results in higher

PM emissions. Thus, the drawback of higher NOx emissions while using biodiesel may be overcome by employing EGR. The objective of his research work was to investigate the usage of bio-diesel and EGR simultaneously in order to reduce the emissions of all regulated pollutants from diesel engines.

A two-cylinder, air-cooled, constant speed direct injection diesel engine was used for the experiments. HCs, NOx, CO, and the opacity of the exhaust gas were measured to estimate the emissions. Various engine performance parameters such as thermal efficiency, BSFC, and Brake Specific Energy

Consumption (BSEC), etc. were calculated from the acquired data. The application of EGR with bio-diesel blends resulted in the reduction of NOx emissions without any significant penalty in PM emissions or BSEC.

Breda (2007) studied the effects of bio-diesel on the emissions of a bus diesel engine. He discussed the influence of bio-diesel on the injection, spray, and engine characteristics, to reduce harmful emissions. The considered engine was a bus diesel engine with an injection system. The injection, fuel spray, and engine characteristics, obtained with bio-diesel, are compared to those obtained with mineral diesel under various operating regimes. The considered fuel was neat bio-diesel from rapeseed oil. Its density, viscosity, surface tension, and sound velocity are determined experimentally and compared to those of mineral diesel. The obtained results were used to analyze the most important injection, fuel spray, and engine characteristics. The injection characteristics were determined numerically under the operating regimes, corresponding to the mode ESC test. The fuel spray was obtained experimentally under peak torque condition. Engine characteristics were determined experimentally under mode ESC test conditions. The results indicate that, by using bio-diesel, harmful emissions

lxxxiv

(NOx, CO, smoke and HC) can be reduced to some extent by adjusting the injection pump timing properly.

Scientific research has projected a significant increase in the amount of major air pollutants in the decades to come, due to the increasing number of internal combustion engine production and usage. To avoid the potential problems related with air pollutants, the U.S. government through the EPA enforced the limit on the maximum amount of pollutants an engine was allowed to generate. The regulations may get more stringent in the years to come. Engine researchers throughout the world have focused on improving engine performance and efficiency while reducing the amount of pollutants generated in order to meet the more stringent regulations. Most of the research reported in the literature looked typically at very specific implementation issues. The literature currently lacks a fundamental study of the use of Jatropha oil and secondary injection port in HCCI engines to study the characteristics of pollutants present in the exhaust gas, and the fuel economy. The research reported in this thesis aims to remedy this deficiency.

2.8 CLOSURE

The literature review shows that the HCCI mode operation in IC engines reduces the emissions, which is the most important criterion for the current environmental norms, and also, the study found that the HCCI mode also decreases the overall fuel consumption. Several researchers have conducted HCCI with different control techniques; still there is a good deal of research work going on in this area for optimizing several factors. There are a limited number of research works available related to the usage of bio-fuels in

HCCI mode operations with PCCI control. In this regard this research work aims to study the emissions, efficiency and fuel consumption in the HCCI mode with conventional diesel and bio diesel as a primary fuel, and diesel, petrol and bio-diesel as a secondary fuel using the PCCI mode.

2.9

lxxxv

MOTIVATION AND OBJECTIVE OF THIS WORK

The objectives of this research work are:

To analyze the efficiency and emissions of a 4 stroke single cylinder DI diesel engine in the HCCI mode using PCCI control and comparing the same with the conventional mode.

To compare the efficiency and emissions with diesel and bio diesel as primary fuel. Diesel, petrol and bio diesel are used as secondary fuel.

SCOPE OF THE STUDY 2.10

Design and fabrication of the pilot injector with a control and measurement unit.

Attaching the pilot injector to the conventional diesel engine.

Conducting the experiment under different operating modes and measuring the variable emission factors and parameters which influence the efficiency and specific fuel consumption.

Comparing the conventional mode with the HCCI mode for efficiency and emission factors.

Comparing the HCCI mode with different primary and secondary fuels.

The experimental set up for the present research work has been explained in detail in chapter 3.

lxxxvi

CHAPTER 3

EXPERIMENTAL SETUP

3.1 INTRODUCTION

The engine experiments are conducted to validate the application of the analysis and to implement strategies for HCCI startability and control. In the present research work, an agricultural water-cooled single cylinder engine which is very well-suited for HCCI operation has been used. Experiments are conducted for different load conditions and different fuel combinations. Based on the results obtained from the experiments, the performance and emission graphs are plotted. The detailed description of the experimental setup, properties of the samples and experimental procedures are discussed in this chapter.

3.2 EXPERIMENTAL APPARATUS

The research engine was based on a single cylinder, Direct

Injection (DI), and four-stroke cycle diesel engine with 0.553 lit of piston displacement (agricultural type water cooled). The specifications and dimensions of the test engine are listed in Table 3.1. The experimental apparatus consists of the primary fuel supply system, the exhaust gas analyzer system, the thermostat air temperature control system and the pilot injection system. A gravity feed fuel system with an efficient paper element filter, and force feed lubrication is also used. Large end bearings and camshaft bush,

lxxxvii suitable for run-through or thermosyphon cooling are also provided in the present system.

Table 3.1 Specifications of the experimental setup

S.No. Details

1 Number of cylinders

2 Bore × Stroke

3 Cubic Capacity

4 Compression Ratio

5 Rated Output as per

BS5514/ISO 3046/IS 10001

One

Specifications

80 × 110 mm

0.553 lit

16.5

3.7 kW (5.0 hp) at 1500 rpm.

6 SFC at rated hp/1500 rpm

7 Lube Oil Consumption

8 Lube Oil Sump Capacity

9 Fuel Tank Capacity

245 g/kWh(180 g/bhp-hr)

1.0% of SFC max.

3.3 lit.

6.5 lit

10 Fuel Tank re-filling time period Every 6 hours of engine running at rated output

114 kg 11 Engine Weight (dry) without flywheel

12 Weight of flywheel 33kg - Standard

13 Rotation while looking at the flywheel

14 Power Take-off

15 Starting

Clockwise.( Optional –

Anticlockwise)

Flywheel end. (Optional-Gear end half or full speed)

Hand start with cranking handle.

Figure 3.1 shows the schematic diagram of the experimental apparatus used for the present research work. The pilot injection system is attached to the air manifold to supply the secondary fuel in the HCCI mode.

The pilot injection system consists of a pilot injection valve and fuel

lxxxviii measurement system. The heating and thermostat arrangements are also attached to the air manifold to preheat and measure the temperature of the air respectively. Diesel, petrol and bio diesel are used as fuels in this research work.

1. Engine

2. Cylinder

3. Brake dynamometer

4. Inlet valve

5. Outlet valve

6. Pilot injector

7. Inlet manifold

12. Air tank

(With heating and thermostat arrangement)

13. Manometer

14. Orifice meter

15. Primary fuel injector

16. Primary fuel injector pump

8. Outlet manifold

9. Secondary injection fuel pump

10. Secondary Fuel Flow Meter

11. Secondary fuel tank

17. Primary fuel flow meter

18. Primary fuel tank

19. Thermocouple

20. Exhaust gas analyzer and

(With heating and thermostat arrangement) 21. Smoke meter

lxxxix

Figure 3.1 Schematic diagram of the experimental setup

The secondary fuel is heated by means of a 1KW water bath provided with an electrical thermostat which is used to maintain the fuel at the desired temperature. The maximum fuel temperature that can be achieved using the current setup is 75°C. This temperature is maintained at 55°C and

75°C for diesel and bio diesel respectively. The air is heated by means of an

800 W air heater placed in the inlet manifold. The air temperature is maintained by means of an electrical thermostat to a maximum of about 80°C.

The secondary fuel is injected into the inlet manifold near the inlet valve using an electric fuel pump through a pilot fuel injector mounted on the inlet manifold whose spray angle is 30° with the inlet air flow and flow rate of

7.2 ml/min. The fuel line pressure is maintained at 6 bars by the fuel injection pump. The current rating of the fuel injection pump is 20 Amps and that of the fuel injector is 0.3 Amps. The valve timing diagram is shown in

Figure 3.2.

xc

Figure 3.2 Valve timing diagram

The pilot fuel injection is controlled by an electronic circuit having a limit switch with a frequency of about 750 cycles per minute. The limit switch is actuated by means of a bolt attached to the inlet valve rocker arm.

An 8 mm travel of the rocker arm is used to control the electrical signal for initiating and closing the secondary injection during the suction stroke. The electrical signal is produced when the rocker arm is positioned at 4.84 mm from the top most position. This signal is used to open the secondary fuel injection through various control devices. When the rocker arm is positioned at 1.97 mm from the bottom most position, the electrical signal is sent to the injector for closing the secondary injection. A photographic view of the experimental setup for the present research work is shown in Figure 3.3.

xci

Figure 3.3 Photographic view of the experimental setup

SAMPLE FUELS 3.3

Diesel, petrol and bio diesel are used for this research work as sample fuels. The specifications of these samples are identified by standard test methods. The properties of diesel, petrol and bio-diesel are listed in

Table 3.2.

Table 3.2 Properties of diesel, petrol and bio-diesel

Properties

Density 15

0

C Kg/m

3

Octane Number

Cetane Number

Kinematic Viscosity at 40°C (m

2

/s)

Flash point °C

Specific gravity

Auto-ignition Temperature (°C)

Molecular Weight

Bio-diesel Diesel Petrol

875 830 760

- - 91

46

4

51

1.3

-

0.55

120

0.88

273

292

55

0.85

263

200

40

0.75

280

103

3.4 INJECTOR DETAILS

The details of the injector assembly and its components are illustrated in Figure 3.4. The photographic view of the primary and pilot fuel injector are shown in Figures 3.5 and 3.6 respectively.

xcii

Injector Nozzle Nuts

Injector Intermediate Plates

Injector Washers

Injector Shims

Figure 3.4 Injector assembly and components of the primary fuel injector

Figure 3.5 Photographic View of the Primary Fuel Injector

xciii

Figure 3.6 Photographic View of the Pilot Fuel Injector

3.5 INJECTION TIMING

The details of the pilot fuel injection timing are as follows:

Circumference of the flywheel : 126.5 cm

Radius of the flywheel : 20.13 cm

The suction stroke duration : 219°

The exhaust stroke duration : 219°

The inlet valve opens at 4.5° before TDC during the exhaust stroke

The inlet valve closes at 34.5° after BDC during the compression stroke

The exhaust valve opens at 34.5° after BDC during the expansion stroke

xciv

The exhaust valve closes at 4.5° after TDC during the exhaust stroke

The manifold injection starts at 61.239° after TDC during the suction stroke

The manifold injection closes at 51.239° before BDC during the suction stroke

The injection duration is for 87.52° during the suction stroke, for about 2/5 of the suction stroke.

3.6 BACKGROUND OF THE ENGINE

In the present research work, an agricultural water cooled single cylinder engine has been used. In the present setup, a four stroke DI engine with a wet sump type lubricating system is also used.

3.6.1 General Characteristics of a DI Single Cylinder Research

Engine

In order to perform the HCCI diesel experimentation, a single cylinder research engine is used. This engine is a single cylinder four stroke engine, with a variable compression ratio and overhead poppet valves mounted on a common bed plate with an electrical dynamometer. The engine is designed for teaching and research purposes in the field of reciprocating IC engines. The engine can be used as an SI engine, as well as a CI engine by making simple modifications. For the scope of the present experimentation, the engine is used as a CI engine (diesel mode).

3.6.2 Functioning Mode

During the compression stroke, air is forced into the chamber. On the way to the chamber, the air picks up a certain amount of heat from the

xcv insulated lower half of the combustion chamber, thereby raising the compression temperature without any loss of volumetric efficiency. The fuel is then injected into the chamber via a high pressure injector.

The cast-iron cylinder is fitted with a hardened high-phosphorus cast-iron liner. The outside surface of the cylinder water jacket slides in a cast-iron guide and the lower portion is screwed to accommodate a nut with worm teeth on its periphery. The rotation of this nut by means of a worm raises and lowers the cylinder relative to the crankshaft, and thus it is possible to vary the compression ratio. The cylinder jacket can be locked by a clamp bolt which closes the guide and so grips the cylinder uniformly around the whole circumference of the jacket.

The lubrication system is of the wet sump type, oil being delivered to the crankshaft and big-end bearings by a gear pump separately driven by an electric motor. The pump is fitted with an adjustable relief valve to control the delivery pressure, while a full-flow filter is fitted to ensure a clean supply of oil to all the bearings. An oil pipe leads up to the top of the cam box and supplies oil at each contact friction between the camshaft and the cylinder head, by means of a timed feed.

3.7 EXPERIMENTAL PROCEDURE

In this research work, experiments are conducted in the conventional and HCCI mode with all the fuel samples. Initially, the engine is operated in the conventional mode using diesel and bio-diesel as a primary fuel; there is no secondary fuel supply. The efficiency, fuel consumption, exhaust gas temperature and emissions are measured under different load conditions. The different percentages of varying loads are: 0%, 20%, 40%,

60%, 80% and 100%.

xcvi

In the HCCI mode, the pilot injector is used to supply secondary fuel to the combustion chamber to get the Premixed Charged Combustion

Ignition (PCCI).The pilot injector works by means of the signal from the rocker ram. The opening and closing of the pilot injector for supplying the secondary fuel depends upon the position of the rocker arm, which depends upon the position of the piston inside the combustion chamber. The air is heated to a temperature of 80°C before it is sent into the combustion chamber.

The primary fuel is supplied to the combustion chamber without heating. In the case of the secondary fuel supply, the fuel is heated before entering the inlet manifold. The heating is done by an electrical heating arrangement. The secondary fuel is supplied into the combustion chamber in the suction stroke.

This will create PCCI during the compression stroke and help to ignite the fuel which is supplied by primary injection and get the HCCI. This HCCI helps to get complete combustion and uniform low temperature throughout the combustion chamber.

The exhaust gas is analyzed by means of a five-gas analyzer to measure the HC, CO, CO

2

, NOx and O

2

present in the exhaust gas. The exhaust gas temperature and smoke contents are measured by means of the thermocouple and smoke meter respectively. The primary fuel, secondary fuel, and air quantity are measured before they enter the combustion chamber.

The experiments have been conducted in the following modes:

3.7.1 Mode 1 - Diesel Conventional

In this mode the experiments are conducted only with primary injection. The diesel is used only as a fuel for combustion. The pilot injector is closed in this mode; there is no fuel supply through the pilot injector. This experiment measures the emissions and fuel consumption in the conventional mode, by preheating the air before supplying it to the combustion chamber.

xcvii

3.7.2 Mode 2 - Bio-Diesel Conventional

In this mode, Bio-diesel (Jatropha oil) is used as a fuel instead of diesel, as in the case of mode 1.

3.7.3 Mode 3 - Diesel-Diesel PCCI-DI Combustion

In this mode of experiment, both the primary and pilot injectors are used to supply the fuel to the combustion chamber. Diesel is used as a fuel in both the injectors. The pilot injector supplied the diesel to the combustion chamber before the actual injection of the diesel into the combustion chamber through the primary injection port. The emissions and fuel consumptions are measured in this mode. The PCCI is created by supplying the fuel through pilot injector after the piston reaches the BDC. The procedure followed in mode 3 has also been followed for all the other modes of experiments which are discussed below.

3.7.4 Mode 4 - Bio-Diesel - Bio-Diesel PCCI-DI Combustion

In this mode of experiment, both the primary and pilot injectors are used to supply Bio-diesel. The pilot injector supplied the bio-diesel to the combustion chamber before the actual injection of the bio-diesel into the combustion chamber through the primary injection port. The emissions and fuel consumptions are measured in this mode.

3.7.5 Mode 5 - Diesel - Bio-Diesel PCCI-DI Combustion

In this mode of experiment, both diesel and bio-diesel are used as primary and secondary fuel respectively. The diesel is injected into the cylinder through the primary injector; the injected diesel is compressed to increase the pressure. The pilot injector supplied the bio-diesel to the combustion chamber before the actual injection of the diesel into the

xcviii combustion chamber through the primary injection port. The emissions and fuel consumptions are measured in this mode.

The same procedure is adopted for different combinations of biodiesel, diesel and petrol combinations such as Mode 6 - Bio-diesel - Diesel

PCCI-DI combustion, Mode 7- Diesel - Petrol PCCI-DI combustion, and

Mode 8- Bio-diesel - Petrol PCCI-DI combustion.

3.8 EXHAUST GAS ANALYSER

The use of a four or five gas exhaust analyzer can be helpful in troubleshooting both emissions and driveability concerns. Presently, shop grade analyzers are capable of measuring from as few as two exhaust gases,

HC and CO, to as many as five. The five gases measured by the latest technology exhaust analyzers are: HC, CO, CO

2

, O

2

and NOx. A photographic view of the exhaust gas analyzer showing a sample result for the present research work is shown in Figure 3.7 and the photographic view of the exhaust gas analyzer used is shown in Figure 3.8.

Figure 3.7 Photographic view of the exhaust gas analyzer showing a sample result

xcix

Figure 3.8 Photographic view of the exhaust gas analyzer

3.8.1 Specifications of the Exhaust Gas Analyzer

A MRU DELTA 1600L five gas analyzer is used for the present research work. Its detailed specifications are listed in Table 3.3.

Table 3.3 Specifications of Exhaust Gas Analyzer

S.No. Details

Exhaust Gas Analyzer Measuring Ranges

Specifications

0 – 25.00% vol 1 Oxygen (O

2

)

2 Carbon monoxide (CO)

3 Carbon dioxide (CO

2

)

4 Hydro carbon (HC)

0 – 15.00% vol

0 – 20.00% vol

0 – 20,000 ppm n-hexane

5 Nitrogen oxide (NOx)

6 Excess Air calculated

According to Brett Schneider

Temperature

7 Rounds per minute

0 – 2,000 ppm

-40°C to +650°C

400 to 10,000 U/min

c

Table 3.3 (Continued)

Specifications S.No. Details

Precision

8 Oxygen (O

2

)

9 Carbon monoxide (CO)

10 Carbon dioxide (CO

2

)

11 Hydro carbon (HC)

12 Nitrogen oxide (NOx)

13 Temperature (T>250 o

C)

14 Rounds per minute

Resolution

15 Oxygen (O

2

)

16 Carbon dioxide (CO

2

)

17 Carbon monoxide (CO)

18 Hydro carbon (HC)

19 Nitrogen oxide (NOx)

20 Rounds per minute

21 Response time T95

22 Mains supply

23 Power consumption

Operating conditions

24 Temperature

25 Humidity

26 Display

27 Tubes

Case

28 Type

29 Dimension

0.1% or 3%

0.06% or 5% of measured value

0.5% or 5% of measured value

12 ppm or 5% of measured value

5 ppm or 5% of measured value

1% (T<150 o

C) 2% (T<250 o

C)

3%

1%

0.01%

0.1%

0.01%

1 ppm

1 ppm

1 U/min (<6000) or 10 U/min (>6000)

15s

110 – 230 V 50/60 Hz or 12 V dc max 100 VA

+5°C to +40°C

0 – 90% v H (not condensing)

8 – line big – display, illuminated

Black Vinton tube, optionally for heating tube

Robust aluminum case with swing – out front

560*330*215 mm

ci modes:

The experiments are conducted in the following eight different

Diesel conventional, Bio-diesel conventional, Diesel - Diesel

PCCI-DI combustion, Bio-diesel - Bio-diesel PCCI-DI combustion, Diesel -

Bio-diesel PCCI-DI Combustion, Bio-diesel - Diesel PCCI-DI combustion,

Diesel - Petrol PCCI-DI combustion, and Bio-diesel - Petrol PCCI-DI combustion. Various parameters such as exhaust gas temperature, HC, NOx,

CO, SFC, BTE, and CO

2 are measured. A detailed analysis and discussion of the results obtained for the above mentioned parameters is included in chapter 4.

cii

CHAPTER 4

RESULTS AND DISCUSSION

4.1 INTRODUCTION

Several experiments have been conducted on the present experimental setup in order to assess the performance of the developed HCCI system. The effects of the load on parameters such as exhaust gas temperature, hydro carbon in exhaust, NOx emission in exhaust gas, specific fuel consumption, brake thermal efficiency, carbon dioxide, and carbon monoxide emission are discussed in this chapter. A detailed description of the graphs comparing: the conventional and HCCI methods, different fuels in the

HCCI mode, variation of the exhaust gas temperature with different fuels, variation of the exhaust gas temperature with load with different fuels, hydrocarbon emissions in the conventional and HCCI methods, hydrocarbon emissions from different fuels in the HCCI mode, variation of hydro carbon with different fuels, NOx emissions in the conventional and HCCI methods,

NOx emissions from different fuels in the HCCI mode, variation of NOx emissions in different fuels, NOx emissions with different loads in different

HCCI modes, specific fuel consumption between the conventional and HCCI methods, specific fuel consumption between different fuels in the HCCI mode, variation of the specific fuel consumption with different fuels, specific fuel consumption with different loads in different HCCI modes, brake thermal efficiency between the conventional and HCCI methods, brake thermal efficiency with different fuels in HCCI modes, brake thermal efficiency with different load conditions for all the modes of operations,

ciii variation of brake thermal efficiency with different fuels, brake thermal efficiency with different loads in different HCCI modes, CO

2 emission in the conventional and HCCI methods, CO

2

emission between different fuels in the

HCCI mode, variation of CO

2

emission with different fuels, brake thermal efficiency with different loads in different HCCI modes, CO emission in the conventional and HCCI methods, CO emission in the different fuels in the

HCCI mode, and variation of CO emission with different fuels is also included in this chapter.

4.2 EFFECT OF THE LOAD ON EXHAUST GAS

TEMPERATURE

The effect of the load on the exhaust gas temperature is measured by a thermocouple attached to the outlet manifold. The variation of temperature with the increase in load for all the cases is measured. Figure 4.1 shows the variation of temperature with an increase in the load for the conventional mode with diesel, bio-diesel and HCCI with the diesel - diesel and bio-diesel - bio-diesel combination mode.

350

300

250

200

150

100

0 20 40 60

Load (%)

Diesel

Bio Diesel

Diesel - Diesel

Bio Diesel - Biodiesel

80 100

Figure 4.1 Comparison of the conventional and HCCI methods

civ

The exhaust gas temperature increased with the increase in the load in all the cases. This could be due to more amount of fuel combustion inside the combustion chamber at a higher load as compared to a lower load and complete combustion of the fuel. Among the four modes, the exhaust gas temperature from the conventional diesel mode is higher than that of the other modes, and that the bio-diesel - bio-diesel mode is lower than that of the other modes for all the load conditions. Table 4.1 shows the variation of the exhaust gas temperature with diesel, and bio-diesel in the conventional mode and the same fuel in the HCCI mode.

Table 4.1 The variation of the Exhaust gas temperature (°C) with an increase of load

Load %

0

20

40

60

80

100

Diesel

213

231

247

265

285

315

Bio-diesel

207

223

239

256

279

306

Diesel -

Diesel

167

202

224

242

258

278

Bio-diesel -

Bio-diesel

143

192

215

223

254

270

For all the load condition, the HCCI mode of operations produced lower exhaust gas temperature compared to the same fuel in the conventional mode. This could be due to the uniform combustion of fuel inside the combustion chamber in the case of the HCCI mode. For 80% load conditions, the exhaust gas temperature of 285°C and 258°C was observed for the conventional diesel mode and diesel with the HCCI mode respectively. The exhaust gas temperature was 270°C and 254°C in the case of bio-diesel in the conventional and HCCI modes respectively.

cv

As compared to the conventional modes with diesel and bio-diesel fuels, diesel and bio-diesel in the HCCI mode produced a lower temperature.

For the maximum load conditions, the exhaust gas temperature from diesel, bio-diesel, diesel - diesel, and bio-diesel - bio-diesel are 315°C, 306°C, 278°C and 270°C respectively.

Figure 4.2 shows the variations of the exhaust gas temperature with an increase of load in the HCCI mode with different primary and secondary fuel supply. Among these, bio-diesel and petrol used with primary and secondary fuels produced a lower exhaust gas temperature as compared to the other three cases such as diesel - bio-diesel, bio-diesel - diesel, diesel - petrol.

The HCCI with different primary and secondary fuels produced lower exhaust gas temperature as compared to the conventional mode.

350

300

250

200

150

Diesel - Biodiesel

Biodiesel - Diesel

Diesel - Petrol

Biodiesel - Petrol

100

0 20 40 60 80 100

Load (%)

Figure 4.2 Comparison between different fuels in the HCCI mode

Table 4.2 shows the variation of the exhaust gas temperature with diesel, bio-diesel and petrol combinations in the HCCI mode of operation.

The lowest exhaust gas temperature was observed for the bio-diesel - petrol combination in the HCCI mode as compared to other modes. For 80% load

cvi condition, the bio-diesel - petrol HCCI mode reduced the exhaust gas with the temperature of 241°C.

Table 4.2 The variation of Exhaust gas temperature (°C) with an increase of load

Load %

0

20

40

60

80

100

Diesel - Biodiesel

149

192

250

275

274

281

Bio-diesel -

Diesel

178

223

272

278

280

285

Diesel -

Petrol

138

179

211

231

255

261

Bio-diesel -

Petrol

140

167

200

225

241

254

Figure 4.3 shows the maximum exhaust gas temperature obtained from both the conventional and HCCI mode of operations with all the fuel samples in this research work. The lowest exhaust gas temperature of 223°C is observed for the bio-diesel - bio-diesel combinations in the HCCI mode of operation, the next lowest temperatures of 225°C and 231°C respectively were observed for the bio-diesel - petrol and diesel - petrol combinations respectively. The bio-diesel - diesel combinations in the HCCI mode produced the maximum exhaust gas temperature compared to the other fuels.

The maximum temperature of 278°C was observed for this sample fuel.

Figure 4.4 shows the variation of the exhaust gas temperature with bio-diesel as a primary fuel and bio-diesel, diesel and petrol as a secondary fuel in the HCCI mode of operation. The exhaust gas temperature for the biodiesel - bio-diesel fuel is between that of diesel and petrol as a secondary fuel.

Diesel as a secondary fuel produced higher exhaust gas temperature as compared to the other two cases.

cvii

300

250

200

150

100

50

0

265

256

242

223

275

278

231

225

Type of Fuel

Figure 4.3 Variation of the exhaust gas temperature (°C) with different fuels as input

300

250

200

150

100

50

0

0 20

Bio Diesel - Biodiesel

Biodiesel - Diesel

Biodiesel - Petrol

40

Load (%)

60 80 100

Figure 4.4 Variation of the exhaust gas temperature with load with different fuels as input

4.3

cviii

EFFECT OF LOAD ON HYDRO CARBON IN EXHAUST

GAS

The hydrocarbon emission from all the modes is observed with different load conditions. Figure 4.5 represents the variation of HC emission with the increase of load from 0% to 100%. The study found that the HC emission from the conventional diesel mode is higher than that of the other modes, and also the HC emission from the HCCI mode is lower than that of the conventional mode. The HC emission increased with the increase of load for conventional mode and decreased with the increase of load for HCCI mode. Among the HCCI mode, the HC emission from the bio-diesel - biodiesel fuel is observed to be lower compared to that of the diesel - diesel fuel for all the load conditions.

120

100

80

60

40

20

0

Diesel

Biodiesel

Diesel - Diesel

Biodiesel - Biodiesel

0 20 40 60

Load (%)

80 100

Figure 4.5 Comparison of hydrocarbon emission between the conventional and HCCI methods

The HC emissions with diesel and bio-diesel as a fuel in both the conventional and HCCI mode are listed in Table 4.3. The study found the conventional mode of operation producing more amounts of HC emissions as

cix compared to those of the HCCI mode. This could be due to complete combustion in the case of the HCCI mode due to the uniform mixture of fuel in the combustion chamber. For the 80% load condition, the HC emissions observed are 103, 51, 25 and 16 ppm with diesel and bio-diesel in the conventional, and diesel – diesel, and bio-diesel - bio-diesel in the HCCI mode respectively.

Table 4.3 Variation of HC (ppm) emission for different loads with different fuels as input

Load %

0

20

40

60

80

100

Diesel

55

85

89

98

103

108

Bio-diesel

43

51

60

35

39

41

Diesel-

Diesel

36

33

30

28

25

24

Bio-diesel-

Bio-diesel

24

30

24

14

16

11

In the conventional mode the HC emission increased from 55 ppm to 108 ppm when the load is increased from 0% to 100%. For the same load interval for bio-diesel – bio-diesel, the emission is observed with the decrement from 24 ppm to 11 ppm.

Figure 4.6 shows the variation of HC emission with the increase of load in the HCCI mode with different fuels as primary and secondary fuels. In all the cases of the HCCI mode the HC emission decreased with the increase of load. Among all the HCCI modes diesel - bio-diesel produced lower HC emissions and bio-diesel - diesel produced higher HC emissions.

cx

160

140

120

100

80

60

40

20

0

Diesel - Biodiesel

Biodiesel - Diesel

Diesel - Petrol

Biodiesel - Petrol

0 20 40 60 80 100

Load (%)

Figure 4.6 Comparison of hydrocarbon emission between different fuels in the HCCI mode

The HC emission with diesel, bio-diesel and petrol combinations in the HCCI mode of operations is listed in Table 4.4. The study observed that the diesel - bio-diesel mode of operation produces lower HC emissions as compared to the other HCCI mode for all the load conditions. The HC emission has decreased with the increase of load conditions for all the HCCI modes except the diesel - bio-diesel mode. In the diesel - bio-diesel mode, a mixed tendency was observed.

Table 4.4 Variation of HC (ppm) emission with different loads with different fuels as input

Load %

0

20

40

60

80

100

Diesel -

Bio-diesel

21

22

36

30

45

35

Bio-diesel -

Diesel

138

95

89

86

75

61

Diesel -

Petrol

84

63

60

46

42

38

Bio-diesel -

Petrol

87

71

66

54

49

43

cxi

Figure 4.7 shows huge variations of HC emissions with the variation of fuel sample and also of operation modes. The lowest HC emission is observed for bio-diesel - bio-diesel in the HCCI mode of operations. The highest HC emission is observed for diesel in the conventional mode of operations. Bio-diesel in the conventional mode of operations produced lower HC emissions as compared to diesel in the conventional mode.

120

100

80

60

40

20

0

98

43

28

14

30

86

46

54

Fuel type

Figure 4.7 Variation of hydro carbon with different fuels as input

4.4 EFFECT OF LOAD ON NOx EMISSION IN EXHAUST GAS

The NOx emission is measured in all modes of operations in this study. Figure 4.8 shows the NOx emission from the conventional mode with diesel, bio-diesel and HCCI with the same fuel for both the primary and secondary injections. In the conventional mode lower NOx emission is observed for bio-diesel based experiment, for all the load conditions as compared to diesel based operations. The HCCI mode operations produced lower NOx emissions as compared to the conventional mode of operations; this could be due to lower and uniform temperature inside the combustion

cxii chamber during the combustion process and due to uniform mixing of the fuel with air inside the combustion chamber.

600

500

400

300

200

100

0

Diesel

Biodiesel

Diesel - Diesel

Biodiesel - Biodiesel

0 20 40 60

Load (%)

80 100

Figure 4.8 Comparison of NOx emission between the conventional and

HCCI methods

The engine is also operated in the HCCI mode with different fuels as primary and secondary fuels. The NOx emissions show that the bio-diesel - petrol and diesel - petrol combinations generate very low NOx emission as compared to other cases of the HCCI mode of operations. The NOx emissions are observed to be of more or less the same value for these two cases for all the load conditions. The NOx emission has increased from 15 to 330 ppm and

19 to 342 ppm for diesel - petrol and bio-diesel - petrol modes respectively, when the load is increased from 0 to 100% as shown in Figure 4.9.

cxiii

600

500

400

300

200

100

0

Diesel - Biodiesel

Biodiesel - Diesel

Diesel - Petrol

Biodiesel - Petrol

0 20 40 60

Load (%)

80 100

Figure 4.9 Comparison of NOx emission between different fuels in

HCCI mode

Figure 4.10 shows the variation of the NOx emission with different fuels as input; the emission variation of the fuel sample is observed to be lower when compared to the HC emissions. The smaller variation is only observed with diesel, bio-diesel in the conventional mode, and diesel - diesel, bio-diesel - diesel in the HCCI mode of operations. Lower NOx emissions are observed for bio-diesel - bio-diesel, diesel - bio-diesel, diesel - petrol and biodiesel - petrol combinations in the HCCI mode of operations. For all these sample fuel operations, the NOx emission is observed to be lower than

250 ppm but for other cases of operations, the NOx emission is observed as higher than 250 ppm.

450

400

350

300

250

200

150

100

50

0

405

387

331

179

226

383

168 171 cxiv

Load (%)

Figure 4.10 Variation of NOx Emission with different fuels as input

Figure 4.11 shows the variation of NOx emission with bio-diesel as primary fuel and diesel, petrol, and bio-diesel as secondary fuels. The NOx emission variation with petrol as a secondary fuel is observed to be lower as compared to the other combinations.

450

400

350

300

250

200

150

100

50

0

Biodiesel - Biodiesel

Biodiesel - Diesel

Biodiesel - Petrol

0 20 40 60 80 100

Load (%)

Figure 4.11 Comparison of NOx emission between different loads with different HCCI modes

cxv

4.5 EFFECT OF LOAD ON SPECIFIC FUEL CONSUMPTION

The Specific Fuel Consumption (SFC) decreases with an increase in the load for all modes of operations (both conventional and HCCI mode).

The SFC is found to be of lower value in the conventional mode as compared to that in the HCCI mode for all the load conditions. This could be due to the secondary fuel supply in the case of the HCCI mode operation as compared to the conventional mode, in which only primary fuel is supplied. Figure 4.12 shows the specific fuel consumption with the increase in the load for the conventional and HCCI modes with the same type of fuel. The study also found that for all the higher load conditions, the specific fuel consumption is observed to be the same for all the cases.

0.8

0.7

0.6

0.5

0.4

0.3

0.2

0.1

0

Diesel

Biodiesel

Diesel - Diesel

Bio Diesel:Biodiesel

20 40 60 80 100

Load (%)

Figure 4.12 Comparison of specific fuel consumption between the conventional and HCCI methods

Figure 4.13 shows that a higher specific fuel consumption is observed for conventional mode compared to the HCCI mode of operation. In the case of the HCCI mode, with different fuels as primary and secondary fuel, the specific fuel consumption is reduced. This could be due to the

cxvi uniform and complete combustion of fuel in the case of the HCCI mode as compared to the conventional mode of operations.

0.45

0.4

0.35

0.3

0.25

0.2

0.15

0.1

0.05

0

0.3452 0.3538

0.4072

0.3804

0.195

0.1967

0.1556

0.1886

Fuel

Figure 4.13 Variation of specific fuel consumption with different fuels as

input

Figure 4.14 shows the comparison of specific fuel consumption with bio-diesel as a primary fuel and bio-diesel, diesel and petrol as secondary fuel. This shows that bio-diesel as a primary and secondary fuel requires more amount of fuel consumption as compared to other modes for the same load conditions. The specific fuel consumption is more or less the same in the case of bio-diesel as a primary fuel and diesel and petrol as secondary fuel. The higher specific fuel consumption in the case of bio-diesel could be due to its lower calorific value and other physical and chemical properties.

cxvii

0.8

0.7

0.6

0.5

0.4

0.3

0.2

0.1

0

Biodiesel - Biodiesel

Biodiesel - Diesel

Biodiesel - Petrol

0 20 40 60 80 100

Load (%)

Figure 4.14 Comparison of specific fuel consumption with different loads in different HCCI modes

4.6 EFFECT OF LOAD ON BRAKE THERMAL EFFICIENCY

The variation of the brake thermal efficiency in the conventional diesel engine is identified by diesel and bio-diesel as a fuel and diesel - diesel, bio-diesel - bio-diesel as a fuel in the HCCI mode with same fuel for primary and secondary injection. The brake thermal efficiency is measured with the increase of load from 20% to 100% of the load condition with the increase the every 20% load. Figure 4.15 indicates the variation of brake thermal efficiency with the increase of load between conventional and HCCI methods.

The study found that for all the modes of operations, the brake thermal efficiency increases with the increase in load. In the case of diesel and bio-diesel in the conventional mode of operation, the brake thermal efficiency has increased from 10.9% to 31.26% and 11.75% to 33.39% respectively when the load is increased from 20% to 100%. For the same fuels, the HCCI mode of operation produced higher thermal efficiency as compared to the conventional mode; this could be due to the lower combustion temperature

cxviii and complete conversion of energy available in the fuel into useful form in the case of the HCCI mode of operation as compared to the conventional mode of operation. The brake thermal efficiency with different fuels as primary and secondary fuel operations in the HCCI mode is represented in

Figure 4.16.

35

30

25

20

15

10

5

0

Diesel

Bio Diesel

Diesel - Diesel

Bio Diesel - Biodiesel

20 40 60 80 100

Load (%)

Figure 4.15 Comparison of brake thermal efficiency between conventional and HCCI methods

40

35

30

25

20

15

10

5

0

Diesel - Biodiesel

Biodiesel - Diesel

Diesel - Petrol

Biodiesel - Petrol

20 40 60 80 100

Load (%)

Figure 4.16 Comparison of brake thermal efficiency between different fuels in the HCCI mode

cxix

In the case of HCCI with different fuels as primary and secondary fuel supply, the diesel and petrol that are used with primary and secondary fuel have produced higher brake thermal efficiency as compared to other mode of operations. Diesel as a primary fuel and bio-diesel as a secondary fuel produced lower brake thermal efficiency. The magnitudes of brake thermal efficiency under all the load conditions for different modes of operations are presented in Figure 4.17. The study identified that for all the load conditions the diesel: diesel and diesel - petrol combinations under the

HCCI mode of operations has produced maximum thermal efficiency. Lower thermal efficiency is observed for diesel conventional and diesel - bio-diesel

HCCI modes.

40

35

30

25

20

15

10

5

0

Diesel

Bio Diesel

Diesel -Diesel

Bio diesel- Biodiesel

Diesel - Biodiesel

Biodiesel - Diesel

Diesel - Petrol

Biodiesel - Petrol

20 40 60 80 100

Load (%)

Figure 4.17 Brake thermal efficiency under different load conditions for all the modes of operations

The brake thermal efficiency for different modes of operations is represented in Figure 4.18. The study found that the brake thermal efficiency at a particular load condition for all the fuels with different primary and secondary fuels, in the conventional and HCCI modes is more or less the same; only very small variations are observed for certain cases. A higher

cxx brake thermal efficiency is observed for the diesel - diesel and diesel - petrol combinations in the case of the HCCI mode of operations. This could be due to the higher energy content and more suitability of petrol and diesel for combustion process as compared to the other combination fuels. Lower efficiency is observed for diesel in the conventional mode and diesel - biodiesel in the HCCI mode.

15

10

5

0

30

25

20

22.84

24.49

22.56

25.18

20.32

20.32

25.97

25.71

Fuel

Figure 4.18 Variation of Brake Thermal Efficiency with different fuels

as input

Figure 4.19 shows the brake thermal efficiency with, bio-diesel as a primary fuel and bio-diesel, petrol, diesel as secondary fuels. The brake thermal efficiency under all the load conditions for these types of fuel exhibits more or less the same trends. The thermal efficiency for the bio-diesel - diesel combination is lower than that the other two cases. But the percentage variation is only observed as a very small value.

cxxi

10

5

0

25

20

15

40

35

30

Bio diesel - Biodiesel

Biodiesel - Diesel

Biodiesel - Petrol

0 20 40 60 80 100

Load (%)

Figure 4.19 Comparison of Brake Thermal Efficiency under different loads indifferent HCCI modes

4.7 EFFECT OF LOAD ON CARBON DIOXIDE

The effect of load on the CO

2

emission in the conventional and

HCCI modes with the same type of fuel is shown in Figure 4.20.

7

6

5

4

3

2

1

0

Diesel

Bio diesel

Diesel - Diesel

Bio diesel - Biodiesel

0 20 40 60 80 100

Load (%)

Figure 4.20 Comparison CO

2

between the conventional and HCCI methods

cxxii

The study found that CO

2

emission from diesel in the conventional and HCCI modes is higher than that of the bio-diesel in the conventional and

HCCI modes. This could be due to the variation in the chemical characteristics of diesel and bio-diesel. The CO

2

emission has increased from

3% to 6.3% and 2.5% to 4.9% for diesel and bio-diesel respectively, when the load is increased from 0% to 100%. For the same load conditions, the CO

2 emission is observed to be 2.7% to 5.7% and 1.8% to 4.8% respectively for diesel and bio-diesel in the HCCI mode respectively. The CO

2

emission from diesel in the HCCI mode is higher than that of bio-diesel in the conventional mode as shown in Figure 4.21.

4

3

2

6

5

Diesel - Biodiesel

Biodiesel - Diesel

Diesel - Petrol

Biodiesel - Petrol

1

0

0 20 40 60 80 100

Load (%)

Figure 4.21 Comparison of CO

2

emission between different fuels in the

HCCI mode

Figure 4.22 shows that a mixed trend of CO

2

emission in the conventional, and HCCI modes, under different fuel conditions is observed.

Bio-diesel in the conventional mode and the same with different fuel combinations in the HCCI mode, produced higher CO

2

emissions as compared to other fuels in the conventional and HCCI modes. Diesel, petrol in the conventional mode and the same with other fuels in the HCCI mode

cxxiii produced lower amount of CO

2

emissions. This could be due to bio-diesel combustion that generates more amount of CO

2

as compared to that of petrol and diesel as shown in Figure 4.23.

3

2

1

0

6

5

4

5.6

4.2

5

3.45

3.2

4.2

1.7

2.5

Load (%)

Figure 4.22 Variation of CO

2

emission with different fuels as input

6

5

4

3

2

1

0

Biodiesel - Biodiesel

Biodiesel - Diesel

Biodiesel - Petrol

0 20 40 60 80 100

Load (%)

Figure 4.23 Comparison of Brake Thermal Efficiency under different loads in different HCCI mode

4.8

cxxiv

EFFECT OF LOAD ON CARBON MONOXIDE EMISSION

The variation of CO emission with the variation in the load for all modes of operations in this study, is represented in Figures 4.24 and 4.25. The study found that there is no variation of CO with the increase in load for all the HCCI modes of operations with different fuels as primary and secondary fuels except the diesel - petrol combination. In the case of the conventional mode of operation, the CO emission initially increased with the increase in load; further increase in the load, decreased the CO emission. Among all the

HCCI modes, the bio-diesel - bio-diesel mode produced very low amount of

CO emission as compared to the other cases of HCCI modes.

Figure 4.26 shows the CO emission in four different HCCI modes of operations with three different fuel combinations under all the load conditions. Bio-diesel as a primary fuel and secondary fuel produced lower amount of CO emissions as compared to petrol and diesel. This could be due to more amount of CO emissions in the case of bio-diesel as a secondary fuel and lower amount of CO emissions in the case of petrol and diesel as a primary fuel.

0.6

0.5

0.4

0.3

0.2

0.1

0

Diesel

Bio Diesel

Diesel - Diesel

Bio Diesel - Biodiesel

0 20 40 60 80 100

Load (%)

Figure 4.24 Comparison of CO emission between the conventional and

HCCI methods

cxxv

0.35

0.3

0.25

0.2

0.15

0.1

0.05

0

0%

20% Load

40% Load

60% Load

80% load

100% Load

Fuel Combinations

Figure 4.25 Comparison of CO emission between different fuels in the

HCCI mode

0.35

0.3

0.25

0.2

0.15

0.1

0.05

0

0.3

0.25

0.2

0.1

0.1

0.1

0.2

0.1

Load (%)

Figure 4.26 Variation of CO emission with different fuels as input

cxxvi

The study found that bio-diesel as fuel could produce more amount of CO. Bio-diesel – petrol produces lesser amount of CO. The reason could be higher O

2

content in the case of bio-diesel and lower O

2

content in the case of bio-diesel - petrol. The conclusion regarding the design of HCCI parameters such as exhaust gas temperature, SFC, brake thermal efficiency, and of emissions of CO, HC, NOx, and CO

2

is discussed in chapter 5.

cxxvii

CHAPTER 5

CONCLUSION

5.1 FINDINGS FROM THE RESEARCH

The research has been comprehensive keeping in mind the current and future needs of HCCI. The result of this exhaustive research has brought out the application of bio-diesel to the emerging needs of HCCI combustion.

The salient details of the research findings are as follows:

1. Eight different modes such as the diesel conventional, Biodiesel conventional, Diesel - Diesel PCCI-DI combustion,

Bio-diesel - Bio-diesel PCCI-DI combustion, Diesel - Biodiesel PCCI-DI Combustion, Bio-diesel - Diesel PCCI-DI combustion, Diesel - Petrol PCCI-DI combustion, and Biodiesel - Petrol PCCI-DI combustion are incorporaded in the proposed HCCI system.

2. Various parameters such as exhaust gas temperature, HC,

NOx, CO, SFC, BTE, and CO

2 are measured and analysed.

cxxviii

3. A MRU DELTA 1600L five gas analyzer is used for conducting the experiment.

4. The PCCI approach is adopted to conduct the HCCI experiment.

5.2

5. The Pilot Injection (PI) system is used.

CONTRIBUTIONS

The imposition of stringent emission norms and depletion of fossil fuel resources led engineers to work out new combustion technologies to substantially reduce harmful emission and to improve the overall efficiency of an IC Engine. The factors to be considered while designing a new combustion process are a higher compression ratio, lean homogeneous air fuel mixture, complete combustion, and instantaneous combustion, which lead to a

Homogeneous Charge Compression Ignition (HCCI). HCCI is a clean and efficient combustion process. In this research work an attempt is made to experimentally analyse the performance and emission characteristics of the

HCCI compression process in the Premixed Charge Compression Ignition

(PCCI) mode, assisted by Pilot Injection (PI) as a combustion initiator.

Several experiments were conducted in a modified single cylinder watercooled diesel engine employing a conceptual system known as Transient State

Fuel Induction (TSFI) with different fuels such as diesel, petrol, and biodiesel. In the present research setup, it is observed that there is a reduction in the emission level of CO and HC with the same power out of the conventional diesel engine. This research study has given a holistic approach to investigate and reduce emission in an HCCI engine in the PCCI mode with different fuels. The major contributions, probably, the first of their kind, are as follows:

1. 5 bio-diesel combinations are tried for emission reduction.

cxxix

2. HCCI with PCCI mode is experimented successfully.

3. gas analyser is used to observe the emission of NOx, CO,

CO

2

, HC and O

2

.

4. The pilot injection of secondary fuel is achieved.

5. The rocker arm is positioned at 1.97 mm from the bottom most position.

6. An electrical signal is used for closing the secondary injection.

7. The conversion from heterogonous to homogenous mixture of the diesel engine is achieved.

8. A higher compression ratio is obtained.

9. A lean homogeneous air fuel mixture is used.

10. Complete combustion is achieved.

11. Instantaneous combustion is achieved and

12. Clean and efficient combustion process is suggested.

5.3 LIMITATIONS

Research and development in this area is cumbersome and takes a lot of time and large-scale effort. The limitations of the present study are that only 8 combinations modes are tried. Another limitation of this study is that friction losses and charge leakage through the cylinder-piston gap are not considered; it may constrain the engine size, and impose minimum engine speed limits.

5.4 SCOPE FOR FUTURE RESEARCH

cxxx

The present investigation has greatly expanded our understanding of HCCI, the controlling mechanisms, and HCCI engine operation strategies.

However, the present research can be extended to various other alternative fuels such as CNG, LPG, biogas, and chemfuels, and the reduction of emission can be studied and analysed. The present system can be slightly modified to design low NOx, low particulate reciprocating engines with improved diesel efficiency. Research and development is needed to develop methods to slow the rate of combustion in HCCI engines at high engine loads, to prevent excessive noise and engine damage. Similar research can also be done to develop concepts to overcome the challenge of ignition in cold HCCI engines without compromising the warm engine performance. The present system can be adopted to develop intake and exhaust manifold designs for multi-cylinder engines to overcome the challenge of maintaining strict uniformity of the inlet and exhaust flows of each cylinder to assure smooth engine operation.

5.5 CONCLUSION

This research work aims to reduce emission in HCCI engines in the

PCCI mode with different fuels. The following conclusions are drawn:

The exhaust gas temperature in the HCCI mode is lower than that of the conventional mode of operation. Among the HCCI mode with different fuel combinations, bio-diesel and petrol are used as primary and secondary fuels respectively, and produced lower temperatures than that of the HCCI mode with other combination of fuels.

cxxxi

The study found that the HC emission decreases with the increase in load for the HCCI mode with all types of fuels. But in the case of the conventional mode the HC emission has increased with the increase in load. Lower HC emission is observed for diesel and bio-diesel as primary and secondary fuels respectively.

Lower NOx emission is observed for all the HCCI modes of operation as compared to that of the conventional mode with different fuel combinations. Among all HCCI modes of operations, diesel and petrol as primary and secondary fuel produced lower NOx emission.

The specific fuel consumption has decreased with the increase in load for all modes of operations. The lower specific fuel consumption is observed for diesel and petrol as primary and secondary fuel respectively. The specific fuel consumption in the HCCI mode is observed to be lower than in the case of the conventional mode.

The brake thermal efficiency of diesel and petrol as primary and secondary fuel respectively in the HCCI mode is higher than that of the other modes of operation with different fuels.

For the same load condition, the same brake thermal efficiency is observed for the conventional and HCCI modes.

The study observed that the CO

2

emission from the HCCI mode is lower than that of the conventional mode; this could be due to the complete combustion of fuel inside the combustion chamber. Among the all HCCI modes, diesel and petrol as primary and secondary fuels respectively, produced lower CO

2

emissions.

cxxxii

There is no change in the CO emission with the increase in load in the case of the HCCI mode with different fuel combinations. The use of bio-diesel reduced the CO emission as compared to the other cases.

The study concludes that the emission from the HCCI mode is lower than that from the conventional mode but the specific fuel consumption varied with different combinations of fuel in the HCCI mode. Diesel and petrol as primary and secondary fuel is better than the other fuel combinations from the environmental, global warming and fuel consumption point of view.

Once properly introduced in the process, the proposed HCCI combustion system should improve performance and reduce the emission and consequently contribute to the efficiency profitability of the automobile industry.

This thesis also suggests that before making any commitment to erect and commission the recommended HCCI combustion engine, the user should carefully study the description, and if possible consult other engine designers who are recommended.

cxxxiii

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LIST OF PUBLICATIONS

INTERNATIONAL JOURNALS

1.

Mohanamurugan S. and Sendilvelan S. (2009) ‘Experimenting of

homogenous charge compression ignition (HCCI) engine with premixed charge compression ignition (PCCI) mode operating at different fuel conditions using petrol’, International Journal of

Engineering Research and Industrial Applications, Vol. 2, No. VII, pp. 1-12, ISSN 0974-1518

2.

Mohanamurugan S. and Sendilvelan S. (2009) ‘Analyzing of

technical feasibility in homogenous charge compression ignition

(HCCI) engine operating at different fuel conditions using a secondary injector’, International Journal of Applied Engineering Research,

Vol. 4, No. 9, pp. 1679-1688, Print ISSN : 0973-4562.

3.

Mohanamurugan S. and Sendilvelan S. (2009) ‘Homogenous charge

compression ignition (HCCI) engine operating at different fuel conditions’, International Journal of Engineering Research and

Industrial Application (Accepted for Publications).

4.

Mohanamurugan S. and Sendilvelan S. (2009) ‘Optimization of

HCCI engine in PCCI mode operating at different fuel conditions’,

International Journal of Applied Engineering Research (Accepted for publications).

5.

Mohanamurugan S. and Sendilvelan S. (2010), ‘Emission and

combustion characteristics of different fuel in HCCI engine’,

International Journal of Automotive and Mechanical Engineering,

Malaysia (Accepted for publications).

6.

Mohanamurugan S. and Sendilvelan S. (2009) ‘Improved emission

and efficiency characteristics of HCCI engine with different premixed fuels using a secondary pilot injector in the inlet manifold’,

International Journal of Automotive Technology (Under Review).

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7.

Mohanamurugan S. and Sendilvelan S. (2009) ‘Effects on HCCI

engine using bio-diesel at direct injection method with other fuels in two stage ignition process’, International Journal of Thermal Science

(Under Review).

8.

Mohanamurugan S. and Sendilvelan S. (2009) ‘Experimenting

technical feasibility in homogenous charge compression ignition engine operating with diesel and bio-diesel using an secondary injector’, International Journal of Gas Turbine and Power (Under

Review).

9.

Mohanamurugan S. and Sendilvelan S. (2009) ‘Experimental

analysis of HCCI engine using bio-diesel at direct injection method with PCCI mode’, The Journal of Engineering Research (Under

Review).

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VITAE

S. MOHANAMURUGAN was born in Tamil Nadu, India on 18

th

June 1972. He received the degree of Bachelor of Engineering in Mechanical

Engineering in 1998 from the Madras University, Chennai, India. He was awarded the degree of Master of Engineering in Energy Engineering and management in 2004 by the Annamalai University, Chidambaram, India.

A few of the honours obtained are (1) University first rank in B.E degree programme. (2) University first rank in M.E degree programme.

(3) Was adjudged the best programme officer of the NSS unit at the Periyar

University.

He put in 16 years of experience in teaching and Industries. He joined as part-time Research Scholar for his Ph.D. at Dr.M.G.R University.

Presently he is teaching in the Mechanical Engineering department at

Velammal Engineering College, Chennai, India. He has guided a number of

B.E and M.E projects in the area of Energy systems. His area of interest includes I.C Engine Emission, Renewable Energy Conversion and Energy

Conservation methods.

He has published 6 papers in internal journals and 15 papers in national and internal conferences. Contributions to professional bodies include life membership in both ISTE and Indian Combustion Institute in

Chennai. He has organised several workshops, seminars, FDP, STTP and conferences.

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