compressed natural gas as an alternative fuel in diesel engines

compressed natural gas as an alternative fuel in diesel engines
University of Southern Queensland
Faculty of Engineering and Surveying
COMPRESSED NATURAL GAS AS AN
ALTERNATIVE FUEL IN DIESEL ENGINES
A dissertation submitted by
WONG, Wei Loon
in fulfillment of the requirements of
Courses ENG 4111 and ENG 4112 Research Project
towards the degree of
Bachelor of Engineering (Mechanical)
Submitted: October, 2005
ABSTRACT
This project aims to investigate the engine performance and exhaust emission of
dual-fuel operation on a single cylinder compression ignition engine. Natural gas is
used as the main gaseous fuel with diesel as the pilot fuel to provide a source of
ignition. The high compression ratios of diesel engines can be achieved without loss
in power together with substantial cost reduction in fuel and conversion kits.
Comparative results of dual-fuel operation with conventional diesel fuel through
experimental results demonstrated its benefits both in the fields of performance and
emission. The engine torque and brake power output show vast improvement and
dual-fuel operation is able to achieve a higher thermal efficiency under all operating
conditions. The emission levels of polluting gases such as carbon monoxide, oxides
of nitrogen (NOx) and carbon dioxide also records an enormous decrease.
i
University of Southern Queensland
Faculty of Engineering and Surveying
ENG4111 & ENG4112 Research Project
Limitations of Use
The Council of the University of Southern Queensland, its Faculty of Engineering
and Surveying, and the staff of the University of Southern Queensland, do not accept
any responsibility for the truth, accuracy or completeness of material contained
within or associated with this dissertation.
Persons using all or any part of this material do so at their own risk, and not at the
risk of the Council of the University of Southern Queensland, its Faculty of
Engineering and Surveying or the staff of the University of Southern Queensland.
This dissertation reports an educational exercise and has no purpose or validity
beyond this exercise. The sole purpose of the course pair entitled "Research Project"
is to contribute to the overall education within the student’s chosen degree program.
This document, the associated hardware, software, drawings, and other material set
out in the associated appendices should not be used for any other purpose: if they
are so used, it is entirely at the risk of the user.
Prof G Baker
Dean
Faculty of Engineering and Surveying
ii
Certification
I certify that the ideas, designs and experimental work, results, analyses and
conclusions set out in this dissertation are extremely my own effort, except where
otherwise indicated and acknowledged.
I further certify that the work is original and has not been preciously submitted for
assessment in any other course or institution, except where specifically stated.
Wong Wei Loon
Student Number : 0050027398
______________________________________
Signature
______________________________________
Date
iii
ACKNOWLEDGEMENTS
I would like to express my gratitude to my supervisors -Dr. Harry Ku who guided
me through this project as well as providing a lot of handy tips, Dr. Fok Sai-Cheong
who provided assistance and help to me regarding the research work during his time
here and Dr. Talal Yusaf from UNITEN who is my local supervisor back in
Malaysia, who gave me permission to access the labs to conduct the experiments for
this research.
iv
TABLE OF CONTENTS
ABSTRACT
i
ACKNOWLEDGEMENTS
iv
LIST OF FIGURES
ix
LIST OF TABLES
xi
NOMENCLATURE
xii
GLOSSARY
xiii
CHAPTER 1 - INTRODUCTION
1
1.1 Project Background
1
1.2 Project Objectives
2
1.3 Project Methodology
3
CHAPTER 2 – LITERATURE REVIEW
2.1 Compressed Natural Gas (CNG)
5
5
2.1.1
Introduction
5
2.1.2
Usage of Compressed Natural Gas
6
2.1.3
Natural Gas Composition
7
2.1.4
Natural Gas Properties
9
2.2 Advantages and Limitations of CNG
11
2.2.1
Advantages
11
2.2.2
Limitations
12
2.3 Safety
13
2.4 Diesel Engine Conversion
13
2.4.1
Introduction to conventional diesel engine
13
2.4.2
Engine conversion types
17
v
2.4.2.1
Bi-fuel engine
17
2.4.2.2
Dedicated NGV engine
17
2.4.2.3
Dual-fuel engine
18
CHAPTER 3 – ENGINE PERFORMANCE
20
3.1 Torque
20
3.2 Input Power
21
3.3 Brake Power
22
3.4 Specific Fuel Consumption
22
3.5 Brake Mean Effective Pressure
24
3.6 Engine Thermal Efficiency
25
CHAPTER 4 – EMISSION
27
4.1 Carbon Monoxide (CO)
27
4.2 Total Hydrocarbon (THC)
29
4.3 Nitrogen Oxides (NOx)
30
4.4 Particulate Matters (PM)
32
4.5 Carbon Dioxide (CO2)
33
4.6 Oxides of Sulphur (SOx)
33
CHAPTER 5 – EXPERIMENTAL SETUP
35
5.1 Engine Preparation
35
5.2 Preparation of Load Banks
35
5.3 Calibration of Digital Thermocouple
36
5.4 Gas Analyzer Setup
36
5.5 Natural Gas Conversion Kit
37
5.5.1 Natural Gas Pressure Regulator
38
5.5.2 Natural Gas Solenoid Valve
38
5.5.3 Natural Gas Mixer
39
vi
5.5.4 Fuel Selector Switch and Gauge
5.6 Experimental Procedures
CHAPTER 6 – RESULTS AND DISCUSSION
6.1 Introduction
39
41
43
43
6.2 Engine Performance for Maximum Load Operating Conditions 44
.
6.2.1 Engine Body Temperature
44
6.2.2 Engine Exhaust Temperature
45
6.2.3 Engine Torque
46
6.2.4 Engine Brake Power
47
6.2.5 Brake Specific Fuel Consumption
48
6.2.6 Engine Thermal Efficiency
50
6.3 Exhaust Emission for Maximum Load Operating Conditions
52
6.3.1 Carbon Monoxide
52
6.3.2 Unburnt Hydrocarbons
54
6.3.3 Oxides of Nitrogen (NOx)
56
6.3.4 Carbon Dioxide
57
6.3.5 Excess Oxygen
59
6.4 Engine Performance for Moderate Load Operating Conditions
61
6.4.1 Engine Body Temperature
61
6.4.2 Engine Exhaust Temperature
62
6.4.3 Engine Torque
63
6.4.4 Engine Brake Power
64
6.4.5 Brake Specific Fuel Consumption
65
6.4.6 Engine Thermal Efficiency
66
6.5 Exhaust Emission for Moderate Load Operating Conditions
68
6.5.1 Carbon Monoxide
68
6.5.2 Unburnt Hydrocarbons
69
vii
6.5.3 Oxides of Nitrogen (NOx)
70
6.5.4 Carbon Dioxide
71
6.5.5 Excess Oxygen
72
6.6 Concluding Discussion
73
6.7 Catalytic Aftertreatment
75
CHAPTER 7 – CONCLUSION
78
7.1
Achievement of Objectives
78
7.2
Recommendation and Future Work
79
LIST OF REFERENCES
80
BIBLIOGRAPHY
84
APPENDIX A – PROJECT SPECIFICATION
86
APPENDIX B – ENGINE SPECIFICATIONS
89
APPENDIX C – GAS ANALYZER SPECIFICATIONS
91
APPENDIX D – EXPERIMENTAL DATA
93
APPENDIX E – SAMPLE ANALYSIS CALCULATIONS 97
viii
LIST OF FIGURES
Figure 1.1
- Project Methodology
Figure 4.1
- Variation of CO emission for different fuels
Figure 4.2
- Ratio of NO2 to NO in diesel exhaust for varying engine load and
speed
Figure 5.1
- Autologic Autogas Gas Analyzer
Figure 5.2
- Natural Gas Converter Kit
Figure 5.3
- Experimental Setup Overview
Figure 6.1
- Engine body temperatures under maximum load operating
conditions
Figure 6.2
- Engine exhaust temperatures under maximum load operating
conditions
Figure 6.3
- Engine torque output under maximum load operating conditions
Figure 6.4
- Engine brake power under maximum load operating conditions
Figure 6.5
- Brake specific fuel consumption under maximum load operating
conditions
Figure 6.6
- Engine thermal efficiency under maximum load operating
conditions
Figure 6.7
- Emission of CO under maximum load operating conditions
Figure 6.8
- Emission of HC under maximum load operating conditions
Figure 6.9
- Emission of NOx under maximum load operating conditions
Figure 6.10
- Emission of CO2 under maximum load operating conditions
Figure 6.11
- Excess of O2 under maximum load operating conditions
Figure 6.12
- Engine body temperatures under moderate load operating
conditions
ix
Figure 6.13
- Engine exhaust temperatures under moderate load operating
conditions
Figure 6.14
- Engine torque output under moderate load operating conditions
Figure 6.15
- Engine brake power under moderate load operating conditions
Figure 6.16
- Brake specific fuel consumption under moderate load operating
conditions
Figure 6.17
conditions
- Engine thermal efficiency under moderate load operating
Figure 6.18
- Emission of CO under moderate load operating conditions
Figure 6.19
- Emission of HC under moderate load operating conditions
Figure 6.20
- Emission of NOx under moderate load operating conditions
Figure 6.21
- Emission of CO2 under moderate load operating conditions
Figure 6.22
- Excess of O2 under moderate load operating conditions
Figure 6.23
- Conversion efficiency of three-way catalyst as a function of air-fuel
ratio
x
LIST OF TABLES
Table 2.1
- Typical Composition of Natural Gas in Percentage
Table 2.2
- Properties of Natural Gas and Diesel
Table 6.1
- Engine Performance for Maximum Load
Table 6.2
- Exhaust Emission for Maximum Load
Table 6.3
- Engine Performance for Moderate Load
Table 6.4
- Exhaust Emission for Maximum Load
xi
NOMENCLATURE
λ
-
Air to fuel ratio
p
-
Cylinder pressure
V
-
Cylinder volume
ηf
-
Engine thermal efficiency
τ
-
Engine torque
Qch
-
Gross heat release rate
Q
-
Heat transfer
Qht
-
Heat transfer rate to the cylinder walls
Q HV
-
Lower calorific value of fuel
mf
-
Mass flow rate of fuel
Qn
-
Net heat release rate
nR
-
Number of crank revolutions for each power stroke per
cylinder
n
-
Number of engine cylinders
γ
-
Ratio of specific heats
hf
-
Sensible enthalpy of fuel
U
-
Sensible internal energy of cylinder contents
cp
-
Specific heat at constant pressure
cv
-
Specific heat at constant volume
Wc
-
Work produced per cycle
xii
GLOSSARY
A
-
Area of engine bore
bmep
-
Brake mean effective pressure
BP
-
Brake power
bsfc
-
Brake specific fuel consumption
C 2 H6
-
Ethane
C 3 H8
-
Propane
CH4
-
Methane
CI
-
Compression Ignition
CNG
-
Compressed Natural Gas
CO
-
Carbon monoxide
CO2
-
Carbon dioxide
CR
-
Compression ratio
DC
-
Direct Current
DPM
-
Diesel Particulate Matter
ECU
-
Electronic Control Unit
H
-
Hydrogen
IP
-
Input power
L
-
Length of engine stroke
LNG
-
Liquefied Natural Gas
N
-
Engine speed
NGV
-
Natural Gas Vehicles
NO
-
Nitric oxide
NO2
-
Nitrogen dioxide
NOx
-
Nitrogen oxides
O2
-
Oxygen
OH
-
Hydroxide
P
-
Power developed by engine
PM
-
Particulate Matter
rpm
-
Revolutions per minute
xiii
SCR
-
Selective catalytic reduction
sfc
-
Specific fuel consumption
SO4
-
Sulfate
SOF
-
Soluble Organic Fractions
THC
-
Total unburnt hydrocarbon
V
-
Voltage
xiv
CHAPTER 1
INTRODUCTION
1.1
PROJECT BACKGROUND
Diesel fuel in compression ignition (CI) engines produces a high level of toxicity in
emission gases (Kojima, 2001) which leads to a health and environmental hazard.
The high level of nitrous oxides (NOx), carbon monoxide (CO), carbon dioxide
(CO2) and particulate matter 10 (PM10) emission associated with diesel fuel has
long been an issue. Although the use of diesel is favorable in fleet vehicles since it
produces a high compression ratio to enable generation of more power, Kelley
(undated) reported that the higher compression ratio causes a significant problem in
starting the engine at low temperatures. Wills (2004) supported the findings and
mentioned that fuel type plays an important role in the ease to start the engine.
Natural gas has been considered as a potential substitute to conventional fuels in
vehicles due to its lower emission of greenhouse gases and safety properties. A
frequent report by natural gas vehicle (NGV) owners is that CNG powered vehicles
have less power and shorter driving range (Graham, 2000). In fact, this is due to the
lower compression ratio when spark ignition dedicated CNG engines are used. The
emission and reduced performance problems of both diesel and dedicated CNG
engines can be eliminated by the use of dual fuel diesel-CNG engines.
1
1.2
PROJECT OBJECTIVES
The main object of this project is to research the effects of using CNG as an
alternative fuel as a replacement for diesel in compression ignition engines. This
means that dual-fuel engine must be used to utilize diesel as the pilot fuel to ignite
CNG. The engine performance and emission qualities are to be investigated by
running the engine at different speeds with varying set of loads. The sub-objectives
of the project are:
a) Research the history of CNG usage worldwide and a literature review on the
engine performance using CNG as the main fuel supply inclusive of the
advantages and limitations.
b) Conversion of the current CI engine to install the CNG fuel system to enable
the use of dual fuel diesel-CNG engine.
c) Study on the effect of using CNG as fuel in terms of power, torque, brake
specific fuel consumption (BSFC), and thermal efficiency. Perform a
comparison analysis on the dual fuel combustion and conventional diesel
fuel.
d) Examine the emission data collected for both fuels and conduct feasibility
study on CNG as a fuel alternative in terms of pollution and economy.
2
1.3
PROJECT METHODOLOGY
Initially, literature review of CNG usage worldwide is compiled and commented.
Information on usage of CNG fuel in vehicles worldwide and locally was gathered
from online sources, journals, magazines and newspaper articles. A concise
summary of properties of natural gas is presented and critically documented. At the
same time, reviews are made on the advantages and disadvantages of CNG as fuel
compared to conventional fuels such as diesel and gasoline.
Next, the engine performance of the dual fuel CNG-diesel engine will be analyzed
with properties such as engine torque, brake power, brake specific fuel consumption,
brake mean effective pressure and engine efficiency being emphasized. The gaseous
emission types are also explained by examining the formation and causes of
pollutants such as nitrous oxides, carbon monoxide, carbon dioxide, and total
unburnt hydrocarbons.
The experimental procedures and setup will then be explained to illustrate the
method of measuring engine performance and emission in the laboratory.
Experimental data will be recorded systematically for different engine speeds and
varying loads to enable comparisons of pure diesel and dual-fuel to be made. The
data collected will be tabulated and relevant graphs plotted. Next, the results will be
critically analyzed and finally a conclusion is made based on the experimental
results.
3
Figure 1.1: Project Methodology
4
CHAPTER 2
LITERATURE REVIEW
2.1
COMPRESSED NATURAL GAS (CNG)
2.1.1
Introduction
Compressed Natural Gas is composed primarily of methane (CH4), and other
hydrocarbons such as ethane, propane and butane. According to Alternative Fuel
Data Centre (2004), the composition of CNG is further enriched with other gases
such as carbon dioxide, hydrogen sulphide, nitrogen, helium and water vapour. The
content of CNG was believed to originate from plants and animal remains which had
decomposed for millions of years (Info Comm, 2005). Natural gas is formed deep
underground trapped between layers of rock and sand in reservoirs underneath the
Earth, like other fossil fuels. Due to its lower density characteristics, CNG will float
above other trapped substances such as crude oil and water.
A drilling rig is used to penetrate the Earth surface to draw out the natural gas. The
extract is then refined to remove impurities and transmitted through a series of
pipelines to processing plants and then sent to transmission companies before
reaching the end-user (NaturalGas.org, 2004).
Natural gas was first used as fuel to boil water, light street lamps and gained
worldwide acceptance to be used in residential homes as water heaters, clothes dryer
and in cooking. The Pacific Gas and Electric Company (2003) reported that the
5
natural gas boom period began in 1950s in the United States where a huge network
of facilities and distribution pipes were constructed for the purpose of promoting the
use of natural gas. The use of natural gas in the transportation sector began as early
as the 1930s and had little development since then (NaturalGas.org, 2004). Without
the support of the public, the use of Natural Gas Vehicles (NGV) is limited mostly
to the public transportation sector.
Natural gas is compressed as CNG to be used as fuel in the vehicles with the
alternative being Liquefied Natural Gas (LNG). The former is most widely used in
alternative fuel vehicles. It promotes environmental friendliness with its low
emission of harmful gases and comparable engine performance (U.S. Environmental
Protection Agency, 2002).
2.1.2
Usage of Compressed Natural Gas
NaturalGas.org (2004) reported that natural gas was originally used by the Chinese
as a fuel to process seawater to separate its salt contents and make it drinkable. In
Europe, Britain was the first country to commercialize the use of natural gas
although its use was limited only to street lighting (Gas-Lite Manufacturing, 2004).
The method then spread to other parts of the world including the United States. After
the Second World War, improvements were made to the transportation and storage
of natural gas with extensive use of the technology available then (NaturalGas.org,
2004). As predicted, with the increased area coverage of natural gas supplies, it
became an increasingly popular source of energy for the public.
Intense research was performed to analyze the feasibility of using natural gas as a
substitute for conventional fuel like gasoline and diesel. According to Shamsudin &
Yusaf (1995), Italy became the leading country in research of natural gas as an
alternative fuel and had around 235 000 vehicles converted to be powered by the
fuel. The United States is now the world’s highest consumer of natural gas with
28.8% consumption followed by the Russian Federation (BP Global, 2005). Almost
130,000 NGVs are operational in the United States with an approximation of 2.5
million vehicles worldwide (NaturalGas.org, 2004). The Pacific Gas and Electric
6
Company (2003) reported that there are more than 40 natural gas vehicles
manufacturing companies available worldwide. This shows a significant expansion
in the field of research of NGVs.
CNG are primarily used as fuel for transit buses, taxi cabs, heavy duty trucks and
other public vehicle fleet worldwide. The high fuel usage of these vehicles makes
conversion to CNG fuel more economical and decreases the payback period of the
conversion cost. Due to the environmental benefits of CNG as an alternative fuel,
the American Government is investing heavily in the field of Research &
Development and providing subsidies to support NGVs in order to encourage the
use of natural gas (Natural Gas Vehicle Coalition, 2005).
In Malaysia, the popularity of NGV is limited and the majority of vehicles are still
running on conventional fuel due to the inadequate fuelling stations and long
fuelling time for natural gas. Although Malaysia has an abundant natural gas reserve
estimated at around 82.5 trillion cubic feet (Autoworld, 2004), the market
penetration remains minimal and the majority of natural gas is exported. However,
steps are being implemented to promote the use of NGVs through publications and
education to instill customer awareness on its vast benefits. PETRONAS Company,
through its subsidiary, PETRONAS NGV Sdn. Bhd., has been promoting the use of
CNG to Malaysian motorists by providing services and fuelling facilities
(Autoworld, 2004). Apart from that, Malaysia has been selected as the host nation
for the Asia Pacific Natural Gas Vehicles Association (ANGVA) conference in 2005
to discuss the technological developments in the use of NGVs.
2.1.3
Natural Gas Composition
Natural gas generally consists of a mixture of hydrocarbons with methane (CH4) as
the main constituent. Ethane, propane, butane, nitrogen and carbon dioxide gases
contribute to the remaining composition while traces of water vapour and hydrogen
sulphide may be present in some natural gases. The properties of natural gas will
vary depending on the location, processing and refining facilities. Usually, the
7
maximum and minimum compositions are specified to enable comparisons to be
made.
Compound
Typical
Maximum
Minimum
Methane
87.3%
92.8%
79.0%
Ethane
7.1%
10.3%
3.8%
Propane
1.8%
3.3%
0.4%
Butane
0.7%
1.2%
0.1%
Nitrogen
2.2%
8.7%
0.5%
Carbon Dioxide
0.9%
2.5%
0.2%
Table 2.1: Typical Composition of Natural Gas in Percentage (Questar Gas, undated)
Current research on the natural gas vehicles found that the engine performance and
emission are greatly affected by varying compositions of natural gas (Ly, 2002). It
was also reported that the heating value, efficiency, and concentration of unburnt
hydrocarbon and other emission particles would highly depend on the source of
supply of natural gas as the main fuel. Ly (2002) also mentioned that this effect is
especially dominant in heavy-duty engines with high compression ratio applications
due to the increased amount of engine “knocking”. Engine knocks are caused by the
pre-mature ignition of the air-fuel mixture in the combustion cylinder, causing the
engine to overheat and run inefficiently.
According to Natural Gas.org (2004), the raw natural gas is processed to remove
impurities such as oil, condensate and water particles. The presence of these
particles may obstruct the smooth flow of fuel into the engine when in use and may
even bring the engine to a halt. ‘Dry’ natural gas, which consists of almost entirely
methane, is then obtained by distilling the other hydrocarbons.
8
2.1.4
Natural Gas Properties
Natural gas in its original form is non-toxic, colorless and odorless (Questar Gas,
undated). A chemical substance called Mercaptan is added to natural gas to add a
scent of rotten egg as a safety precaution so that leakage may be detected by the
human olfactory sense (Info Comm, 2005). Inhalation of natural gas will not
interfere with the body functions or cause detrimental health damage to our body.
Barbotti CNG (2002) mentioned that the natural gas does not emit any aldehydes
and other air toxins, which may be an issue with other fuel types.
Apart from that, natural gas is also lighter than air due to its low density. According
to Clean Air Technologies Information Pool (2005), a natural gas spill would be less
dangerous compared to a gasoline or diesel oil spill since the natural gas vapor
would dissipate into the air and not accumulate on the ground.
The non-corrosive characteristic of natural gas is favorable to prevent oxidation of
storage tanks and hence will reduce the possibility of contamination. Table 2.2
provides a comparison on the physical properties of compressed natural gas (CNG)
and conventional diesel fuel. As can be seen, natural gas comprises primarily of CH4
while the hydrocarbon chains in diesel are longer and more complex. CNG also
shows a lower molecular weight and specific gravity compared to diesel.
Research has shown that natural gas has a narrow combustion limit between 5 to 15
percent (Questar Gas, undated). This implies that combustion of natural gas will
only take place when concentration of natural gas in the air lies in the range
mentioned. Combined with its high ignition temperature, natural gas can be safely
used without the high risk of accidental explosion.
9
Compressed Natural
Conventional
Gas (CNG)
Diesel
Chemical Formula
CH4
C3 to C25
Molecular Weight
16.04
≈ 200
Carbon
75
84-87
Hydrogen
25
13-16
Specific Gravity
0.424
0.81-0.89
Density, kg/m3
128
802-886
Boiling temperature, °C
-31.7
188-343
Freezing point, °C
-182
-40-34.4
Flash point, °C
-184
73
Autoignition temperature, °C
540
316
Lower
5.3
1
Higher
15
6
Specific Heat, J/kg K
-
1800
Property
Composition by weight,%
Flammability limits, % volume
Table 2.2: Properties of Natural Gas and Diesel (Alternative Fuels Data Centre, 2004)
According to P.C. McKenzie Company (undated), the octane number of CNG is 120
compared to 87-93 of gasoline. The octane number measures the potential of
“knocking” in the engine due to fuel selection. A high octane number signifies a
higher resistance to engine “knocking” and increased efficiency of a smooth power
transmission. A direct comparison cannot be made with diesel fuel as it operates in a
compression-ignition engine and is measured with a property called cetane number.
10
2.2
ADVANTAGES AND LIMITATIONS OF CNG
2.2.1
Advantages
Barbotti CNG (2002) reported that CNG is the world’s cleanest operating fuel in
engines due to its low emission levels of nitrous oxides (NOx), carbon monoxide
(CO) and carbon dioxide (CO2) which contributes to the overall greenhouse effect
and global warming. Lewis (2005) added that the CNG is free of benzene and
therefore eliminates the health risk of consumers who may be directly exposed to the
carcinogenic material.
According to NGV.org (2001), the amount of total hydrocarbon (THC) and
Particulate matter 10 (PM10) are greatly reduced with the use of NGVs. The
environmental benefits are one of main reasons why most governments around the
world are promoting the use of CNG as fuel in consumer vehicles (Gwilliam, 2000).
Currently, the Malaysia government is also promoting the use of NGVs by
providing a 25% deduction in road tax for all vehicles running on CNG (Petronas
Dagangan Berhad, 2005).
Another main advantage of NGV is from the economics point of view. The present
price of natural gas is RM0.565 (AUD $0.195) per litre compared to RM1.20 (AUD
$0.414) per litre for petrol and RM0.881 (AUD $0.304) per litre for diesel (Petronas
Dagangan Berhad, 2005). This indicates a substantial 53% and 27% savings in fuel
cost respectively. NGV.org (2001) reported that the price of natural gas is also more
stable than other fuels. The massive cost savings of CNG fuel will definitely
encourage transportation companies and end users to consider purchasing dedicated
NGVs or switching to the alternative fuel.
With an abundant reserve of natural gas and network of dedicated piping systems, it
is convenient for NGV users to gain access to natural gas and refuel their vehicles by
just installing a home refueling system (NGV.org, 2001). Wide usage of natural gas
will also help reduce the dependence on finite petroleum fuels and avoid a steep
price increase in fuels.
11
Kojima (2001) researched that the use of natural gas in buses produces less noise
and vibrations compared to conventional fuel. This will lead to longer service life
and reduced maintenance costs. Fleet operators also reported a 40% savings on
maintenance costs since interval between vehicle check-ups is lengthened (Barbotti
CNG, 2002). Engine performance are also claimed to be superior to gasoline engines
since NGVs encounter less knocking and has a wide range of temperature tolerances
(Barbotti CNG, 2002).
2.2.2
Limitations
NGVs are mostly used in the fleet transportation industry compared to private
vehicle owners due to its high initial cost of engine conversion. According to
Natural Gas Vehicle Coalition (2005), this is caused by low production volumes of
NGVs to accommodate economies of scale. Although the government is paving the
way to encourage CNG usage, customer awareness still remains low due to vague
marketing strategies which lack focus (Natural Gas Vehicle Coalition, 2005).
Researches have shown a slight decrease in engine performance- around 10-15% in
CNG fuelled vehicles (Indian Energy Sector, 2000). Graham et al. (2000) researched
that the lower compression ratios with dedicated CNG engines compared to diesel
engines is the main reason for this power decrease. The spark ignition (SI) engine in
dedicated NGVs will not operate above a 11.5:1 ratio (Clean Air Power, undated)
but the problem may be resolved by using a dual-fuel engine which will be
discussed later.
Murray et al. (2000) revealed that another factor which causes NGVs to be
unpopular among consumers is the lack of refueling station available in most
countries. For instance in Malaysia, the availability of CNG refueling stations are
limited as only a subsidiary company, PETRONAS NGV is currently offering the
facility. From personal communication with NGV vehicle owners, the long
decompression and fill time of CNG fuel usually causes an outstretched queue in
refueling stations, much to the inconvenience of NGV owners.
12
Due to the gaseous form of CNG fuel, an accelerated wear of exhaust valves are also
encountered due to the drying effect (Indian Energy Sector, 2000).
2.3
SAFETY
In addition to its excellent emission quality where toxic gases are reduced and pose a
lower health hazard to the public, Barbotti CNG (2002) reported that the fuel
cylinders used to store CNG in vehicles are designed to withstand impact through
several tightly scrutinized tests. A survey conducted in the United States also
showed a 37% decrease in injury rate for NGVs compared to gasoline-powered fleet
vehicles and no fatality rates (Barbotti CNG, 2002).
The favorable physical property of CNG which enables it to dissipate into the air in
case of a leakage and thus avoiding contamination is also a safety advantage.
However, Graham et al. (2000) argued that CNG vapors formed at low temperatures
from leaks will generate large clouds of flammable vapor and increase the potential
of an explosion coupled with a spark.
Due to its high storage pressure at a range of 20-25 MPa, the refueling process is a
safety issue since 0.2-0.3 kWh of energy per cubic meter of natural gas is required to
compress it (Kojima, 2001). Kojima (2001) further exemplified a recent incident in
India where five people were injured in a NGV during refueling due to inferior gas
cylinder condition.
2.4
DIESEL ENGINE CONVERSION
2.4.1
Introduction to conventional diesel engine
Diesel engines are mostly used in heavy-duty applications and in fleet
transportations due to its higher engine efficiency achieved through the higher
compression ratio (CR). During operation, the compression ratios of diesel engines
13
can reach up to 20:1, compared to compression ratios of 8:1 for spark ignition
engines utilizing gasoline (Kojima, 2001).
Diesel engines in most heavy-duty vehicles are compression ignition engines, which
predominantly operates under a cycle comprising of four strokes:
i.
Induction stroke
The inlet valve is opened and air is forced into the combustion chamber
when the piston moves outwards through atmospheric pressure. As the
piston reaches its bottom dead centre, the intake valve closes.
ii.
Compression stroke
The piston then moves inwards to compress the air present in the
chamber. The air is heated to a temperature as high as 550°C (Shell
Canada Limited, undated) which is above the flashing point of the diesel
fuel. Just before the end of the stroke, fuel is injected into the heated and
compressed air.
iii.
Power stroke
The diesel fuel is ignited and the pressure created pushes the piston
outwards, providing power to the engine via the connecting rod and
crankshaft.
iv.
Exhaust stroke
The piston moves outwards and burnt gases are pushed out of the
combustion chamber through the exhaust valve. As the piston reaches the
top dead centre position, the cycle is repeated.
14
For direct injection diesel engines where diesel fuel is directly injected into the
combustion chamber at the end of the compression stroke, the combustion rates can
be calculated by applying the first law of thermodynamics for the quasi-static
(uniform pressure and temperature) control system (Heywood, 1988):
dQ
dV
dU
−p
+ m f hf =
dt
dt
dt
(2.1)
where
system
dQ
dt
p
boundary
dV
dt
= heat transfer rate across system boundary into the
= rate of work transfer done by the system due to
displacement
mf
= mass flow rate of diesel fuel into the system
hf
= sensible enthalpy of diesel duel
dU
dt
= rate of change of sensible internal energy of cylinder
contents
For heat-release analysis, equation (2.2) applies:
dQn dQch dQht
dV dU
=
−
=p
+
dt
dt
dt
dt
dt
(2.2)
where
dQn
dt
= net heat-release rate
15
dQch
dt
= gross heat-release rate
dQht
dt
= heat-transfer rate to the walls
p
dV
dt
= rate of work transfer done by the system due to
boundary displacement
dU
dt
= rate of change of sensible internal energy of cylinder
contents
If the contents of the cylinder are modeled to be an ideal gas, equation (2.2)
becomes:
dQn
dV
dT
=p
+ mcv
dt
dt
dt
(2.3)
Using ideal gas law, pV = mRT , with R to be constant, it follows that:
dp dV dT
+
=
p
V
T
(2.4)
Substitution of Equation (2.4) into (2.3) can be used to eliminate T:
dQn
c
dV c v dp
= 1+ v p
+ V
dt
R
dt
R dt
or
dQn
γ
dV
1
dp
=
p
+
V
dt
γ − 1 dt γ − 1 dt
(2.5)
16
where
γ
= ratio of specific heats
=
cp
cv
The range for γ for diesel-heat release is usually 1.3 to 1.35 for analysis.
2.4.2
Engine conversion types
2.4.2.1 Bi-fuel engine
A Bi-fuel engine utilizes two fuel systems, usually consisting NGV and gasoline. In
general, spark-ignition engines are easily converted into bi-fuel engines by
retrofitting a NGV kit to the engine system (Autoworld, 2004). A fuel selector
allows the user to choose which fuel to use.
According to Equitable Gas (undated), when natural gas is selected as the running
fuel, the compressed gas is passed through the master manual shut-off valve and
enters the engine compartment. A regulator in the engine compartment reduces the
gas pressure to 1 bar before the gas passes into the fuel-injection system through a
solenoid valve. When gasoline is selected as fuel, the natural gas system is shut off
to avoid any mixture of fuel.
Murray et al. (2000) reported that vehicles using the bi-fuel engine suffer from
power loss of around 10-15% when natural gas is used during wide open throttle.
2.4.2.2 Dedicated NGV engine
Dedicated NGV engines require more modification compared to a vehicle operating
with bi-fuel. Most of the components of a diesel engine need to be replaced as NGV
use a spark ignition engine to ignite its fuel. The gas supply system and ignition
17
system need to be changed and an electronic control unit (ECU) fitted to control the
operations for a dedicated NGV engine.
Autoworld (2004) outlined that the number of manufactured dedicated NGV in
Malaysia is still very low compared to retrofitted bi-fuel engines. According to
Alternative Fuels Data Centre (2004), dedicated NGV show better performance and
superior emissions since the engine system is optimized to run solely on natural gas.
Dedicated NGV only need to carry a single fuel load if compared with other types of
engines and this weight reduction increases the fuel efficiency.
2.4.2.3 Dual-fuel engine
In dual-fuel engines, natural gas and diesel fuel is used simultaneously in the
combustion chamber to produce power. Approximately 80% of natural gas is
consumed while the remainder comes from diesel which acts as a pilot fuel to ignite
the gas in the combustion chamber (Clean Air Power, undated).
To convert existing diesel engines to run on dual-fuel, an electronic control unit
(ECU) needs to be installed. The function of the ECU is to control engine speeds
while monitoring engine temperature and pressure by incorporating electronic and
mechanical sensors to ensure safe operation of the dual-fuel system. Clean Air
Power (undated) added that, in Caterpillar engines which are common in most heavy
duty applications, the ECU installed will communicate with the Caterpillar
Advanced Diesel Management System (ADEM) to determine the quantity and
timing of diesel pilot fuel according to the engine’s RPM signal. Apart from that, a
gas mixer needs to be fitted to the air manifold to allow complete mixture between
air and natural gas (Hybrid Fuel Systems Inc., 2004). In the Garretson fuel system,
the venturi principle is used to obtain the proper gas/air mixture with a sensitive and
properly calibrated fuel controller (Alternate Fuels Technologies Inc., undated).
A rack limiter is also installed to monitor the engine’s load and speed so that the
accurate amount of pilot diesel fuel can be supplied. Sensors and solenoids are
added together with diesel and natural gas injectors so that the injectors are
18
controlled by the ECU through pulse width modulated signals to maximize
efficiency.
The pistons and cylinder heads of the combustion chamber are also modified to
allow proper natural gas and air mixture so that the high compression ratio of the
original diesel engine can be maintained with dual-fuel. The cylinder head is
optimized to allow both diesel and natural gas injectors to operate for a standard
cycle.
Dual-fuel engines have the advantage of providing the same power as a conventional
diesel engine since it retains the high compression ratio and produces lower amounts
of emissions such as NOx and particulate matters. Hybrid Fuel Systems Inc. (2004)
further stated that dual-fuel NGVs have better fuel economy and lower maintenance
costs compared to a dedicated NGV engine.
19
CHAPTER 3
ENGINE PERFORMANCE
3.1
Torque
An engine’s torque is a measure of its rotational force exerted to transmit power
from the engine to the wheels of the vehicle through the drive train. The torque and
power produced by an engine can be measured using a dynamometer which is
mounted to the engine as a separate component. Torque can be improved by addition
of engine cylinders or increasing the capacity of the engine although an increase in
fuel consumption would be significant. The product of torque and angular speed
gives the power developed by the engine:
P=
2πNτ
60
(3.1)
Or
τ=
60 P
2πN
(3.2)
20
where
3.2
τ
=
torque
P
=
power developed by engine
N
=
engine speed
(N. m)
(W)
(rpm)
Input Power
The input power of the engine refers to the maximum rate at which energy is
supplied to the engine. It corresponds with the indicated power calculated from a p-
V diagram based on the work done during compression and expansion process of the
diesel cycle, less the heat loss to exhaust and coolants. The heat of combustion of
fuel is supplied to the engine and assuming the cycle efficiency as unity where all
the chemical energy of the fuel is converted into useful work, the engine input power
is given by
IP = m f × QHV × 10 3
(3.3)
where
IP
=
input power
(kW)
mf
=
mass flow rate of fuel
Q HV
=
lower calorific value of fuel
For diesel fuel, Q HV , Diesel
=
42.5 MJ/kg
For CNG fuel, Q HV ,CNG
=
45 MJ/kg
(kg/s)
(MJ/kg)
The lower calorific value of fuel is used in (3.3) since all water compounds in the
fuel are assumed to be in vapour phase without any condensation.
21
3.3
Brake Power
The brake power is the power output delivered by the engine shaft. It is less than the
indicated power since heat is lost to overcome the total friction generated in the
engine which is summed as friction power. Friction power consists of pumping
friction during intake and exhaust, mechanical friction in bearings, valves and
components such as oil and water pumps. Brake power refers to the rate at which
work is done and shows a maximum value when engine speed is increased close to
maximum before decreasing since friction becomes very significant at high engine
speeds.
Brake power =
Indicated power – Friction power
(3.4)
In a diesel engine, the brake power can be varied by changing the fueling rate or airfuel ratio to produce the desired power for an application. In the experimentation,
brake power is obtained from:
BP =
2πNτ
60 × 10 3
(3.5)
where
3.4
BP
=
brake power
(kW)
N
=
engine speed
(rpm)
τ
=
torque
(N. m)
Specific Fuel Consumption
Specific fuel consumption is the measure of fuel flow rate per unit power output and
relates to the fuel efficiency of an engine. It is inversely proportional to efficiency of
the engine as lower values of specific fuel consumption are favorable for higher
performance. Specific fuel consumption is defined as:
22
sfc =
mf
P
× 3.6 × 10 6
(3.6)
where
sfc
=
specific fuel consumption
mf
=
mass flow rate of fuel
P
=
power output
(g/kW. hr)
(kg/s)
(kW)
In the performance measurement and comparison between engines running on diesel
and dual-fuel for the experimentation, the power output measured is the brake
power. Therefore, brake specific fuel consumption is:
bsfc =
mf
BP
× 3.6 × 10 6
(3.7)
where
bsfc
=
brake specific fuel consumption
mf
=
mass flow rate of fuel
BP
=
brake power
(g/kW. hr)
(kg/s)
(kW)
The brake specific fuel consumption varies with the compression ratio and fuel
equivalence ratio. A higher compression ratio will produce lower bfsc since more
power can be extracted from the burning fuel. Bsfc decreases as engine size
becomes progressively smaller since heat losses from the combustion gas to the
cylinder wall are reduced. Generally, compression ignition engines with diesel fuel
produce a higher amount of energy per unit fuel compared to spark ignition engines.
23
3.5
Brake Mean Effective Pressure
The brake mean effective pressure is a useful measure of the relative performance of
an engine. It refers to the mean pressure to be maintained in the pistons of the
cylinder to produce a power output during each power stroke. The brake mean
effective pressure can be calculated from the torque and is defined as:
bmep =
BP × n R × 60
A× L × N × n
(3.8)
where
bmep =
brake mean effective pressure (kPa)
BP
=
brake power
nR
=
number of crank revolutions for each power stroke per
cylinder (one for two-stroke cycle and two for fourstroke cycle)
A
=
area of engine bore
L
=
length of engine stroke (m)
N
=
engine speed
n
=
number of cylinders
(kW)
(m2)
(rpm)
It indicates the work done per cycle for every unit of cylinder volume displaced and
is the direct measure of brake torque, not engine power. A higher bmep corresponds
to a higher engine output since more pressure is transmitted through the connecting
rods to the crankshaft. However, engine wear increases with increasing bmep and
leads to high mechanical stresses on engine components and imposing high thermal
stresses on combustion chambers.
24
The maximum value of bmep for a compression ignition engine is obtained at the
engine speed where maximum torque is obtained. The bmep of the same engine is
measured to be slightly lower at maximum rated power.
3.6
Engine Thermal Efficiency
Generally, an IC engine loses almost 42% of its energy to the exhaust system and a
further 28% to the cooling system (The Concept IC Engine). The engine thermal
efficiency refers to the ratio of work produced per cycle to the amount of fuel input
to the engine per cycle.
ηf =
Wc
Wc
P
=
=
m f QHV m f QHV m f QHV
ηf
=
engine thermal efficiency
P
=
output power produced per cycle
(kW)
mf
=
mass flow rate of fuel per cycle
(kg/s)
Q HV
=
lower calorific value of fuel
(3.9)
where
(MJ/kg)
In the experimentation, the desired output power per cycle is the brake power.
Therefore, incorporating equation (3.3) into (3.9),
ηf =
BP
IP
(3.10)
25
Another method of obtaining the engine thermal efficiency is by utilization of the
specific fuel consumption. Substituting equation (3.6) into (3.9) gives:
ηf =
3600
sfc × Q HV
(3.11)
where
sfc
=
specific fuel consumption
(g/kW. hr)
QHV
=
lower calorific value of fuel
(MJ/kg)
We can see that the specific fuel consumption is inversely proportional to the total
engine thermal efficiency, as mentioned earlier.
26
CHAPTER 4
EMISSION
4.1
Carbon Monoxide (CO)
Carbon monoxide (CO) is a colourless, odorless, flammable and highly poisonous
gas which is less dense than air. Inhalation of carbon monoxide can be fatal to
humans since a small concentration as little as 0.1% will cause toxication in the
blood due to its high affinity to oxygen carrying hemoglobins. Exposure levels must
be kept below 30 ppm to ensure safety (Environmental Centre, undated). Apart from
that, carbon monoxide also helps in the formation of greenhouse gases and global
warming by encouraging the formation of NOx.
Carbon monoxide forms in internal combustion engines as a result of incomplete
combustion when a carbon based fuel undergoes combustion with insufficient air.
The carbon fuel is not oxidized completely to form carbon dioxide and water. This
effect is obvious in cold weathers or when an engine is first started since more fuel
is needed.
Carbon monoxide emission from internal combustion engines depend primarily on
the fuel/air equivalence ratio ( λ ). Figure 4.1(a) shows the variation of CO emission
for eleven fuels with different hydrocarbon contents. A single curve may be used to
represent the data when using the relative air/fuel or equivalence ratio as represented
in Figure 4.1(b).
27
(a)
Figure 4.1:
(b)
Variation of CO emission for different fuels (a) with air/fuel ratio; (b) with relative
air/fuel ratio (Heywood, 1988)
Both the graphs clearly show that the amount of CO emitted increases with
decreasing air to fuel ratio. Spark ignition gasoline engines which normally run on a
stoichiometric mixture at normal loads and fuel-rich mixtures at full load shows
significant CO emissions. On the other hand, diesel engines which run on a lean
mixture only emit a very small amount of CO which can be ignored. Ferguson
(1986) researched that additional CO may be produced in lean-running engines
through the flame-fuel interaction with cylinder walls, oil films and deposits. Directinjection diesel engines also emit more CO than indirect-injection engines.
However, the CO gas emission increases with increasing engine power output for
both engines.
28
CO formed from hydrocarbon radicals can be oxidized to form carbon dioxide in an
oxidation reaction, in an equilibrium condition (Turns, 1996):
CO + OH = CO2 + H
(4.1)
The emission of CO is a kinetically-controlled reaction since the measured emission
level is higher than equilibrium condition for the exhaust. Three-body radical
recombination reactions such as
H + H + M = H2 + M
(4.2)
H + OH + M = H 2 O + M
(4.3)
H + O2 + M = HO2 + M
(4.4)
are found to be rate-controlling reactions for emission of CO gas.
Reduction of carbon monoxide in internal combustion engines can be achieved by
improving the efficiency of combustion process or utilization of oxidation catalysts
to oxidize carbon monoxide to carbon dioxide. Engine modifications such as
improved cylinder head design, controlled air intake and electronic fuel injection can
help to maintain a lean air/fuel mixture which is favorable.
4.2
Total Hydrocarbon (THC)
Total hydrocarbon (THC) is used to measure the level of formation of unburnt
hydrocarbons caused by incomplete combustion in the engine. The hydrocarbons
emitted may be inert such as methane gas or reactive to the environment by playing
a major role in the formation of smog. The types hydrocarbons emitted from the
exhaust greatly depend on the type and composition of fuel used. Heywood (1988)
added that fuels with a greater concentration of aromatics and olefins compounds
will result in a higher percentage of reactive hydrocarbons.
29
In diesel engines, hydrocarbon emission can be significant under two normal
operating conditions due to the complex nature of fuel-air and burned-unburned gas
mixing in the combustion chamber. Firstly, a mixture of fuel leaner than the lean
combustion limit in the chamber during the ignition delay period will cause
incomplete combustion and hence formation of unburnt hydrocarbons. The locally
overlean mixture of fuel will not autoignite or support a propagating flame, causing
a slow reaction to develop.
On the other hand, undermixing of fuel which occurs when the fuel mixture is too
rich to ignite or support a flame causes hydrocarbon formation during the
combustion cycle. The injector sac volume provides an important contribution to the
hydrocarbon emission in direct-injection engines. The diesel fuel left at the tip of the
injector enters the cylinder at low velocity and does not have enough time to achieve
a standard mixture with air. Experimental results from Heywood (1988) showed that
the amount of hydrocarbon emission is proportional to the injector sac volume.
Generally, hydrocarbon emission in diesel engines are higher during engine idling
and at light loads when the amount of excess air is great (Ferguson, 1986).
Suitable diesel catalysts can be used to oxidize hydrocarbons to carbon dioxide and
water molecules. According to Nett Technologies Inc. (undated), hydrocarbon traps
are also used to capture hydrocarbon emissions especially at low temperatures when
the oxidation catalysts are not functional during engine idling times.
4.3
Oxides of Nitrogen (NOx)
Nitrogen oxides consist primarily of nitric oxide (NO) and nitrogen dioxide (NO2) as
a product of oxidation of atmospheric nitrogen in the combustion chamber. Diesel
fuel contains a significant amount of nitrogen compounds and acts as an additional
source of NO. Formations of nitric oxides from molecular nitrogen are described by
the following equations (Heywood, 1988):
O + N 2 ↔ NO + N
(4.5)
30
N + O2 ↔ NO + O
(4.6)
N + OH ↔ NO + H
(4.7)
Nitric oxides (NO) formed in the combustion chamber can be rapidly oxidized to
form NO2 through the following reaction:
NO + H 2 O → NO2 + OH
(4.8)
At the same time, NO2 will be subsequently converted back to NO via:
NO2 + O → NO + O2
(4.9)
A considerable amount of NO2 is found in diesel engines especially during light
loads or engine idling times. At lower temperatures, the transformation of NO2 back
to NO in reaction (4.9) is quenched by the cooler regions of the chamber and the
ratio of NO2 to NO can go as high as 30% (Heywood, 1988).
Figure 4.2:
Ratio of NO2 to NO in diesel exhaust for varying engine load and speed (Heywood,
1988)
31
Figure 4.2 clearly shows that the maximum amount of NO2 is emitted in a diesel
engine at low engine speed and minimum loads. This can be damaging to the
atmosphere as NO2 formed will contribute to formation of ground-level ozone or
smog and reacts in the air to form corrosive nitric acid.
At the same engine speed, Ferguson (1986) reported that the amount of NOx
produced increases with engine load for a direct-injection engine and the maximum
amount occurs when the fuel mixture is slightly lean. With higher loads, the peak
pressures of the cylinder and temperature distribution are higher and coupled with
enhanced mixing of the diesel fuel, NOx levels are increased. As a rule of thumb, the
emission of NO are roughly proportional to the mass of fuel injected.
Selective catalytic reduction (SCR) can be used to convert NOx emitted to form
oxygen and nitrogen through the use of reducing agents such as ammonia or urea. It
is combined with an oxidation catalyst to oxidize any traces of ammonia which may
escape the system into the atmosphere.
4.4
Particulate Matters (PM)
Particulates are defined as a complex aggregate of solid and liquid material other
than water that can be collected in an exhaust filter. Diesel particulate matters
(DPM) are generally divided into three categories which include dry carbon particles
or soot, Soluble Organic Fractions (SOF) resulting from incomplete combustion,
adsorption and condensation of heavy hydrocarbon onto carbon particles, and sulfate
fraction (SO4).
In addition to these elements, small amounts of zinc, sulphur, phosphorus, calcium,
iron, chromium and silicon are found in particulates. The organic fractions of
particulates are serious hazards to the health of the public and environment. The
primary carbon particles have a nuclei size of around 0.04 to 1 µm which have
adverse health effects when respirated.
32
Generally, the density of PM depends on the engine load, speed and exhaust
conditions. Heywood (1988) mentioned that the highest PM concentration occurs at
the core region of diesel fuel spray in direct-injection engines when the local
equivalence ratio is very rich. Soot concentration decreases with increasing distance
from the centerline.
A compromise must be considered when designing ways to reduce the amount of
PM formation during combustion process since a higher combustion level and
increased in-cylinder temperatures will result in a higher concentration of NOx
formation. A trap oxidizer is used to filter diesel particulate matters in the exhaust as
a method to reduce emission after the combustion process. The trapped particulates
are oxidized to remove them and refresh the filter. Catalysts are used to improve the
filter efficiency by increasing the regeneration capacity of the filters.
4.5
Carbon Dioxide
Carbon dioxide emissions in diesel engines are products of direct combustions or
by-product of oxidizing other unwanted emission gases with the aid of catalysts.
Although diesel engines generally produce low amounts of CO2 compared to other
emission gases, the emission of carbon dioxide must be regulated and controlled to
reduce negative impacts on the environment such as accumulation of greenhouse
gases and global warming.
4.6
Oxides of Sulphur (SOx)
Conventional diesel fuel contains sulphur compounds in its composition and results
in the emission of sulphur oxides in the form of SO2 and SO3 as the products of
combustion. According to Turns (1996), the average content of sulphur compounds
in diesel fuel lies within the range of 0.1 to 0.8% which is very high compared to
gasoline. Sulphur trioxide (SO3) reacts readily with water to form sulfuric acid
33
which accumulates in the exhaust system. On the other hand, sulphur dioxide (SO2)
will be oxidized in the atmosphere to form SO3 before reacting to form sulfuric acid.
Apart from assisting in the corrosion of exhaust metal, the sulfuric acid present
destroys the effectiveness of catalysts used in the exhaust system to reduce other
emission pollutants.
The smog produced from sulphur content in diesel fuel generally cause
environmental and health hazards, and has been liked with respiratory diseases and
illnesses. The sulphur content in the exhaust can be reduced by using low sulphur
diesel fuels which reduces soot emission without affecting the engine performance.
The oxides of sulphur present in the exhaust can be reduced by using limestone
(CaCO3) or lime (CaO) to produce calcium precipitates and carbon dioxide
according to the following equations:
For limestone,
CaCO3 + SO2 + 2 H 2 O → CaSO3 ⋅ 2 H 2 O + CO2
(4.10)
CaO + SO2 + 2 H 2 O → CaSO3 ⋅ 2 H 2 O
(4.11)
For lime,
According to Ly (2002), the sulphur content of natural gas is much lower than
ultralow sulphur diesel levels of 10-50 ppm. Natural gas also does not contain toxic
benzene or 1,3-butadiene compounds. This greatly reduces any oxides of sulphur
from being produced in the combustion process and eliminates the irritating odor of
sulfuric gases.
34
CHAPTER 5
EXPERIMENTAL SETUP
5.1
Engine Preparation
A four stroke, single cylinder air cooled diesel engine is used for the experiment and
the full engine specifications are provided in Appendix A. The engine speed runs up
to 3600 rpm. The engine is attached at one end to the electrical heating element
dynamometer with a drive shaft coupling flange. A throttle is used to control and
increase the speed of the engine as the control variable. The dynamometer and
engine is cooled by an attached cooling tower with cooling fan and heat sump to
dispose generated heat. The other end of the dynamometer is hooked up to the
digital readout system which contains the digital RPM meter, flowmeter, oil sump
temperature and torque meter so that experimental readings can be obtained.
5.2
Preparation of Load Banks
The load bank for the engine which consists of electrical heaters is installed
underneath the digital readout system and connected to the engine. The load bank is
rated at 320 V and is used to provide a 1 kW load increment so that the response of
the diesel engine can be measured. The electrical heaters are immersed in water so
that heat power output from the engine is used to increase the temperature of the
water and care must be taken so that no water spills occur which may short-circuit
the wiring system.
35
5.3
Calibration of Digital Thermocouple
A set of digital thermocouple is used to obtain the temperatures of the engine both at
the body and exhaust. The probe of the thermocouple is inserted into the exhaust of
the engine and on the surface of the engine body respectively to measure the steadystate temperature when the engine runs in normal operating conditions for different
values of loads. Since temperatures fluctuate slightly, a range of readings are taken
over a period of time before finalizing the mean value.
5.4
Gas Analyzer Setup
The emission gas levels are tested and measured using the Autologic Autogas
Emission Analyzer as shown in Figure 5.1. A portable 5 Gas Emission Analyzer
with PC software unit is used for this experiment to measure levels of HC, CO, O2,
CO2, and NOx. Additional units for measurements of RPM, Oil temperature and
Diesel Smoke Meters can be combined to the main gas analyzer unit. The analyzer
performance exceeds standards such as the ASM/BAR 97, OIML, and BAR 90.
Other accessories for the gas analyzer includes a 25 foot sampling hose attached to a
sensor probe, durable high strength aluminium casing for protection of all the
connections, automatic water removal system so that water vapors that may get
trapped will not interfere with the results, and a compatible software for all version
of Windows operating system to allow data to be recorded. The full range of
specifications of the gas analyzer is listed in Appendix C.
The typical warm-up time for the gas analyzer is 2 minutes and it possesses a great
level of accuracy. Initially, the gas analyzer settings are configured and reset to zero
for both diesel and dual-fuel so that consistent results are obtained. The probe of the
gas analyzer is inserted into the exhaust duct of the engine and all connections are
tightened before measurements are taken. The first filter unit for the gas analyzer is
cleaned after every test run for standardization of results.
36
Fig 5.1: Autologic Autogas Gas Analyzer (Autologic Company)
5.5
Natural Gas Conversion Kit
A natural gas conversion kit system is available in the laboratory to enable
compressed natural gas to be used as fuel in the test engine. The conversion kit is of
model TMB Tartarini Natural Gas Conversion Kit and consists of components such
as the gas pressure regulator, solenoid valve, air mixer, fuel select switch and gauges
as shown in Figure 5.2.
37
Fig 5.2: Natural Gas Converter Kit
5.5.1
Natural Gas Pressure Regulator
The pressure regulator decreases the CNG pressure to near atmospheric pressure
from 20MPa in the storage tank to allow natural gas to flow into the gas mixer.
Apart from that, the regulator also acts as a control to modulate the flow of natural
gas to the gas mixer.
5.5.2
Natural Gas Solenoid Valve
The solenoid valve is used as a main control switch to allow natural gas to flow from
the pressure regulator to the gas mixer and engine. The solenoid valve is electronic
timer controlled and contains a built-in gas filter and attached pressure gauge. It also
acts as an emergency shut off device to stop the flow of natural gas into the system
when a leak is detected or when other devices malfunction. Apart from that, the
solenoid valve improves ignition during cold temperature start up of the engine.
38
5.5.3
Natural Gas Mixer
Natural gas is mixed with air in the gas mixer to obtain the optimum ratio for
combustion before being transferred into the combustion chamber of the engine
through the control panel which measures the flow rate of the mixture.
5.5.4
Fuel Selector Switch and Gauge
A fuel selector switch is used for the user to switch between diesel and dual-fuel
system in the experiment. The amount of natural gas in the system can also be
monitored through the attached measurement gauge.
39
Fig 5.3:
Experimental Setup Overview
40
5.6
Experimental Procedures
Figure 5.3 shows the experimental setup to perform the test on diesel and dual-fuel
operating conditions. The engine is connected to the dynamometer through a direct
coupling. The dynamometer is then connected to the main control board which
controls the intensity of the electrical loads applied to the engine.
The effect of diesel and dual-fuel natural gas with pilot diesel fuel on the engine
performance and emission levels are to be investigated. Initially, diesel fuel is used
and the engine is started and no loads are applied for five minutes to allow the
engine to reach a steady-state operating condition. Next, the throttle of the engine is
adjusted so that the engine speed is maintained at 1600 RPM. The measurements for
torque, volume flow rate, body and exhaust temperature of the engine are recorded
when the values reach a steady-state condition. The volume flow rate is calculated
by recording the time taken for the engine to use 10ml of fuel in the flowmeter
attached to the control panel. After calibration, the gas analyzer probe is inserted
into the exhaust duct of the engine and measurements of the levels of emission gases
are recorded. The probe is removed from the exhaust duct after measurements are
taken and cleaned.
After that, the electrical load bank is applied to the engine at 1kW using the selector
switches on the control panel. The same readings are taken from the digital readout
system and gas analyzer. The engine is loaded with additional 1kW loads
progressively and measurements are recorded at the same engine speed until
maximum load occurs, in which the engine fails to support the applied load and
stalls.
The above procedure is repeated for engine speeds of 1800, 2000, 2200, and 2600
RPM so that effective comparisons can be made. For each speed, the engine is
loaded until the maximum condition is reached. Before the performance and
emission test for dual-fuel is conducted, the engine is allowed to cool to room
temperature to obtain standardized results.
41
For the dual-fuel experiment, diesel pilot fuel is supplied to warm the engine for five
minutes. After that, the master shut-off valve of CNG storage tank is opened to
allow natural gas to travel to the gas regulator through the high pressure fuel line.
Natural gas is then supplied to the engine through the on-board control panel and the
mass flow rate of natural gas is recorded using a flow meter. The amount of diesel
pilot fuel is kept constant while the engine speed is controlled by increasing the flow
rate of natural gas to the engine until the specified engine speeds are obtained.
All experimental data for engine speeds of 1600, 1800, 2000, 2200, and 2600 rpm
are recorded and tabulated.
42
CHAPTER 6
RESULTS AND DISCUSSION
6.1
Introduction
The results from the experiments performed on the four-stroke engine for maximum
load operating conditions are shown below in graph form and discussed. The results
are comparable with Talal et al. (2003). For engine performance, the graphs of
engine body temperature, exhaust gas temperature, engine torque, brake power,
brake specific fuel consumption (bsfc) and engine thermal efficiency against varying
engine speeds from 1600 rpm to 2600 rpm are plotted.
On the other hand, the graphs of emission of carbon monoxide (CO), hydrocarbon
(HC), oxides of nitrogen (NOx), carbon dioxide (CO2), and excess oxygen (O2)
against the same range of engine speed are shown.
43
6.2
Engine Performance
Conditions
for
Maximum
Load
Operating
The experimental data was taken at an ambient temperature of 29°C after the engine
was started for five minutes to achieve a stable operating condition. Electrical loads
of 1KW each were progressively added and the results below show the comparison
graphs for both conventional diesel fuel and dual-fuel readings and their respective
explanations under maximum loading conditions for the respective engine speeds.
The body and exhausts temperatures were measured with a thermocouple when the
temperatures become stable.
Engine Body Temperature
6.2.1
Figure 6.1 shows the measured engine body temperatures for both diesel and dualfuel against engine speeds for maximum load operating conditions. The engine body
temperature for diesel fuel is typically higher for all the engine speeds tested
compared to dual-fuel.
Body Temperature under maximum load operating conditions
Body Tem perature (°C)
200
180
160
Diesel
140
Dual-Fuel
120
100
80
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.1:
Engine body temperatures under maximum load operating conditions
44
The maximum difference of temperature occurs at an engine speed of 1800 rpm
where the body temperature for diesel fuel is approximately 42.4°C higher than the
corresponding temperature recorded for dual-fuel under the same loading conditions
and engine speed. On average, the engine body temperature of diesel is 29.0°C
higher than dual-fuel.
It is observed from Figure 6.1 that an increase in engine speed from 1600 rpm to
2600 rpm will effectively increase the body temperature as well. The amount of fuel
injected into the combustion chamber is increased when engine speed is increased
by adjusting the throttle of the engine. As more fuel enters the chamber, combustion
takes place at a higher temperature resulting in an increase in measured engine body
temperature.
The low temperatures measured for dual-fuel body temperature compared to diesel
indicates a higher level of oxygen concentration for combustion since lower
temperatures will cause an increase in density of air. With more oxygen gas present,
dual-fuel can undergo a better combustion and produce higher engine efficiency.
6.2.2
Exhaust Gas Temperature
The exhaust gas temperatures measured with a thermocouple for both diesel and
dual-fuel are plotted in Figure 6.2. Similar to the engine body temperature, the
exhaust temperature follows a similar trend where the temperature increases as the
engine speed is increased. The exhaust gas temperature of diesel fuel is also higher
than dual-fuel, with an average difference of 20.3°C. The measured values also
shows that the temperature difference between diesel and dual-fuel is almost
constant at approximately 7% higher for diesel fuel for engine speeds ranging from
1600 rpm to 2600 rpm.
The high exhaust temperature of diesel engines compared to dual-fuel will result in
unfavorably higher NOx output. The higher exhaust temperature for diesel fuel
operation also indicates that combustion of dual-fuel is leaner than diesel since less
heat is usually produced during lean combustion in a compression ignition engine.
45
Exhaust Temperature under maximum load operating conditions
Exhaust Tem perature (°C)
340
320
300
280
Diesel
260
Dual-Fuel
240
220
200
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.2:
6.2.3
Engine exhaust temperatures under maximum load operating conditions
Engine Torque
Figure 6.3 shows the measured engine torque for diesel and dual-fuel operations.
The torque increases for engine speeds ranging from 1600 rpm to 2100 rpm to a
peak before dropping in magnitude for higher speeds up to 2600 rpm for both diesel
and dual-fuel. It is observed from Figure 6.3 that the measured value from the
dynamometer shows a higher torque for dual-fuel operation for all engine speeds.
Analytically, the average percentage increment using dual-fuel is 12.4% compared
to diesel fuel with the latter displaying a decrease of 1 Nm torque at all the engine
speeds taken.
As engine speed is increased from 1600 rpm to 2100 rpm, the measured torque for
diesel and dual-fuel increases indicating a higher fuel flow rate into the combustion
chamber and an increased energy input available to the engine. The combustion
efficiency must also be taken into account when measuring the output torque
generated by the engine.
46
Engine Torque under maximum load operating conditions
11
Engine Torque (Nm )
10
9
Diesel
8
Dual-Fuel
7
6
5
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.3:
Engine torque output under maximum load operating conditions
The engine output torque drops tremendously for engine speed of 2600 rpm due to
engine knocking. The incorrect ignition timing which results in engine knocks
results in uncontrolled combustion and vibrations which will decrease the overall
torque output available. The high pressure in the combustion chamber due to the
compression stroke will cause CNG to self-ignite to the design temperature and
pressure with high compression ratio. In addition, as the engine speed increases, the
relative air to fuel ratio decreases causing more fuel to be injected into the system.
Accumulation of excess fuel in a high pressured combustion chamber with high
temperatures favors engine knocking, especially at very high loading conditions.
6.2.4
Engine Brake Power
From Figure 6.4, the engine brake power produced for dual-fuel operation is higher
than diesel fuel operation with a maximum value difference of 0.272 kW recorded at
the engine speed of 2600 rpm. In average, the dual-fuel engine produces a higher
engine brake power by as much as 0.213 kW compared to conventional diesel fuel.
47
Brake Power under maximum load operating conditions
2.5
Brake Power (kW)
2.0
1.5
Diesel
Dual-Fuel
1.0
0.5
0.0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.4:
Engine brake power under maximum load operating conditions
As explained earlier, the engine brake power produced decreases slightly due to
engine knocks at the speed of 2600 rpm for both diesel and dual-fuel. The dual-fuel
operation achieved a maximum brake power output of 2.304 kW at speed of 2200
rpm, which is almost 83% of the rated engine output power of 2.8 kW.
6.2.5
Brake Specific Fuel Consumption
Figure 6.5 shows the comparison of brake specific fuel consumption under
maximum load conditions for diesel and dual-fuel operations. The fuel consumption
per unit power output for dual-fuel is lower than diesel fuel with the maximum
occurring at 2600 rpm with 17.1% reduction. On an average basis, dual-fuel
operation reduces the brake specific fuel consumption by 15.6% for all the engine
speeds tested. This phenomenon can be attributed to the chemical properties of
natural gas where the higher octane value of CNG compared to diesel decreases the
amount of fuel required for combustion to drive the engine to support the same
amount of loading.
48
Brake Specific Fuel Consumption under maximum load operating
conditions
Brake Specific Fuel
Consum ption (g/kW.hr)
400.0
350.0
300.0
Diesel
Dual-Fuel
250.0
200.0
150.0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.5:
Brake specific fuel consumption under maximum load operating conditions
It is also observed that the specific fuel consumption decreases when the engine
speed is increased from 1600 rpm to around 2200 rpm, and then shows a slight
increment for speeds above 2200 rpm. At low speeds, the higher values of brake
specific fuel consumption can be explained by the higher frictional forces acting on
the piston and thus consuming more fuel since more heat energy is lost to friction.
The fuel conversion efficiency improves when the engine speed is increased by
reference to the decrease in brake specific fuel consumption for engine speeds from
1600 rpm to 2200 rpm.
By analyzing Figure 6.5, the increase in brake specific fuel consumption for engine
speeds higher than 2200 rpm is mainly due to the engine knocks at 2200 rpm and
2600 rpm. Engine knocks generally decreases the combustion efficiency of the
engine and greatly increases the friction since fuel ignitions are not synchronized,
causing more fuel to be consumed.
49
6.2.6
Engine Thermal Efficiency
The engine thermal efficiency for diesel and dual-fuel operation is depicted in
Figure 6.6 below. From the graph, it can be concluded that the thermal efficiency of
the engine running on dual-fuel is higher compared to diesel fuel for engine speeds
ranging from 1600 rpm to 2600 rpm. This is similar with the result obtained from
Talal et al. (2003). From Figure 6.6, the maximum engine efficiency achieved for
dual-fuel and diesel occurs at the engine speed of 2200 rpm at approximately 36.0%
and 32.5% respectively.
Engine Thermal Efficiency under maximum load operating
conditions
38.0
Overall Effeciency (% )
36.0
34.0
32.0
30.0
Diesel
28.0
Dual-Fuel
26.0
24.0
22.0
20.0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.6:
Engine thermal efficiency under maximum load operating conditions
The improvement in engine thermal efficiency for dual-fuel at an average of 3.6%
can be explained by the higher heating value of natural gas which produces more
heat for combustion for an equivalent mass flow rate compared to diesel.
Apart from that, according to Petronas Dagangan (2005), the current price of natural
gas fuel stands at RM 0.565 (AUD $0.185) per litre compared to the diesel price of
RM 1.281 (AUD $0.442) per liter. Neglecting the small amount of pilot diesel fuel
50
used in dual-fuel operation, the use of natural gas as a substitute for diesel fuel offers
a huge savings of almost 56% coupled with beneficial performance characteristics
such as increased overall torque and brake power output as well as providing an
increase in engine thermal efficiency.
At higher loads where more torque and power is generated, the engine thermal
efficiency improves for both diesel and dual-fuel operations since higher mechanical
efficiency is achieved at higher combustion temperatures. Apart from that, the
higher rate of utilization of fuel and higher air to fuel ratio greatly increases the
combustion rate of the engine since more air and fuel is introduced to the system.
At engine speeds higher than 2200 rpm, the engine efficiency thermal begin to
decrease slightly due to the effects of engine knocks which not only causes output
power losses but also damages the engine. The design of the engine needs to be
modified to prevent engine knocking and substantially improve fuel economy. Early
fuel detonation can be prevented by the design of a smaller and fast-burn motion
combustion chambers and decreasing the air to fuel ratio of the mixture since
combustions with rich mixture produces less heat.
51
6.3
Exhaust Emission for Maximum Load Operating Conditions
A portable gas analyzer which is capable of measuring the amounts of CO, HC,
NOx, CO2, and O2 is used to measure the exhaust emission of the engine running
under diesel fuel and dual-fuel for varying engine speeds. The gas analyzer was
calibrated after each measurement to obtain a high accuracy of the readings taken.
6.3.1
Carbon Monoxide
Figure 6.7 shows the emission characteristics of diesel fuel and dual-fuel operation.
As can be seen, a significant reduction in carbon monoxide emission can be
achieved by running the engine with dual-fuel operation. At lower engine speeds
ranging from 1600 to 2000 rpm, the reduction is more apparent with an
improvement of up to 70% when the engine is running at 1600 rpm.
Generally, the amount of CO present in the exhaust indicates incomplete combustion
in the combustion chamber, mostly due to insufficient air or cold engine
temperatures. The level of CO increases with decreasing air to fuel ratio which
indicates that a rich mixture of fuel will produce more CO pollutants. Under
maximum load operating conditions, diesel fuel produces a high level of CO gas
since the combustion of diesel fuel is richer than dual-fuel.
For diesel fuel, it is observed that a general decrease in CO emission for engine
speeds ranging from 1600 rpm to 2300 rpm. This result is in agreement with the
increasing body and exhaust temperature recorded for diesel fuel operation. A higher
body and exhaust temperature at high levels of engine speed indicates a higher
combustion temperature and hence a more complete combustion and produces less
CO gas together with increased power output. At a high speed of 2600 rpm, the CO
emission level becomes higher due to engine knocking which results in lower
combustion efficiency. This is also due to the fact that engine knocks become
apparent when the engine is operating under high load conditions. The mixing time
52
of diesel fuel and air inside the chamber is greatly reduced when engine knocking
occurs.
CO Emission under maximum load operating
conditions
140
CO (ppm)
120
100
80
Diesel
60
Dual-Fuel
40
20
0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.7:
Emission of CO under maximum load operating conditions
For dual-fuel, the recorded CO emission is lower if compared to diesel fuel
operation for all engine speeds. Natural gas produces a greater combustion
efficiency leading to lower amounts of CO since natural gas in its gaseous state
usually contain less contaminants than diesel fuel. Apart from that, turbulent mixing
between natural gas and air in the engine produces a higher quality mixture since
both are in gaseous phase, thus producing a lower carbon monoxide level generally.
CO emission for dual-fuel operation increases slightly with increasing engine speed
since the residence times of fuel in the combustion chamber is decreased at high
engine speeds, causing higher CO formation. This result shows that dual-fuel
operation is able to achieve a better combustion compared to diesel fuel under
maximum load conditions.
Although the level of emission of CO is greatly reduced by using dual-fuel, the
quantity produced is still above the safety level of 30 ppm. This problem can be
53
solved by using catalytic converters to oxidize the poisonous CO into CO2 and water
or by providing regular maintenance on the engine particularly in cleaning the air
filters and the fuel delivery system.
6.3.2
Unburnt Hydrocarbons
Similar to CO, the amount of unburnt HC present in the exhaust indicates a poor
combustion and excess unburnt fuel which usually occurs when the fuel mixture is
rich. From Figure 6.8, the emission level of HC for dual-fuel operation is
substantially higher than diesel fuel. At a low engine speed of 1600 rpm, the level of
HC emission recorded for dual-fuel is almost five times the amount for diesel. The
reason for this difference is in the composition of natural gas which consists
primarily of methane. The uncontrolled flow of natural gas into the combustion
chamber produces an excess of fuel and the unburnt methane gas accounts for the
majority of HC emissions collected at the exhaust. Overfueling of engine with
natural gas produces a high level of HC since the oxygen gas is insufficient for
complete combustion with a low air to fuel ratio.
For diesel fuel, the level of HC emission records a slight increase with engine speed
but is still at a much lower level than dual-fuel. With increasing load, the amount of
HC produced in the emission will decrease since greater combustion efficiency can
be achieved with increased temperature. This explains the low level of HC for diesel
fuel under maximum load operating conditions.
From the graph, it can be observed that the level of HC for dual-fuel decreases with
engine speed. The level of HC drops because a more complete combustion is present
at high engine speeds, which is in agreement with the exhaust temperature data
measured. The high levels of unburnt hydrocarbon which contains mainly methane
gas for dual-fuel operation poses a great risk since methane gas is a very powerful
greenhouse gas with almost twenty times the global warming effect of carbon
dioxide. Therefore, a suitable oxidation catalyst must be used to counter the effects
of high HC emission of dual-fuel engine. The unburnt hydrocarbon can be easily
54
converted to carbon dioxide and water by retrofitting an oxidation catalyst to replace
the muffler of a vehicle.
HC (ppm)
HC Emission under maximum load operating
conditions
4500
4000
3500
3000
2500
2000
1500
1000
500
0
1500
Diesel
Dual-Fuel
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.8:
Emission of HC under maximum load operating conditions
Another factor which causes engines operating with dual-fuel to produce a higher
level of hydrocarbons is over-scavenging of the cylinder. Diesel engines operate by
igniting the fuel at the end of the compression stroke and usually the exhaust gas is
forced out of the cylinder through blow scavenging, which is operated via an
auxiliary pump to allow clean air to enter the chamber. Since the scavenging process
involves only air, fresh natural gas and air may be forced out into the exhaust when
over-scavenging occurs in dual-fuel operation. The longer time period during overscavenging causes unburnt hydrocarbons to escape into the exhaust port and
increases the overall level of recorded HC.
55
6.3.3
Oxides of Nitrogen (NOx)
As observed from Figure 6.9, the NOx level for dual-fuel operation is considerably
lower than conventional diesel fuel for engine under maximum load conditions. The
use of dual-fuel is favorable in reducing NOx emission levels by an average of 58%.
Generally, an increase in NOx emission is observed when the engine speed is
increased from 1600 rpm to 2600 rpm for both fuels.
NOx (ppm)
NOx Emission under maximum load operating
conditions
200
175
150
125
100
75
50
25
0
1500
Diesel
Dual-Fuel
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.9:
Emission of NOx under maximum load operating conditions
The formation of NOx is largely dependent on the peak temperature in the
combustion chamber as well as the concentration of oxygen and nitrogen gas from
the air intake. Both conditions increase the level of NOx formation in the exhaust
which leads to formation of smog and acid rains. The overall rise of NOx as engine
speed increases for both diesel and dual-fuel corresponds to the increase in body and
exhaust temperature in Figure 6.1 and 6.2. The higher combustion temperature and
pressure in the chamber favors the chemical reaction to form NO and NO2 with
enough oxygen.
56
Engines running with dual-fuel produce less NOx than diesel since diesel fuel
contains high volatile nitrogen compounds in their composition which contributes to
a higher level of nitrogen concentration in the combustion chamber. Since diesel
engines operate primarily in the lean region when diesel fuel is consumed, there is
excess air and oxygen for the nitrogen compounds to form NOx when the
combustion temperature is high.
From Figure 6.9, a steep decline in NOx level for diesel fuel operating at engine
speed of 2600 rpm is observed. This may be caused by the low residence time for
fuel in the combustion chamber at this speed. At high engine speeds, the time for
NO and O2 compounds to react is severely shortened and decreases the overall NOx
emission. Another reason for this occurrence is the low amount of O2 present in the
exhaust for diesel fuel at 2600 rpm. The low O2 concentration of 10.5% at 2600 rpm
indicates a richer combustion with less air in the chamber. Hence, the level of NOx is
greatly reduced due to insufficient air.
The low emission of NOx for dual-fuel engines can be attributed to several factors.
Firstly, the premixed combustion is less intense and produces less activation energy
for nitrogen and oxygen compounds to disintegrate and form NO. The reduced
mixing of air and fuel also lowers the oxidation rate of NO to NO2 in the chamber.
Apart from that, the lower exhaust temperatures present in the dual-fuel system as
indicated in Figure 6.2 compared to diesel fuel reduces the NOx production level.
Finally, the concentration of O2 is reduced in the chamber due to the presence of
gaseous natural gas fuel, which will displace an equal amount of air.
6.3.4
Carbon Dioxide
From Figure 6.10, dual-fuel operation under maximum load operating conditions
produces less CO2 compared to diesel fuel by an average of 1.16%. The most
significant reduction occurs at engine speed of 2200 rpm where dual-fuel provides
reduction of 2.2% in CO2 emission.
57
Generally, CO2 emission levels indicate the quality and composition of fuel that is
being used for combustion. The CO2 emission levels are lower for dual-fuel since
natural gas has a lower carbon content compared to the complex and longer
hydrocarbon chains of diesel. Natural gas consists mainly of simple methane
hydrocarbon in its composition and has less carbon molecules.
Apart from that, the ratio of carbon to hydrogen affects the amount of CO2
production in a typical combustion. Almost 88% of natural gas comprises of
methane with the chemical formula of CH4 while the general formulation for diesel
fuel is CnH1.8n. It is clearly seen that the ratio of carbon to hydrogen atoms in natural
gas is 1 : 4 compared to 1 : 1.8 in diesel which indicates that the carbon content in
diesel fuel is much higher. Therefore, the CO2 concentration is higher for diesel
operation for all range of engine speeds.
CO2 Emission under maximum load operating conditions
6
CO2 (%)
5
4
Diesel
Dual-Fuel
3
2
1
0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.10:
Emission of CO2 under maximum load operating conditions
From Figure 6.10, a trend in CO2 emission is also observed where an increase in
engine speed will in turn cause the CO2 level to rise gradually for both diesel and
dual-fuel operations. This engine behaviour can be explained by the increase in fuel
58
intake for combustion as the engine approaches higher speeds to support the load
connected to it. The addition of fuel to the chamber causes the level of CO2 emission
to be higher at high engine speeds since more carbon molecules are introduced into
the engine.
6.3.5
Excess Oxygen
The level of excess O2 in the exhaust is shown in Figure 6.11. As can be seen, the
level of excess O2 for diesel fuel is higher than dual-fuel for maximum load
conditions. A higher percentage of O2 in the exhaust corresponds to leaner
combustion where the air to fuel ratio is higher than 1. Diesel engines normally
operate at a lean point of stoichiometric for diesel fuel and cause more air to be
present in the combustion chamber and exhaust, producing a higher level of unused
excess O2.
Excess O2 under maximum load operating
conditions
20
O2 (%)
18
16
Diesel
14
Dual-Fuel
12
10
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.11:
Excess of O2 under maximum load operating conditions
59
When the engine operates using dual-fuel, the level of excess O2 decreases by an
average of 1.9% over the range of speeds from 1600 rpm to 2600 rpm compared to
diesel. This result implicates that combustion of dual-fuel is better and leads to
lower levels of excess O2 left in the exhaust since majority of the oxygen supplied is
used for the combustion process.
This result is in agreement with the increasing CO2 emission level shown in Figure
6.10 for both diesel and dual-fuel operating conditions. At higher engine speeds, the
combustion becomes more complete to support the higher torque and output power
levels and produces more CO2 and less excess O2 in the exhaust.
Figure 6.11 also shows a general decline in O2 percentage as engine speed increases.
60
6.4
Engine Performance for Moderate Load Operating
Conditions
The engine performance data for moderate load operating conditions are taken at a
load of 3 kW for every engine speed for both diesel and dual fuel to enable a fair
comparison for the engine performance aspects. Similarly, the torque readings are
taken from the attached dynamometer, the mass flow rate of diesel from the flow
meter and a digital thermocouple is used to obtain the engine body and exhaust
temperatures when the readings become constant.
Engine Body Temperature
6.4.1
From Figure 6.12, it is observed that the engine body temperature of dual-fuel
operating system is generally lower than diesel fuel.
Body Temperature under moderate load operating conditions
130
Body Tem perature (°C)
120
110
100
90
Diesel
80
Dual-Fuel
70
60
50
40
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.12:
Engine body temperatures under moderate load operating conditions
61
On average, the body temperature of diesel fuel is 9.48 °C higher than dual-fuel for
the range of engine speeds tested. The body temperature of dual-fuel increases at a
higher rate than diesel fuel as the engine speed is increased.
The general increase in body temperature at higher engine speeds is caused by an
increase in fuel intake when the throttle is adjusted to increase the engine speed. The
additional fuel input to the engine produces combustion at a higher temperature
albeit still lower than the maximum load condition. On average, the body
temperature for moderate load is 58.6 °C lower for diesel fuel and 39.1 °C lower for
dual-fuel compared to maximum loading conditions. This result shows that there is a
substantial difference in combustion temperature and amount of fuel and air intake
between a similar engine operating under moderate load at 3 kW and maximum load
at a range of 7-8 kW.
6.4.2
Exhaust Gas Temperature
Exhaust Temperature under moderate load operating conditions
Exhaust Tem perature (°C)
240
220
200
180
Diesel
160
Dual-Fuel
140
120
100
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.13:
Engine exhaust temperatures under moderate load operating conditions
62
For the exhaust gas temperature, the general increase in temperature as the engine
speed climbs is also observed. From Figure 6.13, it is shown that the rate of increase
of exhaust temperature in diesel fuel operation is greater than dual-fuel as the engine
speed is increased from 1600 rpm to 2600 rpm. The maximum temperature
difference between diesel and dual-fuel occurs at an engine speed of 2200 rpm with
52.9 °C difference.
Similar to the engine body temperature, the exhaust temperatures for both fuels
under moderate loading conditions are lower than the maximum load, with an
average value of 101.1 °C lower for diesel and 122.8 °C lower for dual-fuel at all
engine speeds.
As mentioned before, the high exhaust temperature for diesel fuel indicates a higher
percentage of NOx production in the combustion products. It is also a direct result of
lean combustion associated with diesel fuel in CI engines which generates more heat
in the combustion process.
6.4.3
Engine Torque
From Figure 6.14, the engine torque measured from the dynamometer shows that
rotational force exerted by the engine to provide power to the loads is higher for
dual-fuel compared to diesel. An increase of 1 Nm of torque for every measured
engine speed is recorded. This shows that the energy input for dual-fuel operation is
higher than diesel fuel since more power is produced.
Similar to the maximum load operating condition, the engine torque for moderate
load increases for engine speeds ranging from 1600 rpm to 2200 rpm for both
operating conditions as the fuel intake is increased and higher combustion
temperature is achieved. Engine knocks results in a slight drop of torque values at
the engine speed of 2600 rpm for both diesel and dual-fuel, although the effect is not
as apparent as the maximum load operating condition.
63
Engine Torque under moderate load operating conditions
9
Engine Torque (Nm )
8
7
Diesel
6
Dual-Fuel
5
4
3
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.14:
Engine torque output under moderate load operating conditions
In comparison, the engine torque produced for moderate loading is lower by an
average of 29.2% for diesel fuel and 26.0% for dual-fuel for all engine speeds
compared to the maximum load operating condition, mainly due to the reduced fuel
intake at lower engine loads.
6.4.4
Engine Brake Power
The engine brake power comparison for both diesel and dual-fuel is shown in Figure
6.15. Proportional to the engine torque, the brake power produced by dual-fuel is
higher than diesel for engine speeds ranging from 1600 rpm to 2600 rpm. The
maximum brake power achieved is 1.634 kW for diesel fuel and 1.906 kW for dualfuel, both at an engine speed of 2600 rpm. The 14.3 % increment in engine brake
power is significant together with its lower specific fuel consumption and emissive
gases to be a suitable alternative fuel to replace diesel. The engine brake power
shows a typical slight decrease compared to maximum load conditions for both
fuels.
64
Brake Power under moderate load operating conditions
Brake Power (kW)
2.0
1.5
Diesel
1.0
Dual-Fuel
0.5
0.0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.15:
6.4.5
Engine brake power under moderate load operating conditions
Brake Specific Fuel Consumption
Figure 6.16 shows the brake specific fuel consumption for moderate load conditions
for diesel and dual-fuel. The bsfc for dual-fuel is lower than diesel from low to high
engine speeds by an average of 19.5%. The higher octane rating of CNG fuel
compared to diesel decreases the amount of fuel needed for combustion, and hence
lower brake specific fuel consumption for dual-fuel is noticed.
The bsfc for both the fuels decrease gradually when the engine speed is increased.
At lower engine speeds, the high frictional forces acting on the piston and cylinder
walls results in a greater fuel consumption rate since the heat of combustion is lost to
friction. This effect is greatly reduced when the engine speed rises above 2000 rpm,
which is the conventional engine speed for the engine of any conventional consumer
vehicle during motion.
65
Brake Specific Fuel Consumption under moderate load operating
conditions
Brake Specific Fuel
Consum ption (g/kW.hr)
400.0
350.0
300.0
Diesel
Dual-Fuel
250.0
200.0
150.0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.16:
Brake specific fuel consumption under moderate load operating conditions
By comparing with the maximum load condition, the bsfc for moderate load is
generally higher by an average of 17.8% for diesel and 12.3% for dual-fuel taking
into account for all the engine speeds tested. The increase in bsfc can be explained
by the lower mechanical efficiency of the engine for lower load conditions where
there is a higher percentage of heat loss due to the lower utilization of fuel. Apart
from that, at lower loading conditions, the lower combustion temperatures justified
by the lower body and exhaust temperatures, and lower air to fuel ratio ( λ ) results
in a lower combustion rate and therefore a lower fuel conversion efficiency.
6.4.6
Engine Thermal Efficiency
Similar to the engine at maximum loading conditions, Figure 6.17 shows that the
engine thermal efficiency for dual-fuel in generally higher than diesel. On average,
the thermal efficiency for dual-fuel is 4.48% higher for moderate load compared to
3.60% for maximum load. This difference can be explained by the lower rate of
engine knocks at lower engine loads since the engine is not burdened to operate at its
maximum potential.
66
Engine Thermal Efficiency under moderate load operating
conditions
36.0
Overall Effeciency (% )
34.0
32.0
30.0
Diesel
28.0
Dual-Fuel
26.0
24.0
22.0
20.0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.17:
Engine thermal efficiency under moderate load operating conditions
The maximum engine efficiency achieved for moderate load conditions for both
diesel and dual-fuel is generally lower than the maximum load operating conditions.
Calculations show that the average drop in thermal efficiency associated with
moderate loading to the engine is 4.32% for diesel and 3.47% for dual-fuel. This
phenomenon can be described by the lower fuel utilization and air to fuel ratio when
the engine load is low, causing the combustion rate to be greatly reduced in
comparison.
67
Exhaust Emission for Moderate Load Operating Conditions
6.5
The discussions for the exhaust emission measured using the gas analyzer is
categorized into different pollutants for both type of fuels. The emission levels of
CO, HC, NOx, CO2, and O2 are compared and discussed between diesel and dualfuel and the effects of load intensity on the emission are also included.
6.5.1
Carbon Monoxide
The emission of carbon monoxide for diesel and dual-fuel when the engine is
running under moderate loading is shown in Figure 6.18. Resembling the maximum
load conditions, the CO emission levels for dual-fuel is considerably lower than
diesel due to the cleaner contents of natural gas fuel. The average reduction level of
CO emission by using the dual-fuel operation stands at 58.0% with the maximum
occurring at the minimum engine speed of 1600 rpm where the reduction level goes
up as high as 79.4%.
CO Emission under moderate load operating
conditions
120
CO (ppm )
100
80
Diesel
60
Dual-Fuel
40
20
0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.18:
Emission of CO under moderate load operating conditions
68
The initial high levels of CO for diesel fuel can be related to the engine body and
exhaust temperatures from Figures 6.12 and 6.13. The lower engine temperatures
indicate lower combustion efficiency since combustion takes place at a lower
temperature. This can be related to the enormous amount of CO production for
diesel fuel during engine idling or cold-startup.
In comparison, the CO emission levels of both the fuels are lower compared to the
maximum loading conditions as the combustion process takes place in the leaner
region. The excess amount of air present in the chamber during moderate load
conditions prevents carbon monoxide buildup since fuel combustion is complete
with sufficient oxygen levels.
The richer mode of combustion when the engine speed increases causes a small
increase in CO emission for both the fuels since less air is being supplied to the
same volume of fuel.
6.5.2
Unburnt Hydrocarbons
The level of unburnt hydrocarbon is also associated with incomplete combustion and
excess of unburnt fuel. From Figure 6.19, the level of HC emission for dual-fuel
operation is extremely high compared to diesel fuel especially at low engine speeds
below 1800 rpm. At a speed of 1600 rpm, dual-fuel produces a vast increase of HC,
at around 145% higher than diesel. The uncontrolled natural gas flow input into the
combustion chamber explains this occurrence. At lower loads and engine speeds,
most of the natural gas is not completely utilized for the combustion process and
most of the methane in the natural gas escapes into the exhaust to produce a high
level of unburnt hydrocarbon detection.
At moderate engine loads, the level of HC produced is by average 96.0% higher for
diesel and 10.5% higher for dual-fuel compared to the maximum load operating
conditions. This result is justified by the lower combustion temperature present for
moderate loading which causes incomplete combustion and excess fuel to be forced
into the exhaust system.
69
HC Emission under moderate load operating
conditions
6000
HC (ppm)
5000
4000
Diesel
3000
Dual-Fuel
2000
1000
0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.19:
Emission of HC under moderate load operating conditions
As mentioned earlier in the maximum load operating conditions discussions, the
level of HC emission reduces with increasing engine speed due to the increase in
quality of combustion for higher engine speeds due to the increasing combustion
temperature. It is thus shown that the level of unburnt hydrocarbon is heavily
dependent on the combustion temperature compared to the air to fuel ratio.
6.5.3
Oxides of Nitrogen (NOx)
Figure 6.20 shows the level of NOx produced by both the fuels under moderate
loading conditions. It is clear that the use of dual-fuel significantly reduces the
harmful NOx gases due to its lower nitrogen composition compared to diesel. The
reduction by an average of 68.2% is a direct result of lower combustion
temperatures for dual-fuel which impedes NOx production.
The steady increase in NOx emission as the engine speed increases can also be
attributed to the increase in engine combustion temperature. A higher ignition
temperature favors the chemical reactions to form NOx when enough air and time is
70
present. The emission levels increase by 110% for diesel and 228% for dual-fuel
when then engine speed is increased from 1600 rpm to 2600 rpm.
NO x (ppm )
NOx Emission under moderate load operating
conditions
160
140
120
100
80
60
40
20
0
1500
Diesel
Dual-Fuel
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.20:
Emission of NOx under moderate load operating conditions
By making comparisons with the same engine speeds, the average percentage of
NOx emission under moderate loading is 10.7% lower for diesel and 22.5% lower
for dual-fuel compared to the maximum loading conditions, due to the influence of
combustion temperature as well.
6.5.4
Carbon Dioxide
The level of CO2 emission for diesel fuel is generally higher than dual-fuel from
Figure 6.21. The maximum level of carbon dioxide emission for diesel fuel occurs at
the engine speed of 2600 rpm with 4.12% emission while the dual-fuel operation
recorded 2.29% maximum at the same engine speed. The highest CO2 emission is
recorded at the maximum engine speed measured because the fuel intake is at its
highest.
71
CO2 Emission under moderate load operating
conditions
5
CO2 (%)
4
3
Diesel
2
Dual-Fuel
1
0
1500
1700
1900
2100
2300
2500
2700
Engine Speed (rpm)
Figure 6.21:
Emission of CO2 under moderate load operating conditions
The higher carbon content and percentage for diesel fuel accounts for its higher CO2
emission at all range of engine speeds. In comparison, the level of CO2 emission is
reduced by an average of 0.65% when running the engine at moderate loading
conditions compared to the maximum load where the fuel intake is maximum.
6.5.5
Excess Oxygen
Figure 6.22 shows the level of excess oxygen present in the exhaust after
combustion for both fuels under 3 kW loading condition. The higher level of excess
O2 detected implies a lower combustion efficiency since most of the oxygen is not
used up when there is an excess of fuel from the unburnt hydrocarbon readings.
The data of excess oxygen for diesel and dual-fuel records only a slight difference
between them, with only a 0.13% percentage difference on average. Generally, as
the combustion temperature rises with engine speed, the level of excess oxygen
decreases since more air is being used for combustion. The higher level of excess
72
oxygen present during moderate loading also proves that the combustion rate is
lower compared to maximum loading.
O 2 (%)
Excess O2 under moderate load operating conditions
20
19
18
17
16
15
1500 1700 1900 2100 2300 2500 2700
Diesel
Dual-Fuel
Engine Speed (rpm)
Figure 6.22:
6.6
Excess of O2 under moderate load operating conditions
Concluding Discussion
The average data values for engine speed ranging from 1600 rpm to 2600 rpm for
both maximum and moderate load operating conditions are tabulated below for
comparison between diesel and dual fuel.
Table 6.1 shows the engine performance summary for maximum loading which
shows an improvement in engine torque, engine brake power and thermal efficiency
while a reduction of specific fuel consumption is observed. On the other hand, Table
6.2 sums up the emission characteristics for maximum operating load which shows a
reduction in carbon monoxide, oxides of nitrogen and carbon dioxide. It is also
observed that the level of unburnt hydrocarbon emission increased by almost two
and a half fold when dual-fuel is used.
73
Percentage
Increase
Engine Performance
Diesel
Dual-Fuel
Engine Torque (Nm)
8.20
9.20
12.20
Engine Brake Power (kW)
1.75
1.96
12.23
Brake Specific Fuel Consumption (g/kW.hr)
288.75
243.43
-15.70
Engine Thermal Efficiency (%)
29.62
33.25
3.63
Table 6.1:
Engine Performance for maximum load
Diesel
Dual-Fuel
Percentage
Decrease
Carbon Monoxide (ppm)
102
55
46.08
Unburnt Hydrocarbon (ppm)
918
3200
-248.58
Oxides of Nitrogen, NOx (ppm)
121
52
57.02
Carbon Dioxide (%)
3.61
2.45
32.13
Excess Oxygen (%)
14.49
12.59
13.11
Exhaust Emission
Table 6.2:
Exhaust Emission for maximum load
On the other hand, Table 6.3 shows the engine performance summary for the
experiment conducted with moderate loading conditions. The increase in engine
torque, brake power and engine efficiency is observed as well, similar to the
maximum load operating condition. There is more improvement in the engine
performance for the moderate load operating condition as the engine thermal
efficiency is improved by an amount as high as 4.48%. Apart from that, the exhaust
emission for moderate loading conditions records a better outcome with a further
reduction in the toxic and greenhouse gases while the unburnt hydrocarbon level is
only increased by 122% when dual-fuel is used.
74
Engine Performance
Diesel
Dual-Fuel
Percentage
Increase
Engine Torque (Nm)
5.80
6.80
17.24
Engine Brake Power (kW)
1.26
1.47
16.94
Brake Specific Fuel Consumption (g/kW.hr)
337.35
271.11
-19.64
Engine Thermal Efficiency (%)
25.30
29.78
4.48
Table 6.3:
Engine Performance for moderate load
Diesel
Dual-Fuel
Percentage
Decrease
81
35
56.79
Unburnt Hydrocarbon (ppm)
1608.4
3570.4
-121.98
Oxides of Nitrogen, NOx (ppm)
102.6
33.8
67.06
Carbon Dioxide (%)
2.94
1.82
38.10
Excess Oxygen (%)
17.51
17.38
0.74
Exhaust Emission
Carbon Monoxide (ppm)
Table 6.4:
Exhaust Emission for moderate load
In general, for both maximum and moderate load, the engine thermal efficiency is
greater for dual-fuel operation compared to diesel and more engine torque and brake
power is produced. Apart from that, there is a hefty decrease in major pollutant such
as carbon monoxide, oxides of nitrogen (NOx) and carbon dioxide. A lower level of
excess oxygen signifies a higher combustion rate for dual-fuel. However, the level
of unburnt hydrocarbon in dual-fuel operation for both maximum and moderate load
record an incredibly 250% and 120% increase respectively. This is a direct result of
uncontrolled natural gas flow into the engine which produces methane gas
accumulation in the exhaust.
6.7
Catalytic Aftertreatment
The high level of unburnt hydrocarbon produced for dual-fuel operation for both
moderate and maximum loading can be controlled by using catalytic aftertreatment
75
techniques which uses noble metals such as platinum, palladium and rhodium to
allow reactions to take place that oxidizes carbon monoxide and hydrocarbons to a
safe level (Turns 1996). Washcoat such as alumina and promoters are added to
increase the reaction level and efficiency of oxidation. The absence of sulphur in
natural gas fuel allows the use of catalytic aftertreatments since sulphur compounds
are known to interfere with the catalytic action and reduce the efficiency.
A three-way catalyst which controls the pollutant levels of three emission gases
namely carbon monoxide, hydrocarbon and oxides of nitrogen can be used in the
aftertreatment. The main criterion which governs the feasibility of this three-way
catalyst is the air-fuel ratio which must be near stoichiometric value for an effective
reduction in all three pollutant gases as shown in Figure 6.23.
Figure 6.23:
Conversion efficiency of three-way catalyst as a function of air-fuel ratio
(Ferguson 1986)
76
Since CNG fuel produces a low level of oxides if nitrogen (NOx) emission due to its
low nitrogen content in its fuel, the lean operating condition of diesel engines will
not cause a major problem since the catalyst conversion efficiency for both
hydrocarbon and carbon monoxide is still high for lean air-to-fuel ratio as shown in
Figure 6.23. The reduction of unburnt hydrocarbon and carbon monoxide remain the
primary concern in using CNG fuel to reach the safety emission level imposed by
the authorities.
Ammonia (NH3) and hydrogen sulphide (H2S) are both by-products of three-way
catalysts generated by the reduction of oxides of nitrogen and sulphur dioxide
respectively which are common in normal diesel engines. A high catalyst
temperature and rich fuel mixture promotes the formation of both the by-products
which causes irritation and discomfort to humans. The low level of NOx emission
coupled with the zero sulphur content of natural gas fuel causes the emission of
ammonia and hydrogen sulphide gases to be low. The lower exhaust temperature for
dual-fuel operation further reduces the catalyst by-products to a negligible level.
Apart from that, the high accumulation level of methane in the engine due to the
uncontrolled intake of natural gas must be controlled through an automated valve
which comprise of sensors of the gaseous fuel level required for combustion so that
the amount of hydrocarbon in the exhaust can be reduced. Hydrocarbon filters and
coalescers in stand alone vessels can also be set up in the natural gas fuel line supply
to separate solid hydrocarbon and moisture in the air to allow efficient operation of
the engine and its lubrication systems. Good engine maintenance and tune-up is also
necessary to reduce the level of emission gases in the exhaust of the CNG fuel
system.
77
CHAPTER 7
CONCLUSION
7.1
Achievement of Objectives
The main objectives of the Project Specification in Appendix A were achieved. A
compression ignition standard diesel engine was successfully converted to utilize
CNG fuel by retrofitting a conversion kit. Dual-fuel operation which utilizes diesel
as a pilot fuel as a source of ignition for natural gas proves to be an excellent
substitute for conventional diesel fuel with its superior engine performance and
exhaust emission characteristics.
The results obtained from the experiments conducted on both the maximum and
moderate load operating conditions show substantial reduction in carbon monoxide,
oxides of nitrogen, carbon dioxide, and excess oxygen levels in the exhaust. Apart
from that, the engine performance is also improved when using dual-fuel operation
with increased engine torque, brake power and engine efficiency alongside an
improvement in specific fuel consumption. The extreme levels of unburnt
hydrocarbons produced when utilizing dual-fuel operation due to the uncontrolled
fuel input to the engine is controlled using catalytic aftertreatment process which
reduces the level of carbon monoxide and oxides of nitrogen simultaneously through
application of a three-way catalyst.
78
The price of natural gas which costs less than half the price of conventional diesel
fuel provides massive savings in terms of fuel costs and the payback period for the
initial CNG conversion kit cost is reduced with these huge savings.
7.2
Recommendation and Future Work
The engine performance characteristics comparison between conventional diesel
fuel and dual-fuel can be improved by consideration of the heat loss of combustion
to the cylinder walls and also the friction generated which contribute to the total
energy loss of the system through use of measurement instruments such as a
calorimeter. The corrected engine output power can then be obtained to achieve a
more accurate value of the overall efficiency of the engine.
Devices which are able to detect other emission gases such as particulate matter
(PM), oxides of sulphur, ammonia, hydrogen sulphide and other chemical
compounds in the exhaust of the engine can be used to monitor the emission levels
of both fuels so that a more comprehensive study on the emission characteristics can
be performed to ensure that no unusually high pollutant levels are associated with
dual-fuel operation.
Apart from that, the catalytic aftertreatment chosen can be installed on the exhaust
of the engine to allow emission data to be collected so that the reduction of the high
level of unburnt hydrocarbon can be monitored and the apparatus can be modified if
necessary to ensure that the safety standards are reached. Software simulation on the
engine performances and the process of combustion in the engine can be used to
design the optimum engine to incorporate dual-fuel by varying parameters such as
the air-fuel ratio, compression ratio and other engine parameters.
79
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12) Rashidi, M n.d., ‘Compressed Natural Gas Vehicles’, University of Shiraz,
Iran, viewed 12 March 2005, <http://succ.shirazu.ac.ir/~motor/ngv1.htm>
13) Sigall, J 2000, ‘Analysis of Alternative Fuel Technologies for New York City
Transit Buses’, New York, United States of America, viewed 20 April 2005,
<http://www.cleanairnet.org/infopool/1411/articles35609_analysis_alternative.pdf>
14) Traver, M.L. n.d., ‘Emission Formation in Compression Ignition Engines’,
College of Engineering and Mineral Resources, West Virginia, United States
of America, viewed 22 May 2005,
<http://www2.cemr.wvu.edu/~englab/Tutorials/EmissTut/diesel.html>
15) Union Gas 2004, Union Gas Limited, Ontario, Canada, viewed 29 March
2005, <http://www.uniongas.com/aboutus/aboutng/ngv/ngvhistory.asp>
16) United States Environmental Protection Agency 2002, ‘Clean Alternative
Fuels: Compressed Natural Gas’, viewed 29 March 2005,
<http://www.epa.gov/otaq/consumer/fuels/altfuels/420f00033.pdf>
85
APPENDIX A
PROJECT SPECIFICATION
86
University of Southern Queensland
FACULTY OF ENGINEERING AND SURVEYING
ENG 4111/4112 Research Project
PROJECT SPECIFICATION
FOR:
WONG, Wei Loon
TOPIC:
COMPRESSED NATURAL GAS AS AN ALTERNATIVE
FUEL IN DIESEL ENGINES
SUPERVISORS:
Dr. Harry Ku (USQ)
Dr. Fok Sai Cheong (USQ)
Dr. Talal Yusaf (UNITEN)
ENROLMENT:
ENG 4111 – S1, X, 2005
ENG 4112 – S2, X, 2005
PROJECT AIM:
This project aims to investigate the effects of using a dual-fuel
mixture of compressed natural gas (CNG) and diesel in a
compression ignition (CI) engine performance
PROGRAMME:
Issue B, 23rd April 2005
a) Research the history of CNG usage worldwide and a literature review on the
engine performance using CNG as the main fuel supply inclusive of the
advantages and limitations.
b) Conversion of the current CI engine to install the CNG fuel system to enable
the use of dual fuel diesel-CNG engine.
c) Study on the effect of using CNG as fuel in terms of power, torque, brake
specific fuel consumption (BSFC), and thermal efficiency. Perform a
comparison analysis on the dual fuel combustion and conventional diesel
fuel.
d) Examine the emission data collected for both fuels and conduct feasibility
study on CNG as a fuel alternative in terms of pollution and economy.
As time permits,
e) Conduct a Matlab program to effectively analyze the heat and mass transfer
and emission data for comparison with experimental results
f) Develop a CFD FLUENT analysis on the CNG flow rate and its effects on
engine performance
87
AGREED:
_______________ (Student) _____________, _____________ (Supervisors)
__ / __ / __
__ / __ / __
__ / __ / __
88
APPENDIX B
ENGINE SPECIFICATIONS
89
Model
Y170F Vertical 4 Stroke Diesel Engine
Cooling
Air cooled single cylinder
Combustion System
Direct Injection
Bore
70 mm
Stroke
55 mm
Displacement
211 cc
Maximum Engine Speed
3600 rpm
Maximum Output
2.8 kW (3.755hp)
Continuous Output
2.5 kW (3.353hp)
Net Weight
27 kg
Dimensions
324 x 410 x 416 mm
Table B.1
Engine Specifications
90
APPENDIX C
GAS ANALYZER SPECIFICATIONS
91
Model
Part No
Autologic Autogas Portable 5 Gas
Emission Analyzer with software
310-0120
Measuring Item
HC, CO, O2, CO2, and NOx
Method
Non-dispersive Infrared (NDIR)
Range
Warm-up Time
HC
0-30000 ppm
CO
0-15%
0-25%
O2
CO2
0-25%
0-5000 ppm
NOx
HC
1 ppm
CO
0.001 vol%
O2
0.01 vol%
CO2
0.01 vol%
NOx
1 ppm
0-90 % of 8 seconds for NDIR
measurements
Less than 2 minutes
Operating Temperature
0-50 °C
Humidity Level
Up to 95% (non-condensing)
Altitude
-300 to 2500 m
Vibration
1.5G Sinusoidal 5-1000 Hz
Power Supply
AC 90-230 V, 50-60 Hz
Resolution
Response Time
Table C.1
Gas Analyzer Specifications
92
APPENDIX D
EXPERIMENTAL DATA
93
Engine Speed
1600
Time for 10ml fuel
(s)
78
Volume Flow Rate
(m³/s)
1.28x10-7
Body Temp (°C)
124.6
Exhaust Temp
(°C)
248.8
Torque (Nm)
7
Brake Power (kW)
1.1730
Mass flow rate
(kg/s)
1.13x10-4
Input Power (kW)
4.7949
BSFC (g/kW.hr)
346.2483
BMEP (kPa)
415.5844
Efficiency
24.4639
Table D.1
2000
2200
2600
69
64
55
58
1.45x10-7
140.8
1.56x10-7
149.6
1.82x10-7
165.3
1.72x10-7
177.9
255.5
9
1.6967
259.6
9
1.8852
298.1
9
2.0737
330.7
7
1.9061
1.28x10-4
5.4203
270.6052
534.3228
31.3024
1.38x10-4
5.8438
262.5716
534.3228
32.2601
1.60x10-4
6.8000
277.7617
534.3228
30.4959
1.52x10-4
6.4483
286.5503
415.5844
29.5606
Engine Performance Data for Diesel Fuel under Maximum Load
Operating Conditions
Engine Speed
Time for 10ml fuel
(s)
Diesel Mass flow
rate (kg/s)
Body Temp (°C)
Exhaust Temp
(°C)
Torque (Nm)
Brake Power (kW)
CNG Mass flow
rate (kg/s)
Input Power (kW)
BSFC (g/kW.hr)
BMEP (kPa)
Efficiency
Table D.2
1800
1600
1800
2000
2200
2600
290
249
230
202
272
3.03x10-5
93.9
3.53x10-5
98.4
3.83x10-5
125.9
4.36x10-5
144.7
3.24x10-5
150.2
225.8
8
1.3406
237.4
10
1.8852
240.9
10
2.0947
287.5
10
2.3041
299.8
8
2.1785
1.35x10-4
4.7949
287.3534
474.9536
27.9588
1.54x10-4
5.4203
231.0534
593.6920
34.7804
1.66x10-4
5.8438
224.2024
593.6920
35.8446
1.92x10-4
6.8000
237.1270
593.6920
33.8843
1.74x10-4
6.4483
237.4353
474.9536
33.7835
Engine Performance Data for Dual-Fuel under Maximum Load
Operating Conditions
94
Engine
Speed
1600
1800
2000
2200
2600
Table D.3
Diesel
CO
HC
NOX CO2
O2
CO
HC
NOX
(ppm) (ppm) (ppm) (%)
(%) (ppm) (ppm) (ppm)
122
758
70
2.37 17.56
36
4015
15
117
566
96
2.99 17.25
42
3657
26
83
922
155 3.15 14.32
55
3094
50
88
1150
180 5.35 12.87
68
2816
78
99
1196
106 4.18 10.47
74
2420
91
CO2
(%)
1.29
1.68
2.46
3.17
3.64
Exhaust Data Comparison for Diesel and Dual-Fuel under Maximum
Load Operating Conditions
Engine Speed
Time for 10ml fuel
(s)
Volume flow rate
(m3/s)
Body Temp (°C)
Exhaust Temp
(°C)
Torque (Nm)
Brake Power (kW)
Mass flow rate
(kg/s)
Input Power (kW)
BSFC (g/kW.hr)
BMEP (kPa)
Efficiency
Table D.4
Dual Fuel
1600
1800
2000
2200
2600
130
80
71
59
68
7.69x10-8
71.3
1.25x10-7
79.8
1.41x10-7
86.8
1.69x10-7
105.4
1.47x10-7
121.9
125.7
4
0.6703
153.6
6
1.1311
187.5
6
1.2568
201.5
7
1.6129
218.8
6
1.6338
6.77x10-5
2.8769
363.5607
237.4768
23.2990
1.10x10-4
4.6750
350.0955
356.2152
24.1951
1.24x10-4
5.2676
355.0264
356.2152
23.8590
1.49x10-4
6.3390
332.9105
415.5844
25.4440
1.29x10-4
5.5000
285.1456
356.2152
29.7062
Engine Performance Data for Diesel Fuel under Moderate Load
Operating Conditions
95
O2
(%)
15.68
13.14
12.74
10.78
10.62
Engine Speed
Time for 10ml fuel
(s)
Diesel Mass flow
rate (kg/s)
Body Temp (°C)
Exhaust Temp
(°C)
Torque (Nm)
Brake Power (kW)
CNG Mass flow
rate (kg/s)
Input Power (kW)
BSFC (g/kW.hr)
BMEP (kPa)
Efficiency
Table D.5
Engine
Speed
1600
1800
2000
2200
2600
Table D.6
1600
1800
2000
2200
2600
462
354
340
322
300
1.90x10-5
58.7
2.49x10-5
64.5
2.59x10-5
78.6
2.73x10-5
100.1
2.93x10-5
115.9
105.9
5
0.8379
119.6
7
1.3196
135.4
7
1.4663
148.6
8
1.8433
168.0
7
1.9061
8.19x10-5
2.8769
275.9697
296.8460
29.1237
1.27x10-4
4.6750
284.2620
415.8544
28.2276
1.42x10-4
5.2676
288.1396
415.8544
27.8355
1.67x10-4
6.3390
275.6569
474.9536
29.0789
1.50x10-4
5.5000
231.5298
415.8544
34.6572
Engine Performance Data for Dual-Fuel under Moderate Load
Operating Conditions
Diesel
Dual Fuel
CO
HC
NOX CO2
O2
CO
HC
NOX CO2
O2
(ppm) (ppm) (ppm) (%)
(%) (ppm) (ppm) (ppm) (%)
(%)
68
1984
65
2.19 19.04
14
4857
18
1.25 18.99
72
1776
87
2.47 18.41
29
4215
25
1.58 17.45
38
3259
29
1.97 17.69
79
1682
105 2.38 17.24
45
2940
38
2.01 16.47
89
1402
119 3.55 17.09
97
1198
137 4.12 15.78
49
2581
59
2.29 16.30
Exhaust Data Comparison for Diesel and Dual-Fuel under Moderate
Load Operating Conditions
96
APPENDIX E
SAMPLE ANALYSIS
CALCULATIONS
97
Calculations for engine performance with engine running at maximum
operating conditions:
A.
FOR DIESEL FUEL
At 1600 rpm,
Time for 10 ml of diesel = 78s
Volume flow rate
=
10 × 10 −3 × 10 −3 m 3
78s
= 1.282 × 10 −7 m 3 / s
Density of diesel
= 880 kg/m3
Mass flow rate
= 1.282 × 10 −7 m 3 / s x 880 kg/m3
= 1.128 × 10 −4 kg / s
a) Input Power
IP = m f × QHV × 10 3
where
IP
=
input power
(kW)
mf
=
mass flow rate of fuel
Q HV
=
lower calorific value of fuel
(kg/s)
(MJ/kg)
For diesel fuel, Q HV , Diesel = 42.5 MJ/kg
For CNG fuel, Q HV ,CNG
Input Power
= 45 MJ/kg
= 1.128 × 10 −4 × 42.5 × 10 3
= 4.794 kW
98
b) Torque
The torque is directly measured from the electronic control board mounted to
the dynamometer.
Torque , τ = 7.0 Nm
c) Brake Power
2πNτ
60
Brake Power, BP =
where
τ
=
torque
BP
=
power developed by engine
N
=
engine speed
BP
(N. m)
=
(W)
(rpm)
2π × 1600 × 7.0
60
= 1.173 kW
d) Specific Fuel Consumption
sfc =
mf
P
× 3.6 × 10 6
where
sfc
=
specific fuel consumption
mf
=
mass flow rate of fuel
P
=
power output
(g/kW. hr)
(kg/s)
(kW)
99
sfc
=
1.128 × 10 −4
× 3.6 × 10 6
1.173
= 346.2 g/ kW.hr
e) Brake Mean Effective Pressure
bmep =
BP × n R × 60
A× L × N × n
where
bmep =
brake mean effective pressure (kPa)
BP
=
brake power
nR
=
number of crank revolutions for each power stroke per
(kW)
cylinder
(m2)
A
=
area of engine bore
L
=
length of engine stroke (m)
N
=
engine speed
bmep =
π
4
(rpm)
1.173 × 2 × 60
× (0.07 ) × 0.055 × 1600 × 1
2
(for single cylinder)
= 415.63 kPa
f) Overall Efficiency
Efficiency =
=
BP
IP
1.173
4.794
= 0.245
= 24.5%
100
B.
FOR DUAL FUEL
At 1600 rpm,
Time for 10 ml of diesel pilot fuel = 290 s
Volume flow rate of diesel
=
10 × 10 −3 × 10 −3 m 3
290 s
= 3.448 × 10 −8 m 3 / s
Density of diesel
= 880 kg/m3
Mass flow rate of diesel
= 3.448 × 10 −8 m 3 / s x 880 kg/m3
= 3.034 × 10 −5 kg / s
Diesel ratio
=
Diesel mass flow rate in dual fuel
Diesel mass flow rate in pure diesel
=
3.034 × 10 −5 kg / s
1.128 × 10 − 4 kg / s
= 0.269
CNG Ratio
= 1 - 0.269
= 0.731
a)
Input Power
The input power is taken to be the same for both diesel fuel and dual fuel since
the engine is operating under the same speed and load conditions.
∴ Input power = 4.794 kW
From the input power, the mass flow rate of CNG can be calculated:
101
[(Diesel ratio × m
f , Diesel
[(0.269 × 3.034 × 10
]
× QHV , Diesel ) + (CNG Ratio × m f ,CNG × QHV ,CNG ) × 10 3 =4.794
]
)
× 42.5 + (0.731 × m f ,CNG × 45) × 10 3 = 4.794
−5
m f ,CNG
b)
= 1.352 × 10 −4 kg / s
Torque
The torque is directly measured from the electronic control board
mounted to the dynamometer.
Torque, τ
c)
= 8.0 Nm
Brake Power
BP
=
2πNτ
60
=
2π × 1600 × 8.0
60
= 1.340 kW
d)
Specific Fuel Consumption
sfc
=
sfc
=
mf
P
× 3.6 × 106
0.269 × 3.034 × 10 −5
0.731 × 1.352 × 10 −4
+
1.340
1.340
× 3.6 × 10 6
= 287.44 g/kW.hr
102
e)
Brake Mean Effective Pressure
bmep =
bmep =
BP × nR × 60
A× L× N × n
π
4
1.340 × 2 × 60
× (0.07 ) × 0.055 × 1600 × 1
2
(for single cylinder)
= 474.81 kPa
f)
Overall Efficiency
Efficiency =
=
BP
IP
1.340
4.794
= 0.2795
= 27.95%
103
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