DEVELOPMENT OF A COMPRESSED NATURAL GAS (CNG) MIXER FOR A
TWO STROKE INTERNAL COMBUSTION ENGINE
DEVARAJAN A/L RAMASAMY
UNIVERSITI TEKNOLOGI MALAYSIA
PSZ 19:16 (Pind.1/97)
Universiti Teknologi Malaysia
BORANG PENGESAHAN STATUS TESIS
JUDUL: DEVELOPMENT OF A COMPRESSED NATURAL GAS (CNG) MIXER
FOR A TWO STROKE INTERNAL COMBUSTION ENGINE
SESI PENGAJIAN: 2004/2005-2
Saya DEVARAJAN A/L RAMASAMY
(HURUF BESAR)
mengaku membenarkan tesis (PSM/Sarjana/Doktor Falsafah)* ini disimpan di PerpustakaanUniversiti Teknologi Malaysia dengan syarat-syarat kegunaan seperti berikut:
1. Tesis adalah hakmilik Universiti Teknologi Malaysia.2. Perpustakaan Universiti Teknologi Malaysia dibenarkan membuat salinan untuk tujuan
pengajian sahaja.3. Perpustakaan dibenarkan membuat salinan tesis ini sebagai bahan pertukaran antara
institusi pengajian tinggi.4. **Sila tandakan ( )
SULIT(Mengandungi maklumat yang berdarjah keselamatan ataukepentingan Malaysia seperti yang termaktub di dalamAKTA RAHSIA RASMI 1975)
TERHAD (Mengandungi maklumat TERHAD yang telah ditentukanoleh organisasi/badan di mana penyelidikan dijalankan)
TIDAK TERHAD
Disahkan oleh
(TANDATANGAN PENULIS) (TANDATANGAN PENYELIA)
Alamat Tetap: No. 70, Jalan Batu Nilam 9 PM DR. ROSLI ABU BAKARKaw. Perumahan Bukit Tinggi Nama Penyelia42000, Klang, Selangor
Tarikh: Tarikh:
CATATAN: * Potong yang tidak berkenaan.** Jika tesis ini SULIT atau TERHAD, sila lampirkan surat daripada pihak berkuasa
/organisasi berkenaan dengan menyatakan sekali sebab dan tempoh tesis ini perludikelaskan sebagai SULIT atau TERHAD.Tesis dimaksudkan sebagai tesis bagi Ijazah Doktor Falsafah dan Sarjana secara penyelidikan, atau disertasi bagi pengajian secara kerja kursus dan penyelidikan,atau Laporan Projek Sarjana Muda (PSM).
I hereby declare that I have read this thesis and in my
opinion this thesis is sufficient in terms of scope and quality for the
award of the degree of Master of Engineering (Mechanical)
Signature :
Name of Supervisor : PM. DR. ROSLI ABU BAKAR
Date :
PART A – Confirmation Of Cooperation*
It was confirmed that this thesis research project was accomplished with the cooperation
between __________________ and _________________.
Confirmed by:
Signature : ____________________________ Date: _______________
Name : ____________________________
Position : ____________________________
(Official Stamp)
* If the thesis/project research involves cooperati on.
PART B – For The School Of Graduate Studies Office Usage
This thesis had been checked and approved by:
Name and Address of External Examiner : __________________________________
__________________________________
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Name and Address of Internal Examiner : __________________________________
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Name of Other Supervisor (If Available) : __________________________________
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Confirmed by Assistant Registrar of SPS:
Signature : ____________________________ Date: _______________
Name : ____________________________
UTM(PS)-1/02
Sekolah Pengajian Siswazah
Universiti Teknologi Malaysia
PENGESAHAN PENYEDIAAN SALINAN E-THESIS
Judul tesis : DEVELOPMENT OF A COMPRESSED NATURAL GAS (CNG) MIXER FOR A TWO STROKE INTERNAL COMBUSTION ENGINE
Ijazah : SARJANA KEJURUTERAAN MEKANIKAL (AUTOMOTIF)
Fakulti : FAKULTI KEJURUTERAAN MEKANIKAL
Sesi Pengajian : 2004/2005-2
Saya DEVARAJAN A/L RAMASAMY mengaku telah menyediakan salinan e-thesis sama seperti
tesis asal yang telah diluluskan oleh panel pemeriksa dan mengikut panduan penyedian Tesis dan
Disertasi Elektronik (TDE), Sekolah Pengajian Siswazah, Universiti Teknologi Malaysia, Januari
2005.
___________________________
(Tandatangan pelajar)
_____________________________
(Tandatangan penyelia sebagai saksi)
Alamat tetap:
N0 70, Jalan Batu Nilam 9,
Bukit Tinggi,
41200 Klang,
Selangor
Nama penyelia: PM DR. ROSLI ABU BAKAR
Fakulti: FAKULTI KEJURUTERAAN MEKANIKAL
Nota: Borang ini yang telah dilengkapi hendaklah dikemukakan kepada SPS bersama penyerahan CD.
DEVELOPMENT OF A COMPRESSED NATURAL GAS (CNG) MIXER FOR A
TWO STROKE INTERNAL COMBUSTION ENGINE
DEVARAJAN A/L RAMASAMY
A thesis submitted in fulfilment of the
requirements for the award of the degree of
Masters of Engineering (Mechanical)
Fakulti Kejuruteraan Mekanikal
Universiti Teknologi Malaysia
OCTOBER 2005
ii
I declare that this thesis entitled “Development of a Compressed Natural Gas (CNG)
Mixer for a Two Stroke Internal Combustion Engine” is the result of my own research
except as cited in the references. The thesis has not been accepted for any degree and
is not concurrently submitted in candidature of any other degree.
Signature: ....................................................
Name : ....................................................Devarajan A/L Ramasamy
Date : ....................................................
iii
This work is dedicated to my beloved ones,
My Father Mr Ramasamy Rengasamy
My Mother Mrs. Ganam Govindan
My Brother Mr.Saravanan Ramasamy
My Sisters Ms Kalaivani Ramasamy and Ms Karthikayeneee Ramasamy
And
My sweetheart Ms Gayadri Krishanan
iv
ACKNOWLEDGEMENT
I would like to express my greatest appreciation and gratitude, to my
supervisor, Dr. Rosli Abu Bakar for his gu idance, patience for giving advises and
supports throughout the progress of this study. Throughout this work, I have learned
much from him. Special thanks are also given to my co-supervisor, Dr Normah
Ghazali for the guidance to comment my thesis.
Not forgotten, special thanks to En Mardani and all the technicians at the
Automotive Laboratory of Universiti Teknologi Malaysia, especially Wak Sairaji, En.
Mazlin, En. Subki, En. Hishamudin and Tuan Haji Wahab. They were not hesitant to
answer all my doubts and spending their time to guide me during my experimental
work.
A great appreciation is acknowledged to the Ministry of Science, Technology
and Innovation for the funding under the Intensified Research in Prioritized Area
(IRPA), vot 74516. The project is hoped to succeed under the guidance and
knowledge of Prof. Ir. Dr Azhar Abdul Aziz.
Last but not least, I would like to thank all of my friends and teammates
especially Mr. Wong Hong Mun, Mr. Gan Leong Ming, Mr. Chong Chin Lee and
Mr. Fadzil Rahim, for their support and encouragement given to me, especially
during the hard times.
v
ABSTRACT
Compressed Natural Gas (CNG) has been accepted widely as an alternative to
gasoline. More importantly the use of CNG in two stroke engines will drastically
reduce the high emission output from these engines as these engines are widely used
around the world. A conversion kit is used to apply the fuel in engines. A bi-fuel
conversion system converts engines without much modification to other systems.
They are normally produced for four stroke application. This kit has to be studied to
be modified for two stroke application. The part that connects the engine to the kit is
called a gaseous fuel mixer. This part mixes the air and fuel due to its venturi shape.
A mixer provides fuel suction at different engine speeds due to pressure difference at
the throat. The optimisation of the throat is important as a small throat will cause
poor performance at high speeds while a large throat will reduce fuel suction. The
smaller throat size creates higher velocity and lower pressure. This low pressure
creates fuel suction into the mixer. The mixer was designed for a two stroke engine
air flow. Computer aided design (CAD) and computational fluid dynamic (CFD)
software were used as a tool for the design. The design is optimised for inlet and
outlet angles, number and size of the hole at the throat circumference and also the
throat size. The prototype design was manufactured based on optimised dimensions
of the mixer that were obtained from CFD analysis. The mixer was validated to show
that the CFD analysis was correct. Testing apparatus were used to do the validation.
The apparatus consists of a laminar flow element (LFE), a smoke generator, a digital
manometer and a gaseous flow meter. It was used to validate the flow pattern,
pressure drop from the mixer and the air fuel ratio given by the mixer.
vi
ABSTRAK
Gas Asli Termampat (CNG) telah diperaku i sebagai satu alternatif kepada
petrol. Penggunaan gas in dalam enjin dua lejang mampu mengurangkan pengeluaran
pencemaran tinggi dari enjin ini. Ini kerana penggunaan enjin dua lejang adalah
banyak di dunia. Bahan api ini digunakan pada engine melalui kit penukaran.
Penukaran enjin petrol ke CNG perlu dilakukan dengan modifikasi kecil pada enjin
asal. Oleh itu, kit penukar CNG dwi-bahanapi digunakan. Unit ini dibuat lazimnya
untuk enjin empat lejang, oleh itu, ia perl u dikaji bagi penggunaan dalam enjin dua
lejang. Bahagian pada alat ini yang bersambung kepada enjin dinamakan sebagai
pencampur bahanapi bergas. Ia menyebabkan gas bercampur pada bahagian yang
berbentuk venturi. Pencampur ini memberikan sedutan gas kepada enjin pada halaju
enjin yang berbeza disebabkan perbezaan tekanan pada bahagian yang dipanggil
leher. Ubahsuai leher adalah penting bagi operasi alat ini. Ubahsuai leher adalah
perlu kerana leher yang kecil akan menyebabkan prestasi enjin yang rendah pada
kelajuan tinggi manakala leher yang besar tidak dapat memberi sedutan gas yang
diperlukan. Tekanan rendah menyebabkan sedutan pada pencampur ini. Pencampur
direkabentuk untuk aliran udara pada enjin dua lejang. Rekabentuk berbantukan
computer (CAD) dan Dinamik Bendalir berbantukan computer (CFD) digunakan
sebagai alat rekabentuk. Rekabentuk pencampur diubahsuai dengan menggunakan
CFD pada sudut masukan dan keluaran, bila ngan lubang dan saiz lubang pada leher
serta saiz leher itu sendiri. Prototaip dibuat berdasarkan dimensi pencampur yang
diperolehi daripada analisis CFD. Untuk membuktikan analisis CFD pengesahan
telah dilakukan. Peralatan ujikaji telah digunakan untuk melakukan pengesahan ini.
Ia terdiri daripada elemen aliran laminar (LFE), penghasil asap, manometer digital
dan meter aliran gas. Peralatan ini digunakan bagi tujuan pengesahan bentuk aliran,
kejatuhan tekanan dan nisbah udara kepada bahan api yang diberi oleh pencampur
ini.
vii
CONTENTS
CHAPTER TITLE PAGE
TITLE i
DECLARATION ii
DEDICATION iii
ACKNOWLEDGEMENT iv
ABSTRACT v
ABSTRAK vi
CONTENTS vii
LIST OF TABLES xi
LIST OF FIGURES xii
LIST OF APPENDICES xiv
LIST OF SYMBOLS xv
1 INTRODUCTION 1
1.1 Problem Statement 2
1.2 Objectives 3
1.3 Scope 3
1.4 Methodology 3
viii
2 LITERATURE REVIEW 5
2.1 Two Stroke Engine 5
2.2 CNG as Fuel for Two Stroke Engines 6
2.2.1 CNG as an Alternative Fuel 7
2.2.2 Combustion Characteristics of CNG 10
2.2.3 Emission Reduction from CNG Usage in
Two Stroke Engines 11
2.2.4 Other Issues Regarding CNG Usage 13
2.3 CNG Mixer 14
2.3.1 Current Trends in CNG Mixer Design 15
2.3.2 Sizing of the Mixer Throat 18
2.3.3 Pressure Drop in the Mixer 19
2.3.4 CNG Mixer and Engine Conversion Kits 23
2.4 Summary of Literature Review 24
3 DESIGN OF A VENTURI BURNER MIXER 25
3.1 Conceptual Design 26
3.2 Procedure of Mixer Design 28
3.2.1 Initial Throat Size 29
3.2.2 CFD Simulations of the Mixer 30
3.2.3 Inlet and Outlet Angles of the Mixer 34
3.2.4 Number of Holes at Throat Circumference 36
3.2.5 Size of Hole at Throat Circumference 37
3.2.6 Throat Size Optimisation 37
3.3 Prototyping the Mixer 38
3.4 Validating the Mixer Design 39
3.4.1 Testing Apparatus 39
3.4.2 Testing Procedure 42
3.4.2.1 Smoke Mixing in Mixer 43
3.4.2.2 AF Ratio Test 43
ix
3.4.2.3 Pressure Drop Test 46
4 RESULT AND DISCUSSION 47
4.1 Designing of the Mixer 47
4.1.1 Initial Throat Size 47
4.1.2 CFD Simulation of the Mixer 48
4.1.3 Inlet and Outlet Angles of the Mixer 48
4.1.4 Number of Holes at Throat Circumference 52
4.1.5 Size of Hole at Throat Circumference 56
4.1.6 Throat Size Optimisation 58
4.2 Prototyping the Mixer 63
4.3 Validating the Mixer Design 65
4.3.1 Smoke Mixing in Perspex Prototype 65
4.3.2 AF ratio Testing of Mixer 67
4.3.3 Pressure Drop Testing of Mixer 69
5 CONCLUSION AND RECOMMENDATION 72
5.1 Conclusion 72
5.3 Recommendation 73
REFERENCES 74
APPENDICES 77
Appendix A 77
Appendix B 79
x
Appendix C 109
Appendix D 117
Appendix E 125
Appendix F 128
xi
LIST OF TABLES
TABLE NO. TITLE PAGE
2.1 Energy content of alternative fuels relative to petrol
and diesel 8
2.2 Proven natural gas reserves 8
2.3 Average natural gas composition in Malaysia 9
2.4 Methane gas properties 10
2.5 Typical 2-stroke emissions 12
2.6 Current regulation that is available for two-stroke
engines 12
2.7 Fuel price 13
3.1 Specification of the analysed engine 29
3.2 Properties of air 33
5.1 Specification of the mixer designed 73
xii
LIST OF FIGURES
FIGURE NO. TITLE PAGES
1.1 Methodology 4
2.1 Operation of a two stroke engine 6
2.2 Type of CNG mixers currently being used in the market 15
2.3 Power test results for different mixer designs 16
2.4 Venturi upstream of the carburettor 18
2.5 Mixer after throttle in intake system of injection engine. 18
2.6 Schematic plot of velocity and pressure across a venturi 20
2.7 Pressure profile during intake stroke of an engine 21
2.8 Pressure drop in air cleaner and intake manifold 22
3.1 Methodology for designing the CNG mixer 25
3.2 The concept models 27
3.3 Proposed shape of the mixer 28
3.4 Location of throat diameter 30
3.5 Simulation steps for each simulation 32
3.6 Overall simulation stages done on the mixer 34
3.7 Simulation model for inlet and outlet angles 35
3.8 Schematic diagram of flow test rig to measure air flow 40
3.9 Schematic of smoke generator connected to test rig 41
3.10 Schematic diagram of pressure measurement 42
4.1 Pressure plot along the centre line of the mixer at different
inlet and outlet angles 49
4.2 Lowest pressure at the throat diffuser wall 50
4.3 Pressure ratios of each model inlet and outlet angle changes 51
4.4 Eight holes mixer model 53
4.5 Ten holes mixer model 54
xiii
4.6 Twelve holes mixer model 55
4.7 Effect of AF ratio on hole sizes at throat circumference
at all speed range 57
4.8 Effect of throat diameter size on air fuel ratio 60
4.9 Simulation pressure drop due to different throat size
at all engine speed 62
4.10 Perspex model for flow testing 63
4.11 Assembled view of Aluminium mixer 64
4.12 Components of Aluminium mixer 64
4.13 Simulation of smoke at 1000 rpm, 2000 rpm and 3000 rpm
air speed 66
4.14 Experiment and simulation results of AF ratio 68
4.15 Simulations and experiment pressure drop 71
xiv
LIST OF APPENDICES
APPENDIX TITLE PAGES
A Thesis Gantt Chart 77
B CFD Analysis 79
C Apparatus and Experiments 109
D Technical Drawings 117
E Material Selection 125
F Mesh Independant Analysis 128
xv
LIST OF SYMBOLS
AF Air fuel ratio -
1A Area in inlet m2
2A Area at throat m2
C Viscosity constant -
Cv Specific Heat J/kgK
Dr Delivery ratio -
HL Losses in pipe Pa
k Turbulent kinetic energy J/kg
1m Inlet mass flow rate kg/s
N Engine speed rpm
aQ Volumetric air flow rate m3/s
1Q Measured flow rate m3/s
2Q Actual flow rate m3/s
atmp Atmospheric pressure Pa
QH Heat source per unit volum e J/m 3
qi Diffusive heat flux J/s
Si Mass-d istributed external force per unit mass N/kg
U Fluid velocity m/s
1v Velocity at inlet m/s
2v Velocity at throat m/s
p Pressure drop Pa
airP Pressure drop in the air cleaner Pa
uP Intake pressure drop upstream Pa
thrP Pressure drop across throttle Pa
xvi
valveP Pressure drop across intake valve Pa
1 Air density at inlet kg/m 3
f Turbulent viscosity factor. -
ij Kronecker delta function -
Turbulent dissipation J/s
Angle º
ik Viscous shear stress tensor Pa
Dynamic viscosity kg/m s
l Dynamic viscosity kg/m s
t Turbulent eddy viscosity kg/m s
CHAPTER 1
INTRODUCTION
Current trends in the automotive industry are ever changing especially
regarding the usage of alternative fuels. The search for the best alternative fuel that
produces the least amount of emission has sparked concerns to many researchers.
Maxwell (1995) stated that many studies on alternative fuel have been carried out
and researchers are looking at natural gas, liquefied petroleum gas (LPG), methanol,
ethanol, and hydrogen. All of these fuels have their advantages and disadvantages
which are cost, availability, environmental impact, usage in vehicle, safety and the
acceptance by consumers.
Current fuel price inflation and also current oil crisis, drastic moves were
taken by many countries to reduce petroleum usage and finding other alternatives to
its usage. In developing countries, the concern of finding alternative fuels has
started and already had become an issue. With gas reserves three times more than
petroleum oil, Malaysia is increasingly turning its attention towards natural gas. The
national petroleum company of Malaysia, PETRONAS has embarked on the
Natural Gas for Vehicles (NGV) program where NGV dispensing facilities are
available at some selected PETRONAS service stations, located in high traffic
density areas of Kuala Lumpur and Johor Bahru. The government support for the
NGV program was seen in 25% reduction on car road tax for using NGV as well as
requiring new taxis in the Klang Valley to use CNG by engine conversion systems.
2
In automotive applications, natural gas can be used in three forms based on
how the natural gas is stored. One of the most popular forms of natural gas is the
compressed natural gas (CNG), which is natural gas in pressurised form. The other
least popular methods of obtaining natural are liquefied natural gas and the
absorption natural gas.
CNG is a good alternative to petrol and diesel. Consumers would easily
accept this form of alternative as it has low operational cost due to subsidised price
and its usage could provide cleaner engine emissions. The main reason behind CNG
fuel being cleaner is that natural gas is principally comprise of 90% methane, which
is the simplest form of hydrocarbon. Even so, the CNG fuel available today still
lack in some qualities compared to petroleum fuel. For example, CNG fuelled
engines normally possess lower engine performance compared to petrol.
The main reason is that CNG fuelling systems creates a lot of losses in
terms of volumetric efficiency. This happens as CNG must be supplied to the
engine through a mixing device before the mixture of CNG and air is drawn into the
engine. This causes less fuel in the combustion chamber and reduces volumetric
efficiency. Currently petrol fuelled engine are converted into a CNG fuelled engine
by means of a fuel mixing device.
1.1 Problem Statement
Currently, there are no specific CNG mixers specifically designed for two
stroke engines in the market. All of the conversion kits that are available for four
stroke engines only. A proper CNG mixer should be designed for two stroke engine
application. A supercharged 150 cc two stroke engine has been chosen for CNG
conversion. Direct usage of a conventional four stroke engine CNG mixer for two
stroke engines is not possible as they are too large a size for a small two stroke
engine air requirements. The design of the mixer has to consider the whole range of
engine operating condition in order to provide a complete view of its performance.
3
The existing four stroke engine CNG mixers are usually not properly refined
and optimised to enable good air fuel mixing. In addition, the efficiency of the
current mixer design is also an issue as it is designed for simplicity which only
offers practicality but lack in efficient air flow performance throughout the engine
speed. Therefore, a straight forward conversion is not possible.
1.2 Objectives
The objectives of the study are as follows:
1) To design a venturi burner type CNG mixer for a two stroke engine
according to the engine’s air requirement using CFD.
2) To fabricate the optimised prototype of the CNG mixer and test it on a
flow bench machine.
1.3 Scope
The scopes of the research are as follows:
1) Preliminary design of the CNG mixer.
2) Optimising the CNG mixer design using CFD as a design tool .
3) Fabrication of the prototype CNG mixer.
4) Testing and validation of the CNG mixer design.
1.4 Methodology
A general methodology was followed in the research as indicated in the flow
chart as shown in Figure 1.1:
4
Start
Literature review
Concept design
Designing of mixer
Figure 1.1 Methodology
Prototyping the mixer
Meet design criteria
No
Yes
Validating the mixer design
End
CHAPTER 2
LITERATURE REVIEW
2.1 Two Stroke Engine
Air pollution in many Asian cities is increasing due to the proliferation of
vehicles powered by simple two-stroke cycle engines.In Asia, there are an estimated
70-100 million two stroke engines operating which are motorcycle, tricycle and auto
rickshaws to name a few.
In a two stroke engine, there is a power stroke in every revolution whereas
there is only one power stroke for every two revolutions in four stroke engines. With
this, it can be said the two stroke engine can produce power better than the four
stroke engine (Bryan, 2002). The higher power production allows the two stroke
engine to have a higher power to weight ratio and are simpler in design. Evolutions
of the two stroke engines have seen many changes in their designs. There are designs
of two stroke engines with single or multi cylinders, with turbochargers or even with
superchargers.
The basic operation of a two stroke engine has not differed much from how
the first two stroke engine was designed. Figure 2.1 describes an operation of a two
stroke engine with a supercharger system.
6
Compress
Air drawn intothe engine driven supercharger to be
compressedIgnition
Piston descends due to Compressed air sent
to the aircompartment
combustion
Figure 2.1 Operation of a two stroke engine
As seen in Figure 2.1, the intake of a two stroke engine is based on the
scavenging process. Scavenging occurs when the intake ports are uncovered, the
compressed air and fuel pushes in and displaces the remaining exhaust gases. There
are many types of scavenging process that is used in a two stroke engine among them
are cross, loop or uniflow scavenging (Ferguson, 2001).
2.2 CNG as Fuel for Two Stroke Engines
CNG is a type of alternative fuel to petrol and is vastly accepted around the
world. With current fuel price inflation and also current oil crisis that is going on,
drastic moves have been taken by many countries to reduce petrol usage and find
other alternatives to its usage.
Compressed air fuel mixtureforced into
cylinder through intake port
Exhaust escapes through exhaust
valves
Fuel is mixed with compressed air
in air compartment
Scavenging
7
2.2.1 CNG as an Alternative Fuel
By definition, alternative fuels are fuels that can be derived from non-crude
oil resources. Crude oils are petroleum based fuels. Some of the types of alternative
fuels that are available today which are not from crude oil are Natural Gas (NG),
Liquefied Petroleum Gas (LPG), Methanol, Ethanol, Hydrogen and others (Maxwell,
1995).
Amount energy released by burning of fuel in air is dependant on the type of
fuel. This was seen in Table 2.1, which shows in lower heating values for some
known alternative fuels. Judging by their energy density, ethanol and methanol must
be burned more to produce energy equal to that of petrol. Other fuels such as NG,
LPG and hydrogen also have a lower density compared to petrol. With this they can
provide more energy per kilogram equivalent of petrol. Vehicles running in any of
these fuels are more efficient as they produce more energy for the same given mass
of petrol. Here engine design plays an important role in providing the efficiency to
the engine to make full use of the higher energy content derived from these fuels.
Due to its high energy content and large availability, NG has been chosen to
be studied further worldwide. Table 2.2 shows the distribution and availability of NG
throughout the world. NG has already been used in more than 1 million vehicles in
the world since 1993 and the number is increasing as more consumers are exposed to
the numerous benefits of NG. It has been used for domestic and industrial sectors
(heating, thermal energy production, chemical industries). NG has also been labelled
as a clean fuel in ecological considerations (Poulton, 1994).
In Malaysia, the national oil and gas company PETRONAS had introduced
the NGV program in 1986. Now there are almost 4,000 taxis called the Enviro 2000
vehicles in Kuala Lumpur since 1998. It was expected that the Malaysian car
manufacturer, PROTON would be producing around 40,000 NGV’s by 2004 for the
Malaysian market (Taib Iskandar Mohamad, 2003).
8
Table 2.1 Energy content of alternative fuels relative to petrol and diesel (Maxwell,
1995)
Fuel Density
(kg/m3)
Energy Content
(MJ/m3)
Energy Relative
to equivalent
mass of Petrol
Energy Relative
to equivalent
mass of Diesel
Petrol 621.8 4257 100% 91%
Diesel 622.2 4694 110% 100%
LPG 422.1 3113 115% 109%
Methanol 658.5 2100 49% 45%
Ethanol 652.5 2813 66% 60%
NG 351.2 2814 120% 113%
Table 2.2 Proven natural gas reserves, 1991, (Poulton, 1994)
Area
Trillion
Cubic
Meters
Billion
Tonnes Oil
Equivalent
Share of Total
(%)
North America 7.5 6.7 6.1
Latin America 6.8 6.1 5.4
Western Europe 5.1 4.6 4.1
CIS/E Europe 50.0 45.0 40.4
Middle East 37.4 33.7 30.1
Africa 8.8 7.9 7.1
Asia/Australasia 8.4 7.6 6.8
Total 124.0 111.6 100.0
The growth and development of NGV industry in Malaysia was slow due to
the lack of refuelling stations and unavailability of original equipment manufacturer
(OEM) of gas conversion kits. On the contrary, the government is supportive as they
had given import duty and sales tax exemption on the conversion kits. Tax reductions
of 25% for bi-fuel vehicles and 50% on monogas vehicles were given to boom the
industry by the government.
9
Malaysia is currently using EURO II emission regulation and will continue
the standard until 2010 (Jitendra Shah, 2001). The exhaust emission by NGV
vehicles are well below EURO II limits on carbon monoxide, hydrocarbon and
nitrogen oxide. In addition, these vehicles can travel up to 170 km per filling of NGV
rendering its performance equivalent to a petrol-powered vehicle. More importantly,
these NGV vehicles can generate a significant saving on fuel expense as natural gas
is cheaper compared to petrol (Yusoff Ali, 2003).
The composition of CNG also varies between countries. The principle
ingredient of CNG is methane. Methane makes up to 90 percent of CNG. Aside of
methane, the composition contains small portions of other gases such as ethane,
propane, butane, pentane and hexane. It can also contain nitrogen, helium, carbon
dioxide and hydrogen sulphide. Malaysia owns the 12th largest NG reserve in the
world according to Gas Malaysia (2003). The average NG composition in Malaysia
is shown in Table 2.3, while the properties of methane are shown in Table 2.4.
Table 2.3 Average natural gas composition in Malaysia (Gas Malaysia, 2003)
Natural Gas Composition Percentage
Methane 92.73
Ethane 4.07
Propane 0.77
n-Butane 0.06
i-Butane 0.08
Other Hydrocarbon 0.01
Nitrogen 0.45
Carbon Dioxide 1.83
10
Table 2.4 Methane gas properties (Gas Malaysia, 2003)
Methane characteristics
Density (kg/m3) 0.715
Gross Calorific Value (kCal/cm3) 9530
Burning Velocity (m/s) 0.3
Upper Flammability Limit 15.4
Lower Flammability Limit 4.5
Auto Ignition Temperature (oc) 640
Theoretical Air Requirement (m3) 9.74
2.2.2 Combustion Characteristics of CNG
Fuel energy can be harvested only through combustion. Any fuel that
combusts completely will produce simple by-products. A complete burning of fuel in
air to produce this condition is called a stoichiometric condition. When a
stoichiometric air fuel mixture combust, it will produce energy from combustion,
water vapour, carbon dioxide while other composition such as nitrogen and inert
gases will remain constant. Theoretically, combustion burns the fuel completely
without much emission such as carbon monoxide, hydrocarbon from excess fuel and
nitrogen oxides from reaction with surrounding air.
Stoichiometric air fuel (AF) ratio was calculated by first knowing the
combustion chemical reaction of the fuel in air. AF ratio is defined as the mass of air
to the mass of fuel in a mixture. Knowing CNG consists of 90% methane (CH4), the
chemical reaction of CNG fuel is therefore given by:
OHCOOCH Energy2224 22 (2.1)
The equation states that 1 mole of methane would completely combust with 2 moles
of oxygen to produce energy as well as 1 mole carbon dioxide and 2 moles of water.
11
Air is a composition of many gases, where 20.95% of the whole composition
is oxygen and the rest are nitrogen and other gases. Since the mass of 1 mole of air is
28.96 g, therefore 1 mole of oxygen is contained in a mass of 138.23 g of air (Lenz,
1992). On the other hand, the mass of 1 mole of CNG or methane from equation 2.1,
is given as (1 X 12.01 + 4 X 1.008) or 16.042 g. By knowing the mass of each mole
of air and CNG, the stoichiometric AF ratio was found as stated below.
(2.2)
23.17042.16
46.276
042.16
23.1382
ratioAFfmam
According to Yeap (2002), for idling at normal engine operating temperature,
the engine still demands a fairly rich mixture, for which the AF ratio usually in the
range of 11:1 to 13:1. A good mixer must be able to meet these requirements. The AF
ratio becomes too rich and incombustible for the engine at 9.77:1. At too rich
conditions the engine will have less oxygen to burn with the fuel, causing a rise in
hydrocarbon emission or in the worst case it will stall the engine.
2.2.3 Emission Reduction from CNG Usage in Two Stroke Engines
The main problem that occurs in a two stroke engine is they are characterised
by very high levels of hydrocarbon (HC), carbon monoxide (CO), and particulate
matter (PM) emissions. Table 2.5 shows the current emissions of two-stroke engines.
Clearly from Table 2.5, it can be seen that the usage of CNG in two stroke engines
can drastically reduce the emissions produced by normal petrol fuelled engines.
12
Table 2.5 Typical 2-stroke emissions (Bryan, 2002)
Type Of Engine CO
g/kW-h
HC
g/kW-h
NOx
g/kW-h
55 kW 2-stroke
outboard engine 185 134 2.74
7.4 kW 2-stroke
Johnson OMC-J10RCSE
Engine
519 236 0.73
Cooper GMVC-10C 2-stroke
natural gas engine1.88 9.1 0.67
Comparisons of emission are generally based on standards followed by
regulating bodies of a country. All vehicles should follow these regulations in term
of emission. The guideline for two stroke engine emission in India and Unites Stated
of America is shown in Table 2.6. The overall emission of the two stroke engine
using CNG is lower than standard emission regulation of two stroke engines. It was
only slightly higher in term of HC emission of a small engine regulation standard in
California. The California emission standard is considered one of the highest
standards in emission regulations in the world.
Table 2.6 Current emission standard that is available for small engines by
experiments (Bryan, 2002)
CO
g/kW-h
HC
g/kW-h
NOx
g/kW-h
Indian 2-stroke
Genset regulation 603 603 -
US EPA 2-stroke
emission factor, 1999 522 206 0.67
California 4-stroke emission
factor < 19 kW (25 hp) 322 5.4 2.4
Euro II Standard * 4 7 0.15
* Euro II standard is the emission standard followed by Malaysia
13
2.2.4 Other Issues Regarding CNG Usage
Apart from having lower emissions than petrol, the CNG fuel also has its
advantages and disadvantages in other areas. Advantages of using CNG fuel in
vehicles include (Rosli, 2002):
a. Higher Octane number in the range of 120 to 130, which is
considerably higher than 93 to 99 Octane for petrol. A high Octane
number ensures that CNG fuel can run at high compression ratio
without any knocking phenomena to the piston that will cause damage
to the engine.
b. Higher flammability compared to petrol that makes it appropriate to run
on lean burn technology.
c. Safer; as it is lighter and dissipates quickly. Due to this it ignites
quickly, but only when the fuel to air ratio was between 5 – 15% by
volume.
d. Because it is a clean burning fuel, it reduces the required maintenance
cost of vehicle; it can be half of petrol—oil changes can be done for
more than 15,000-30,000 km, spark plug points can be changed at
intervals up to 120,000 km.
e. Plenty of reserve; there is an estimated 65-70 year supply of natural gas.
Besides made from fossil fuel, natural gas can also be made from
agricultural waste, human waste and garbage.
f. Cheaper per litre equivalent than petrol, in Europe 14-17% less than
petrol and 12-74% less expensive than diesel. In Malaysia, the CNG
price is half less compared to petrol as shown in Table 2.7
Table 2.7 Fuel price
FuelPump Price/Litre 1)
CNG RM 0.56Petrol (Unleaded) RON 97 RM 1.41Petrol (Unleaded) RON 92 RM 1.36Diesel RM 0.83
Note: 1) Source: Petronas pump price (2005)
14
However, CNG fuel has some disadvantages that limit its potential to achieve
the optimum engine performance, are as stated below:
a. Since CNG is available in gaseous form, it has a low density. CNG
from the mixer drawn into the engine cylinder displaces approximately
8% to 10% of Oxygen by volume. This reduces the amount of Oxygen
due to larger space occupied by the CNG in the combustion chamber.
b. CNG has a low flame speed. Its burns slower than conventional fuels,
such as petrol and diesel. As much as 60% decrease in burning velocity
has been measured. This prolongs the total combustion duration
compared with diesel and petrol. This can cause a further reduction in
the engine output of 5 to 10%.
Even though CNG has disadvantages, the advantages outweigh the
disadvantages.
2.3 CNG Mixer
The principle characteristic of a CNG mixer was known to analyse the
operation of the mixer. The operation of a mixer is that the change in velocity
causes a change in pressure in the contraction passage which in turn effects a
change in flow of the fuel to join and mix with the main airflow in the required
proportion (Heywood, 1988).
The sizing of the mixer was basically based on the air flow that is drawn by
any engine (Maxwell, 1995). A mixer would also cause pressure drop to the air flow
as it was a device that restricts air flow to create fuel suction. A general
understanding of pressurised flow was also studied to understand pressure drop in
the CNG mixer.
15
2.3.1 Current Trends in CNG Mixer Design
The venturi mixer acts as a carburettor to meter the amount of fuel to the
engine. One research that show the characteristic consideration of the CNG mixer
was done while studying the exhaust gas recirculation (EGR) system. Baert (1999),
found that requirements for a good mixer are as follows:
1) A compact design of the mixer for the EGR system
2) Minimal flow restriction during the intake process. This means less
pressure difference.
3) A good suction pressure mainly in the throat due to venturi effect
from the pressure difference. This will enable more fuel to be sucked
into the system.
CNG mixers have also evolved in design. There are many types of mixer
available. The main three types of mixer design were identified as the venturi, fan
and venturi-burner mixer as shown in Figure 2.2 (Roslia, 2002).
Venturi Fan
Venturi-BurnerFigure 2.2 Type of CNG mixers currently being used in the market
According to (Mardani, 2003), the design with combination of venturi and
burner mixer had better performance than all the other mixer type. By measuring
16
the power output a deduction was found that the venturi-burner type of mixer had
the 5% value closer to petrol power output. This is as shown in Figure 2.3.
15
20
25
30
35
40
45
50
55
60
1500 2000 2500 3000 3500Engine Speed (rpm)
Pow
er (
kW)
Petrol
CNG venturi-burner
CNG venturi
CNG fan
Figure 2.3 Power test results for different mixer designs (Mardani, 2003)
In Mardani’s research, the inlet and outlet angles for the design are obtained
by looking at the pressure drop caused by the angle. The angle that produced lower
pressure drop was chosen as it does not restrict the engine. Apart from this, the
angles are also sized by the amount of turbulence and velocity at the throat. The
angle that has higher turbulence with lower velocity would cause a homogenous
mixture of AF. The design which consists of the combination of 60° of inlet angle
and 30° of outlet angle with 8 holes at the throat was proved to increase the engine
performance using CNG to standard of petrol engines (Mardani, 2003).
Apart from the angle, the holes of a CNG mixer are normally designed at the
throat of the mixer. This was found by Luiz (1996), who developed the Mercedes-
Benz Natural Gas Engine M 366 LAG with a lean burn system by incorporating a
venturi type mixer. The gas was added into the air stream from the holes of the mixer
at the narrowest section of the venturi (throat). The mixer was placed after the
turbocharger, near the intake valve. The air fuel ratio was controlled by the
17
characteristic of the pressure regulator, mixer and the gas flow valve which is
commanded by the engine management system.
Another important parameter is the size of the throat. The throat size affects
the engine performance. Maxwell (1995) observed this when testing on the GMC
(5735cc) engine. The engine was installed on a SuperFlow Model SF-800
dynamometer to obtain data such as engine speed, brake torque, brake horse power,
water temperature, inlet air temperature, exhaust temperature and barometric
pressure. The CNG mixer also gives lower flow rate compared to the General Motors
throttle body. This means the mixer causes a large restriction to the airflow. Peak
power was reduced by 16.2% and 19.7% for small and large venturi respectively
when running on CNG. They concluded that the large venturi produced less suction
than the small venturi hence induced less fuel to the engine. The power loss of about
10 percent was due to natural gas occupying significant portions in the intake system.
Other parameters that are also considered important in a mixer design are
nozzle distance from inlet, venturi tube diameter, throttle opening and mixture ratio.
This was observed by Mikio (1998), after looking at the mixing of natural gas and air
in a two-dimensional CNG mixer. The Schliren method was used to see the mixing
effect. The parameters were found to affect the mixing of natural gas and air in a
CNG vehicle. The inadequate mixing gave an adverse effect on engine combustion
and emission characteristics.
The location of the mixer in the intake system of the engine is another
parameter to note. The positioning can be done by various methods. According to
Lenz, (1992), the location of the mixer for a carburettor engine is as shown in Figure
2.4. It is applied before the throttle. In a fuel injection engine the mixer was usually
put after the throttle as shown in Figure 2.5.
18Air Inlet
GasSupply
Figure 2.4 Venturi upstream of the carburettor
Figure 2.5 Mixer after throttle in intake system of injection engine
AirInlet
GasSupply
2.3.2 Sizing of the Mixer Throat
As mentioned earlier, the sizing of the mixer throat was done based on the
air requirements of two stroke engine. The required air flow (Qa) depends on the
engine speed (N), the displacement of the engine (Vd) and the delivery ratio (Dr).
Knowing this, the air flow rate in a two stroke engine was calculated as,
Da VNDrQ (m3/s) (2.3)
From the equation 2.3, the delivery ratio is important as it depicts the
amount of air that will be drawn into the engine (Ferguson, 2001). The delivery
ratio was defined as,
Delivery ratio, Dr =densityambientvolumedisplaced
cycleperairdeliveredofmass (2.4)
19
Different engines provides different delivery ratio. The delivery ratio was
found to be as low as 85% of the airflow for naturally aspirated engine or it was as
high as 150% for supercharged engines (Willard, 1997).
Mohamed (1998), cited the highest suction that was created by the air stream
occurs only when the air velocity is traveling in near boundary of compressible and
incompressible flow, which is at 150 m/s. The contraction of the throat causes a rise
in velocity linearly and the highest velocity at the throat could not be increased above
150 m/s without neglecting the change in density of the flow. Knowing throat
velocity as 150m/s, the area of the throat was found from the continuity equation.
Thus, by solving the circular cross sectional area of the mixer’s throat, the diameter
of the mixer throat is found.
2.3.3 Pressure Drop in the Mixer
As mentioned previously in Bernoulli principle, the faster the fluid travels at
the throat of the venturi, the lower the surrounding pressure at that region as shown
in Figure 2.6.
Pressure at the throat is lowest when the velocity is the highest. When the
flow exits the venturi the pressure does not return to its original value. This is the
overall pressure drop due to the restriction of the throat as shown in the Figure 2.6.
This also occurs in a CNG mixer as the shape of the mixer is similar to the venturi.
20
321
Diffuser angleVenturi
Y
X
Pressure
Velocity
FlowDirection
Figure 2.6 Schematic plot of velocity and pressure across a venturi
Applying an ideal full sized venturi in the engine is almost impossible as the
venturi needs a long length to reduce pressure loss in the flow. To compromise this
carburettors use air funnels at the throat to provide minimal pressure loss and
creating the largest vacuum at the narrowest point. The overall pressure loss was
found very little during low flow separation occurring at funnel diffuser angles of 7°
to 12°. The usage of funnel can reduce diffuser length and the overall length of a
mixer (Lenz, 1996).
The pressure drop in a venturi can be calculated by using the Bernoulli’s
equation which is,
LHgzvP 2
21
= constant (2.5)
Since the axis mixer was along the horizontal axis, the effect of gravity was
neglected and a more simplified Bernoulli was arrived for two points as in Figure 2.6
as,
LHvPvP 222
211 2
1
2
1 (2.6)
21
The pressure drop is obtained by subtracting the pressure of the flow entering
the mixer to the flow exiting the mixer. This can be given by,
LHvvPPP 21
221221 2
1 (2.7)
Losses in pipes occur as the term HL and computations of the losses needs
also to be included. A simulation and experimentation was seen a good way to
compute the losses and also the pressure drop (Lenz, 1996)
A clearer picture of how the pressure drop occurs in an engine is shown in the
profile of pressure during the intake in Figure 2.7.
P0
airP Po, To
IntakeManifold
uP
thrP TDC BDC
valveP
P P
Engine
Figure 2.7 Pressure profile during intake stroke of an engine (Heywood, 1988) Po, To
is pressure and temperature at atmospheric condition; airP , is the pressure loss in
the air cleaner; is the intake losses upstream of throttle; , is the losses
across throttle; losses across intake valve.
uP thrP
valveP
Heywood (1988), also found that a large pressure drop would cause
performance problems for the overall engine operation. This is because a venturi
arrangement can only meter fuel over a certain range of flow rates and pressures.T
22
The selection of small throat size of mixer will cause a higher restriction of air
entering to the intake manifold at high engine speeds. As flow rates increase, the
venturi will begin to "choke". Consequently, the engine will not achieve desired
operation with respect to increased throttle opening.
At the other extreme, when the velocity of the air in the venturi was very low
for example during idle or start-up, the pressure drop across the venturi becomes
small. A too small pressure drop due to a larger throat also causes less suction to at
the mixer throat. This extremity concerns with engine starting, idle and low-speed
throttle response to inadequate suction. The advantage of slightly oversized mixer
appears on over-the road applications. This was due to the fact that it achieves
optimum performance at higher engine speeds.
In a specific throat size of the mixer the overall pressure drop of the device
can be seen. The pressure drop across the venturi will increase as the engine speed
increases. This condition is similar to any component that restricts air flow to the
engine. Figure 2.8 shows how the pressure drop graph is plotted for components in
the intake manifold (Heywood, 1988).
0
10
20
30
40
50
60
70
0 1000 2000 3000 4000 5000 6000 7000 8000
Engine Speed,rev/min
DP,
mm
Hg
Air Cleaner
Patm
Pthrottle
Throttle
Patm-Pthrottle
Figure 2.8 Pressure drop in air cleaner and intake manifold, Patm is the atmospheric
pressure and Pthrottle is the pressure after the throttle (Heywood, 1988)
23
Since the pressure drop will increase as the air flow in the mixer increases the
fuel flow will also increase due to suction created by the lower pressure at the throat.
This reduces AF to a richer mixture. Generally engines require a high amount of fuel
at high speeds and the mixture of air and fuel tends to get richer as the engine speed
increases.
The mixer that can create more suction of fuel can have smooth engine
operation as fuel flow will not bottleneck the performance of the engine at high
speeds. Nevertheless, a too rich mixture would also cause the engine to stall as the
fuel is more than oxygen to burn. As discussed earlier on CNG fuel mixture, a
stoichiometric mixture is important to enable proper fuel combustion. With this the
mixer should allow the air fuel mixture to be near the stoichiometric range.
2.3.4 CNG Mixer and Engine Conversion Kits
The CNG mixer is the part that connects the engine with a CNG conversion
system. The system itself is made from many components that functions to bring fuel
to the mixer when the engine is operated. There are many types of conversion that
can be used to convert an engine to run on CNG. The conversion systems can be
grouped into three groups which are:
1. Bi-fuel engines, this is a spark ignition petrol engine converted to
natural gas by fitting various components such as a gas
mixer/carburettor, regulator, shut-off valves, control systems and fuel
storage tanks. This arrangement retains petrol fuel system which can be
used when CNG refuelling facilities is unavailable.
2. Dedicated natural gas engines, which are engines optimised for natural
gas. They can be made from petrol engines or be designed mainly for
CNG usage.
3. Dual-fuel engines, these are diesel engines that operate with mixture of
natural gas and diesel fuel. Here, a pilot fuel (diesel) is injected within
the gas mixture so that ignition can occur. This injection is required
because natural gas has a low Cetane rating which is suitable for
24
compression ignition engines. The pilot fuel will ignite first to start the
ignition process.
In the market bi-fuel conversion was a popular option as the user can change
between the fuels without making drastic changes to the existing engine. In the
research the bi-fuel conversion kit was chosen to be studied as it can be obtained
readily from the market as in Appendix C.
2.4 Summary of Literature Review
The literature obtained for two stroke engine and CNG fuel was a preliminary
knowledge for the development of the mixer. The different designs of CNG mixer
showed that the venturi-burner mixer was better and has better performance in actual
testing.
The working principle of the mixer was based on the venturi principle from
Bernoulli equation. It creates a low pressure in the throat area so that natural gas can
be sucked into the engine. The calculation of throat size was seen possible by
knowing the air speed and usage of continuity equation.
Pressure drop causes fuel suction. Due to this, fuel and air mixing
characteristic must reach a certain characteristic that will enable proper operation of
the mixer in the engine. A near to stoichiometric operating condition will allow the
mixer to work efficiently by mixing the fuel and air to the required burning condition
of the engine so that complete combustion of the fuel occurs.
CHAPTER 3
DESIGN OF A VENTURI BURNER MIXER
The research is based on the CNG mixer development for the two-stroke
engine. The flow of the research was carried out as shown in Figure 3.1. The findings
of the study have aided the researcher to begin the designing process of the mixer.
Start
Conceptual Design
Procedure of Mixer Design
Computational Fluid Dynamics Simulation
Design and Fabrication
Validating the Mixer Design
End
Figure 3.1 Methodology for designing the CNG mixer
26
3.1 Conceptual Design
Firstly the design process starts with a concept design. The initial idea for the
shape
ed for 150 cc supercharged two stroke engines.
ent.
at
l angles were used for the inlet and outlet angles to reduce
made along the circumference of the throat to dispense the
nents rather than a single part as the
terial was considered to be used to
s the venturi burner type of mixer shape was for design, many ideas were
though
he first concept in Figure 3.2(a) shows a two inlet mixer, the two inlets
require
design was not developed and needs some improvements.
of the mixer was done from some sketches. The sketches were developed
based on the design criteria. This had been predetermined for the mixer design by
considering machining process involved in the prototyping of the mixer. The design
criteria were set as follows:
The mixer was siz
The length of the mixer must be 60 mm to fit the engine compartm
It has to be a venturi burner type CNG mixer with a diffuser angle of 7°
the throat.
Symmetrica
mixer size.
Holes were
CNG fuel to the air stream evenly.
The model was divided into compo
process of machining would be easier.
A light weight and easy to fabricate ma
make the mixer.
A
t for the shape of the mixer. These shapes are as shown in Figure 3.2. The
shapes were in a development stage, where a constant improvement of the previous
design was made. Each shape was discarded until a final acceptable design was
reached.
T
modification to the piping of the CNG kit. Having two pipes for CNG also
creates a burden for installation of the mixer. The complexity of the shape also
demands high machining cost as the process of machining identified are casting and
rapid prototyping. This increases the production cost of the design. Due to this, the
27
Figure 3.2 The concept models
Figure 3.2(b) show ngential porting with one
CNG inlet. This idea was to give a swirl to the CNG flow but was quiet hard to be
fabrica
bricated. Figures 3.2(c) is a
oncept model with a section view. The model was split in components rather than a
whole
(b)
(c)
(a)
s the concept of the mixer with ta
ted with conventional machining. A casting process was still needed. The cost
of fabricating this shape is also estimated to be too high.
The design had to be cost effective and easily fa
c
part for machining. This also reduces the machining cost as complex
machining is not needed. A turning process was chosen to create this concept. This
concept model also uses circular holes made along the circumference of the throat.
28
The holes are easily made by drilling process. Casting process was not needed for
this concept.
The last concept model was accepted as the initial shape of the mixer. This
design was found to be easier to be fabricated and cheaper due to its simple
machining process. The concept was further refined until the shape in Figure 3.2 as
the proposed design. Further simulation in CFD will continue to optimise the
proposed design.
Proposed shape of the mixer
.2 Procedure of Mixer Design
ed to start the design of the mixer. The design
riteria that were outlined were fulfilled in the process of designing. The design starts
Inlet
Outlet
Throat
CNGInlet
Hole at the throat
Figure 3.3
3
The conceptual design was us
c
by finding the initial throat size of the mixer by using equations found from
literature. The other dimensions of the mixer was then finalised from CFD simulation
and analysis. CFD was a tool to simplify the design process.
29
3.2.1 Initial Throat Size
based on the specification given by a two stroke engine.
ns of the engine are as shown in Table 3.1. Hoses that fit the two
stroke
an, 2002)
The mixer was sized
The specificatio
engine were found to be of 36 mm diameter. The size was used to target the
outer diameter of the mixer so that the mixer can fit the hoses available in the market.
A thickness of 2 mm was given to the mixer walls for connecting purposes, thus
making the inner diameter of the mixer to be sized at 32 mm.
Table 3.1 Specification of the two stroke engine (G
Volumetric displacement 150 cc
Bore x Stroke 57.6 mm
Maximum engine speed 8000 rpm
Type of scavenging system Uniflow scavenging
Effective compression ratio 8.2866
Maximum combustion pressure 49 bar
Maximum temperature 2688.88 K
Type of fuel Petrol
Supercharger Blower
Since the inner diameter was predetermined by hose size, the initial size of
e mixer throat was calculated from air flow characteristics of the two stroke engine.
In this
th
engine the air flow was increased by the use of a blower mechanism. Due to
this the engine air flow is calculated by taking the deliver ratio of the two stroke
engine as 150% as suggested by Willard (1997). The equation for air flow of the two
stroke engine at any given rpm was given by,
60
5.1 Da
VNQ (m3/s) (3.1)
By knowing the air flow of the engine, the initial approxim tion of the throat
ize was computed by using continuity equation. An assumption of inviscid and
a
s
30
steady flow was used. As known earlier in Chapter 2, the throat size was the
minimum cross section area of the mixer that makes the air flow reaches speed of
near compressible flow. Figure 3.4 shows where the throat is located in the mixer.
Maximum air flow rate of the engine was used to find the throat area.
Figure 3.4 Location of throat diameter
V1
ThroatInlet Outlet
V2D2D1
The maximum equation 3.1, using
e maximum engine speed of 8000 rpm. The air speed at the throat was taken as 150
m/s as
theoretical air flow rate was found from
th
discussed in literature. Thus, the throat area was obtained by solving the
continuity equation for incompressible flow as,
22
112 V
Q
V
VAA a (3.2)
The initial throat diameter, D2 was then obtained by solving the circular area
btained from equation 3.2. The throat diameter was used as a reference for the
.2.2 CFD Simulations of the Mixer
G mixer continues to achieve the criteria set
arlier in the concept stage. Simulations need to be done to save time and cost of
designi
o
beginning step of designing the concept model of the mixer.
3
The process of designing the CN
e
ng the mixer. As the air flow in the mixer is governed by fluid mechanics, the
simulation of fluid motion would be needed to analyse the mixer. A computational
31
method is chosen to calculate the fluid motion, by using a commercial CFD package,
CosmosFloworks 2001 (Rosli, 2004).
In CFD, the fluid motions are determined by solving the the Navier-Stokes
equation of mass, momentum and energy equation. The three equations can be
written in the conservation form as follows:
0ukkxt
(3.3)
ii
ikkik
i Sx
Puu
xt
u (3.4)
Hkkiikkkk
QuSuquPExt
E (3.5)
here u is the fluid velocity, W is the fluid density, Si is a mass-distributed external
tforce per unit mass, E is the to al energy per unit mass, QH is a heat source per unit
volume, ik is the viscous shear stress tensor and qi is the diffusive heat flux.
Turbulence normally occurs in any fluid flow. Prediction of turbulent flow
was done by the Reynolds averaged Navier-Stokes equations. Through the averaging
procedure, extra terms known as the Reynolds stresses would appear in the
equations. Following Boussinesq's assumption, the Reynolds stresses was defined in
this model as:
ijijl
l
i
j
j
it
Rij k
xu
x
u
xu
32
32
(3.6)
t is defined using two basic turbulence properties, namely the turbulent kinetic
energy, k and the turbulent dissipation rate, . A constant is also derived empirically
as, C in the equation with value of 0.09.
2kCft (3.7)
hus, theT k model is used to solve turbulence by the software package.
Generally, the CFD analysis follows steps shown in Figure 3.5.
32
Figure 3.5 Simulation steps for each simulation
After the initial mixer modelling w CAD, the CFD analysis continues with
mesh generation. It is done by splitting th volume of fluid in the model into small
volume
boundary conditions inputs were given to the CFD software to begin the calculations.
The ac
t
ith
e
s which is called mesh and iterates the calculation for each volume The CFD
package uses rectangular mesh to the volume of the fluid in the design (Rosli, 2003).
When a specific a mesh was reached as in appendix F, the mixer model’s
curacy of the boundary is important as it would determine the simulation
outcomes. Three boundaries are needed to analyse the CNG mixer they are,
1. Air inlet
2. Flow rate at outlet
3. CNG inle
Modelli erng of mix
Meshing of model
Sett ioning boundary condit
Iterations
Satisfactory
YesNo
Start
Ana ltslysing resu
End
33
1. ir inlet
oundary is the air inlet. A static air inlet was given assuming the air
at the starting of the simulation was not moving. The properties of air were taken as
shown
Pressure 101325 Pa
A
The first b
in Table 3.2. The values of pressure and temperature were taken from an
aneroid barometer from the lab. The other values are interpolated from fluid dynamic
air properties found in any fluid properties table at the given pressure and
temperature. The air property was used as the first input to the inlet of the mixer.
Table 3.2: Properties of air
Temperature 299.15 KSpecific Heat, gKCv 1006 J/kDensity, 1.225 kg/m3
Dynamic viscosity, 1.80x10-5 kg/ms
2. Flow
e engine operates the air will get sucked into the intake
anifold. This causes the flow rate at the outlet of the mixer. The suction flow rate
was cal
undary was the CNG fuel inlet. The fuel is in ambient pressure as
in actual usage a bi-fuel system delivers the fuel at ambient conditions. This
bounda
ulation strategy of the mixer in CFD was done in stages to determine the
required dimensions of the mixer. The dimensions are to meet the design criteria set
based on fabrication needs. The simulation strategy would follow Figure 3.6.
rate at outlet
When the two strok
m
culated from equation 3.1. This flow rate is the second boundary to the mixer.
The boundary was calculated and simulated for engine intervals of 1000 rpm.
3. CNG inlet
The third bo
ry allows the software to compute the amount of fuel sucked into the mixer
by means of the venturi shape and the air flow of the two stroke engine by the mixer.
Methane properties that are available in the software are similar to that found in
literature.
Sim
34
Start
Stage 1 Inlet and outlet angle analysis
Stage 2 Number of hole umferences at throat circ
Stage 3 Size of hole a rcumferencet throat ci
Stage 4 Throat size isationoptim
End
Figure 3.6 Overall simulation stages done on the mixer
.2.3 Inlet and Outlet Angles of the Mixer
Firstly, simulations done on Stage 1 are to obtain the inlet and outlet angles
e needs to be at maximum when the
ngine speed is at 8000 rpm using equation 3.1. This airflow rate was applied at the
mixer o
ith a 7°
diffuser angle to reduce flow separation. A 5 mm gap was at least needed to enable
pipe fit
3
for the mixer. For the analysis, the airflow rat
e
utlet while ambient pressure opening was given to the inlet. The air in the
mixer is travelling at the highest velocity according to the top engine speed.
During the simulation for the angle analysis the throat size was kept at the
initial size obtained from equation 3.2. The throat was given a design w
tings to the mixer at both ends. After giving the minimum gap at the ends, the
angles were computed using simple trigonometry. The throat length was
predetermined to be 20 mm from comparison of existing four stroke mixers
available. The angle was then varied to make the mixer fit in 60 mm. Figure 3.7
show where the inlet and outlet angles was placed in the mixer for CFD analysis.
35
Figure 3.7 Simulation model for inlet and outlet angles
The pressure across the mixer was found for the whole length of the mixer.
The pressure p cs of the air
ow in the mixer. A graph would show the effect of air flow pressure at each angle
inlet and outlet are not same in the
ixer. The most effective angle is the angle that produces the highest suction per
pressur
he hole drilled at this location would allow the mixer to operate
most efficiently as the suction of fuel is the highest.
result was subtracted from the
inlet pr ssure, P1 to get the highest suction reading. The overall pressure drop also
Inlet angle Outlet angle7°Diffuser
Gap5 mm
60 mm
P1 P2P3
lot was needed to see the overall pressure characteristi
fl
and to lead to the choosing of the best angle.
The criteria for choosing the right angle was by looking at a pressure ratio.
The reason for the criteria is that pressure at the
m
e drop. The ratio was found by finding the pressure difference from the inlet
to the highest suction and dividing it with overall pressure drop in the system. The
highest suction occurs at the location of the fastest air movement in the mixer. This
location happens at the throat because it restricts the air. As the mixer needs holes at
the throat based on the design criteria, the location of the highest suction was seen on
the throat wall, P2.
A pressure plot along the wall was needed to show exactly the point of
maximum suction. T
Knowing this location, a surface integral by the software was then used to
find the averaged pressure at this cross section. The
e
36
can be found by subtracting the pressure of the inlet, P1 from the outlet, P3. The ratio
was obtained from the equation:
Pressure ratio =31
21
PP
PP (3.8)
e efficiency of each angle that could
produce the higher amount of suction per pressure drop needs to be obtained from
FD analysis. The highest efficiency was chosen to be the angles used for the
sign.
of Holes at Throat Circumference
Obtaining the angle and location of the hole leads the design process to Stage
i at circumference of the mixer
esign. A CNG inlet was added to the design and the boundary of fuel is applied to
this inl
t speed, the mixer generates the lowest suction. The
low suction must be sufficient at the fuel inlet to enable the fuel to flow due to
suction
two holes at once. Then, the hole size were
determ itting of the number of holes at the throat. With even numbers,
the num
A pressure ratio graph that shows th
C
de
3.2.4 Number
2. Wh ch is to find the number of holes at the thro
d
et as in Figure 3.3. The size of the inlet needs to match the hose that connects
to the available conversion kits.
For this stage, the boundary of air should be for lowest engine speed of 1000
rpm. This is because at the lowes
generated in the throat. The profile of fuel mixing with air must be qualitative
and needs visualisation using CFD.
Firstly, the numbers of holes were taken as even numbers to ease machining
as the drill can be used to make
ined by the f
ber of holes that covers the most area at the throat needs to be chosen. This
was to not waste too much material between the angles at the circumference. After
this consideration, the number of holes was increased and simulations are done for
each number. The number of hole that gives a good mixing of fuel qualitatively at
the throat is finalised for the number of hole. A cross-sectional plot of the mixer at
37
throat was used to visualise the mixing. The colour contour plot from CFD
simulation decides which number hole gives better mixing.
3.2.5 ize of Hole at Throat Circumference
tage 3 of the process in designing the mixer was to enhance the size of hole
the number of holes was determined
Stage 2, the size of the holes are based on commonly available drill sizes. The size
uel flow at the fuel inlet should be in the range of 9.77 at the
chest limit and 17.23 at the leanest limit. A lower than 9.77 AF would cause the
mixer t
oat Size Optimisation
The last simulations in Stage 4 were conducted to find the correct throat size
determined. A variation of the throat size was
one by giving a small increment to slowly increase the throat size and looking at
S
S
to get good fuel and air mixing at the throat. As
in
of the hole would affect the AF ratio of the model. AF ratio given by the different
hole sizes was seen for the entire operation of the mixer from 1000 rpm to 8000 rpm.
To choose a suitable hole size the AF ratio is compared between the hole sizes
simulated at the throat.
The AF ratio obtained from dividing the surface integral of the mass airflow
at the air inlet and mass f
ri
o have too much fuel suction and this would cause the engine to have less
oxygen to complete the combustion. Meanwhile, a too lean AF ratio above 17.23
means less fuel sucked by the mixer and the hole size restricts the fuel suction. This
could cause incapability of the engine to combust the mixture. The AF ratio obtained
must fall in this range of rich at 9.77 and lean limit of 17.23 to choose the hole size to
be drilled.
3.2.6 Thr
after all the other parameters were
d
CFD results that were obtained by the enlargement process. Two parameters are
obtained by this enlargement to find the best throat size. The first being the AF ratio
and the second was the pressure drop.
38
The choosing criteria for the throat size was the first parameter, this is the AF
ratio. The throat size was increased from the initial throat size to the appropriate
throat size that gives AF ratio near to the stoichiometric limit of 17.23. The limit
nables
at gives the restriction to the air flow a pressure drop result is
expected in this simulation. The pressure drop value can be used for the validation of
the sim
.3 Prototyping the Mixer
After the CFD predictions were made, the mixer was built as a prototype. The
a xer drawings from the CAD drawings was sent for
brication of the prototype.
or better known as Perspex. With this material, a
transparent mixer was fabricated. The prototype needs to be created transparent to
visualis
use it is easy to machine, light weight and easily available
in the market. In addition, Aluminium is also corrosion resistant. These are reasons
why Al
e the mixer to mix air and fuel to the correct combustion value. A venturi is
predicted to give range of AF ratio and is constant. Due to this, a range of AF ratio
near stoichiometric value at a certain range of engine speed needs to be considered.
The mixer throat that can give AF ratio range near stoichiometric condition is chosen
to be fabricated.
The second parameter obtained from throat enlargement was the pressure
drop. As the thro
ulation results. Two validations were done to prove that the mixer simulation
is accurate in real life.
3
inform tion of the optimised mi
fa
Two prototypes were aimed to be generated. The first was a mixer made from
polymethylcrylate (PMMA)
e the flow of gas.
The second mixer was targeted to be made entirely from Aluminium. This
material was chosen beca
uminium was chosen as the material for the second prototype.
39
Fabrication process for machining the prototype are turning and drilling as
described earlier in the conceptual design of the mixer. The process was applied for
both m er prototypes.
3.4 alidating the Mixer Design
he two prototypes were made to validate the simulation results. Tests were
o tions to get a firm confirmation of CFD and
ctual testing. The first test was using the Perspex prototype of the mixer. This
.4.1 Testing Apparatus
The first apparatus was a flow test rig that is readily available at the
ntre (ADC) lab in UTM. The flow rate of air was
easured by this test rig. A flow test rig was used to test the mixer with air flow
similar
ix
V
T
done n the prototypes in three varia
a
prototype was used to visualise the CNG mixing in the mixer as the prototype is
transparent. Smoke was generated and flowed into the CNG inlet to replicate the gas
movement. The second test was the AF ratio that the mixer provides during its
operation. The AF ratio shows the amount of air and fuel that the mixer induced due
to suction from the engine. The third test was the pressure test, where the overall
pressure drop was found. The results were compared with simulation results. A set of
equipment were targeted to do all three tests.
3
Automotive Development Ce
m
to the engine conditions. The schematic diagram of the test rig is shown in
Figure 3.8. A centrifugal blower was used to generate suction and blowing to the
mixer prototype. The electronically controlled blower can produce the equivalent air
flow of the two stroke engine. The unit of measurement in the flow test rig is in litres
per minute (LPM). Air flow given by the blower is measured by means of a
calibrated laminar flow element.
40
Laminar Flow ElementMixer
DigitalDisplayUnit
PVC Piping
Centrifugal BlowerInverter
Figure 3.8 Schematic diagram of flow test rig to measure air flow
As the flow element of the flow test rig was used for a long time, a
recalibration needs to be done. Calibration was done by measuring the air flow at the
flow outl he flow
rate, Q ound from the test rig was converted into velocity. A velocity meter was
used to
) Measurement of AF ratio
i)
he first test is to visualise the mixing of two fluid flows in the mixer. As
CNG is colourless, the fuel was replaced with smoke to see the mixing clearly. The
et area. The blower was then set to the predetermined flow rate. T
f
measure the velocity again at this outlet. Finally, the two velocities are
compared to get a correction of the air flow rate given by the test rig.
The calibrated flow test rig was used with other measuring instruments to do
all three tests. The following tests were carried out for the mixer prototype:
i) Flow visualisation
ii
iii) Pressure drop across the mixer
Flow visualisation
T
41
vis ation was done by cualis ombining the use of the test rig with a smoke generator to
visually analyse the mixing phenomenon. The generator was available at the
Thermodynamic Lab in UTM was chosen as it provides continuous smoke at
atmospheric pressure. Figure 3.9 shows the schematic location of the smoke
generator when it is assembled to the test rig.
LaminarFlowElement
Figure 3.9 Schematic of smoke generator connected to tes
The gaseous
ow meter was used to read the flow rate at the fuel inlet of the mixer. The meter
. The gaseous flow meter measures the flow rate in
LPM also. The gaseous flow meter is as shown in Appendix C. With this instrument
both th
across the mixer
he third test was the pressure drop test. As the mixer causes pressure drop to
the air flow, a pressure test was needed to be done by combining the test rig to a
used by connecting it directly to the hoses,
t rig
Smokegenerator
ConeAdapter
Black screen
Mixer
Adapter
Camera
ii) Measurement of AF ratio
second test was the measuring of the AF ratio in the mixer. A
fl
was also available at the ADC lab
e measurements of air flow rate and fuel flow rate were obtained to do the AF
ratio measurement.
iii) Pressure drop
T
digital manometer. A manometer was
42
giving reading of pressure difference. The manometer was able to measure the range
of the p
Figure 3.10 ent
he chosen apparatus was used to do expe set of testing
procedures. The procedures are followed as a guide to conduct the tests.
he goals of the experiment were to validate the CFD simulation. Each
experim nt was done based on the air input calculated for each engine speed. The
t rig. Air speed was taken when the system comes to a
teady condition. The minor flow fluctuation also needs to be minimised. The errors
were re
ressure drop given by the mixer. A schematic of the manometer connected to
measure pressure drop is shown in Figure 3.10.
Schematic diagram of pressure measurem
2
1 3
Mixer
Digital manometer
T riments based on a
3.4.2 Testing Procedure
T
e
values were set at the flow tes
s
duced by doing many tests and averaging all the results obtained.
43
3.4.2.1 Smoke Mixing in Mixer
In the first experiment of smoke mixing of mixer, the air flow was set in the
ow test rig and pictures of smoke were taken at different engine speeds. The smoke
hows the path taken by the CNG if it was used in the system.
he smoke was generated by a smoke generator connected to the test rig. A
blowin
smoke flow from the
surroundings.
the second experiment, the combination of laminar flow element and
gaseous flow meter readings gives flow rate of air induced by the blower and the
flow rate of CNG. In actual test, the fuel that needs to be used to find AF ratio was
rig. Due to some limitations, the CNG fuel could not be used
the test rig. The limitations are as follow:
water to the regulator. As the mixer
2)
fl
s
T
g condition was used in the flow test rig to eliminate smoke from entering the
blower. A flow pattern was seen to determine whether the design can induce mixing.
A picture was taken with a black screen to differentiate the
3.4.2.2 AF Ratio Test
In
CNG from a bi-fuel test
in
1) The bi-fuel conversion kit needs heated water to heat the decompressed
gas from the storage tank. As the gas is regulated from 200 MPa to
atmospheric pressure the regulator becomes extremely cold and reduces
the flow of gas due to condensation in the regulator. Normally the engine
cooling system provides heated
needs to be tested separately a heating unit needs to be designed. The
limitation in resources did not allow the design of a heating unit.
The operation of the bi-fuel kit also needs input of spark timing from the
engine. This timing enables the regulator to energise the solenoids to
open and flow the CNG fuel to the mixer. This was seen as a limitation
as a timing device needs to be designed to operate the bi-fuel system if
the mixer was tested separately.
44
3) Apart from this, the usage of CNG in open testing imposes some safety
issues. The gas could not be contained after the testing and the highly
combustible gas would ignite if strict safety procedures are not applied.
main task here is to measure thThe e AF ratio even though the fuel is not
availab
readily ava
properties s e
to change the air properties to methane properties. The flow similarity follows
Reynol
or the first step, since the flow meter is calibrated to methane, the reading
shows
r manual (Sierra, 1994).
From the manual a K, factor was used to correct the flow rate measured as Methane
to the a
le. As CNG or methane cannot be used, air is used to replace the fuel as it is
ilable. Nevertheless the composition of air does not share the same
uch as density and viscosity with methane. A similarity analysis was don
ds number similarity (Andreas, 2001).
Two steps are taken to analyse the fuel flow. The first was to convert the fuel
flow rate measured by the gaseous flow meter to actual air flow rate. The second step
was to do the similitude analysis to find the fuel flow rate.
F
the flow rate of methane instead of air. This produces errors to the readings if
the correction of the flow is not done. The actual air flow rate was found by
calculations procedures given by the gaseous flow mete
ctual flow rate of air from the meter.
air
Methane
K
K
Q
Q
2
1 (3.9)
Subscripts 1 and 2 denote the measured and actual flow rates respectively. MethaneK is
0.72 and airK is 1.00 from the manual. After the air flow rate is obtained the value
was converted to air velocity by looking at
The air velocity was used in step two to change it back to fuel velocity.
pro
the cross sectional area at the fuel inlet.
For the second step, similarity of the flow rate at the CNG inlet need to be
used. The totype tested with air is geometrically similar to the model tested with
methane. For making both prototypes similar, a dimensionless property must be
made similar. The dimensional property is Reynolds number given as,
45
vd(3.10)Re
Where, is the fluid density, d is the diameter of the pipe, v is the fluid
velocity and is the viscosity of the fluid.
Using flow sim
ethane and the model with air. This can be written as,
ilarity the Reynolds number is same for the model with
m
AirMethane
vdvd (3.11)
Thus, the corrected air velocity can be converted to methane velocity by,
airair
MethaneairMethane vv
Methane
(3.12)
he velocity of methane is then converted to mass flow rate of fuel to the mixer,T
inletCNGMethaneMethanef Avm (3.13)
After getting the mass flow rate of fuel the mass flow rate of air has to be
und in order to find the AF ratio. The suction of the blower combines the two air
ing
the air flow rate, the air flow rate from inlet, te
flow test rig, . The subtracted flow rate is, . The
e AF ratio. The results
a gr
experiment and simulations. The air induced by the blower at the inlet is obtained by
fo
flow from the CNG inlet and the air flow from ambient condition, Q . For findTotal
the CNG , was subtrac d with the 2Q
readings shown on by Total air
airflow need to be converted into air mass flow rate, am by including air density.
The results am are divided by fm to obtain th are
calculated for each engine speed and aph of AF ratio was compared between
subtracting the flow rate obtained by the gaseous flow meter.
Q Q
46
A repetition of 10 times was done. Again the average readings were is taken.
The AF ratio was calculated and graph of AF ratio versus engine speed was plotted.
A comp
The third experiment was done to measure pressure drop at different engine
peeds. The pressure drop was measured in Pascal (Pa), as shown in Figure 3.9
3. Points 1 and 3 are tapped at 20 mm away from the
mixer so as not to be too far away from the mixer inlet and outlet. The CNG inlet,
point 2
f 1000 rpm. A plot of pressure drop
versus engine speed was produced.
ined and discussed.
arison was done on the two graphs.
3.4.2.3 Pressure Drop Test
s
previously for points 1 and
, was left open to atmospheric condition as the given input in the simulation
model is methane at atmospheric conditions.
Since there is fluctuation in the reading, the experiment was done 10 times.
The average of the 10 readings was taken for each pressure reading obtained.
Pressure readings were taken for intervals o
Validation was done on both the simulation of pressure drop and AF ratio.
Graphs of CFD analysis were compared with the corresponding experimental graphs.
The error of the two graphs was obta
CHAPTER 4
RESULTS AND DISCUSSIONS
The prototype of the mixer was based on the methodology planned in Chapter
3. Each step of the method was followed and results at each step were used to
continue on the other step. The steps had led the research to fabricate the prototype
of the mixer.
4.1 Designing of the Mixer
The process of design continues with the final concept of the mixer as seen in
Chapter 3. Calculation and CFD simulation was done to do the design for reduction
of cost in producing the mixer.
4.1.1 Initial Throat Size
Mixer throat size was determined by the maximum air flow characteristics of
the two stroke engine. The purpose was to find the correct inner diameter of the
throat which would be used to design the mixer prior to CFD analysis. The size of
the throat from maximum engine speed of 8000 RPM was given by,
s
mV
NQ Da
3
03.060
5.1
48
With throat air speed at compressible limit, was assumed as 150 m/s. The
throat area was given by:
2v
2A =2v
Q= 2.0 x 10-4 m2,
The diameter was found as: mA
d tt 016.0
4
The throat diameter was found to be 16 mm. The size is a guide to further
develop the design to meet the design criteria set earlier in Chapter 3.
4.1.2 CFD Simulation of the Mixer
A number of simulation were done to find the design dimensions. The results
were obtained for the following dimensions:
1) Inlet and outlet angle of the mixer
2) Number of holes at throat circumference
3) Size of holes at throat circumference
4) Throat size optimisation
4.1.3 Inlet and Outlet Angles of the Mixer
The results of CFD were plotted in Figure 4.1 as a pressure graph. The graph
showed the effect of the changes for each angle that was analysed. The range of the
angles for the inlet and outlet started by including a predetermined 5 mm gap to fit
the hoses at the inlet and outlet of the mixer. This was shown in Figure 3.7 earlier.
With this the simulations were done with the angles were varied from 60°, 70°, 80°
and 90°.At angles less than 60° the configuration does not meet the 5 mm gap
decided from design criteria and these angles would not be chosen to analysed. At
more than 90° the angle produced large opening from the inlet to the throat which
was seen as a waste of material to fit the 60 mm length.
49
84000
86000
88000
90000
92000
94000
96000
98000
100000
102000
0 0.01 0.02 0.03 0.04 0.05 0.06
Mixer Length (m)
Pres
sure
(Pa
)60 deg70 deg80 deg90 deg
Outlet
Inlet
2
1 3
Inlet and outlet angles
P3
P2Diffuser
P1
Figure 4.1 Pressure plot along the centre line of the mixer at different inlet and outlet
angles; 1 indicates inlet, 2 indicates throat diffuser and 3 indicates outlet
The graphs showed clearly the effect of the constriction at the throat on the
pressure profile. The graph also proves Bernoulli effect that the contraction at the
throat size cause higher fluid flow velocity at the throat which causes pressure to be
lowered. The converged air stream lines exert pressure to the wall. This causes more
pressure to be built up at the inlet of the mixer. When the air passes the throat, the
high pressure difference forces the air velocity to increase and reduces the pressure at
the throat. The pressure curve was found similar to the theoretical curve found in
literature as shown in Figure 2.6.
The stream lines will eventually diverge by the angle given at the throat
diffuser and outlet. The diverged flow looses its velocity and eventually increases the
pressure at the outlet to almost back to initial pressure at the inlet. The pressure does
not return back to its original condition because energy was used to overcome the
restriction. Judging by Figure 4.1, the lowest pressure drop occurred at 90° angle, but
the amount of pressure drop from the intake also was considered high comparatively
to the other angles. A choosing criteria needs to be followed at this stage.
50
The criteria to be met was that the angle must give high suction but with
lower overall pressure drop by looking at the pressure ratio. To find the pressure ratio
the point of highest suction must be found first. An accurate value of this lowest
suction was found by looking at the pressure plot along the throat wall in Figure 4.2.
83000
84000
85000
86000
87000
88000
89000
90000
91000
92000
0 0.005 0.01 0.015 0.02
Diffuser Length (m)
Pres
sure
(Pa
)
60 deg70 deg80 deg90 deg
MIN
Inlet and outlet angles
Figure 4.2 Lowest pressure at the throat diffuser wall.
From Figure 4.2 it can be seen that the lowest pressure occurs at the same
location for each angle. The location was at a point 4 mm from the diffuser. The
location also measured 24 mm from the mixer inlet if seen from Figure 4.1. This
point was taken as the point to make the holes along the circumference of the throat.
The location of the holes was then used to find the integral at the cross-
section. A pressure ratio was found by including the average surface integral at the
inlet throat and outlet of the mixer. The pressure ratio was obtained from the
equation 3.10. The angle that gives best pressure ratio was then found from a graph
plotted as shown in Figure 4.3.
51
1
1.1
1.
1.
1.
1.
1.
Pres
sure
Rat
io, (
P
2
3
4
5
6
1.7
1.8
50 60 70 80 90
1- P
2)/(
P2-
P3)
Inlet and Outlet Angles, º
Figure 4.3 Pressure ratios of each model inlet and outlet angle changes
From Figure 4.3, it was obvious that the 60° gave more relative suction
pressure to the overall pressure drop. The high suction was seen important as the
mixer was used to suck in fuel. On the other hand, the prediction of low pressure
drop ensures less energy was used by the mixer to overcome the restriction given.
The saving of energy here translates into a more efficient engine operation. The ratio
also depicts the efficiency of the angles. The efficiency reduces as the angle
increased due to more flow separation that occurs during the expansion at the outlet.
This shows that sharp angles are bad for smooth airflow as more turbulence were
produced.
As a result the inlet and outlet angle of the main body of the mixer was
optimised at 60 . The simulation also showed that the lowest suction was
approximately at 24 mm from the inlet of the mixer for all angles simulated. The
reason being the flow entering the throat restriction produces a vena contracta. This
causes the highest velocity in the mixer happens at a slight distance from the throat
and not at the throat. This is as shown in Appendix B. The high velocity in turn
Two views of mixing air and fuel were seen. The views are a 3 dimensional
view for overall comparison of the mixing and a cross-sectional side view at the
throat holes. The results are shown in Figure 4.4, 4.5 and 4.6 for the three numbers of
holes simulated. The quality of the flow from the simulation plot shows how well the
mixing occurs.
As the largest number was 12, a 3 mm hole size was used to start the
simulation. A larger than 3 mm size does not fit the mixer throat with 12 holes. It
causes jagged edges at the throat due to the holes being too close to each other.
Simulations were done to find the number of holes at the throat
circumference. Since the number of holes was made by considering even number of
holes, 8 holes is the most likely number to be simulated first. When increasing the
number of holes it was found that they caused decimals when dividing with 360°.
For example, if 8 hole numbers was chosen it can be easily divided with the 360° in
the throat giving 45° per hole. If the number was 14 holes the division is 25.714° per
hole. It is difficult to fabricate the design if there are decimals in the design.
Simplification without decimal was done. Twelve holes were found to be the
maximum increment without causing decimals in the angle calculated. Due to this
the numbers of holes were simulated for three conditions of 8, 10 and 12 holes.
4.1.4 Number of Holes at Throat Circumference
reduces the local pressure in the area. The location of the low pressure was used to
drill the location of the holes along the circumference of the throat.
52
1
82
Thr
oat
= 1
6 m
m
Hol
e =
3 m
m
73
64
5F
igur
e 4.
4 E
ight
hol
es m
ixer
mod
el
53
12
10
39
Thr
oat
= 1
6 m
m
Hol
e =
3 m
m
84
57
6
Fig
ure
4.5
Ten
hol
es m
ixer
mod
el
54
55
4
3 5
2 6
1
78
12
11 9
10
Thr
oat
= 1
6 m
m
Hol
e =
3 m
m
Fig
ure
4.6
Tw
elve
hol
es m
ixer
mod
el
With the number of holes determined, the CFD analysis continued to find the
better hole size based on available drill bits in the market. The largest size of the drill
was considered by the size govern by the 12 holes chosen earlier in the design stage,
while the smallest is the smallest easily available drill bit in the market. The drill bits
were identified as 1mm, 1.5 mm, 2.0 mm, 2.5 mm and 3.0 mm. The results are
shown in Figure 4.7. The results to the 1 mm hole size AF ratio was not plotted in the
graph as the values are too lean.
4.1.5 Size of Hole at Throat Circumference
From the results, twelve hole gives most coverage of fuel at the throat. This is
true from the colours obtained from the simulation. The blue colours around the
edges are less comparing to the 8 and 10 holes. The blue colour is the location that
the fuel does not mix with the air. The occurrence of this location shows the less
coverage of the fuel at the throat cross section. With this 12 holes will be done to on
the mixer throat to enable good mixing at the throat.
All the models were capable of mixing the fuel around the mixer throat. The
induction of fuel was due to the lower pressure region at the throat. As CNG moves
from the higher pressure at the CNG inlet to lower pressure at the throat, the fuel was
virtually sucked into the mixer. It was found that the model predicted for the location
of the hole was correct and does provide adequate suction to the mixer. This was
because the mixer simulation shows fuel entering the mixer without the pressure
from the fuel inlet. It is suction that pulls the fuel in. The fuel was then seen mixing
with the air which is shown by the gradual change in colour of the simulation from
red (CNG fuel) and blue (pure air) to colour of light blue (mixed CNG and air).
56
57
510152025
010
0020
0030
0040
0050
0060
0070
0080
0090
00
3.0
mm
AF=
9.7
7
2.5
mm
2 m
m
1.5
mm
2.5
mm
2.0
mm
1.5
mm
AF=
17.
23
3 m
mD
rill
Size
Fig
ure
4.7
Eff
ect o
f ho
le s
izes
on
AF
ratio
at t
hroa
t cir
cum
fere
nce
at a
ll sp
eed
rang
e
Eng
ine
spee
d (r
pm)
AF Ratio
58
The overall trend of results show that the AF ratio becomes richer as the
speed was increased. The AF ratio also becomes rich when the area of the hole was
increased using the larger drill size. This shows that more suction of fuel occurs with
higher engine speeds and larger flow area. The limits for choosing the better hole are
drawn in the graph at AF ratio of 9.77 and 17.23.
From the graph, the area covered by hole of 1.5 mm produces a lean mixture
with most of the points above the stoichiometric limit. The reason would be the small
opening restricts the flow of fuel into the air stream. The reverse was seen happened
for the 2.5 mm and 3 mm hole as the opening area is larger. The amount of fuel was
too much and shows the overall rich condition.
From the result, the hole size of 2 mm was the best available option as the
range was found in between the very rich and stoichiometric limit. With the hole size
determined, the last stage simulation was done which is the throat size optimisation.
4.1.6 Throat Size Optimisation
Since the initial diameter was 16 mm the throat size was increased to find the
size which is in the stoichiometric range. The throat was increased slightly and each
increment was simulated to do the optimaisation. An increment of 1 mm was seen
the best for this small increment. This was because it was small enough to be
machined easily but enough produce the required resolution of results until the
stoichiometric limit.
Considering the increment, the throat size range was found to be 16 mm, 17
mm 18 mm, 19 mm, 20 mm and 21 mm. At size larger than 21 mm, the simulated
result was found too lean as it was too much above the stoichiometric limit. Due to
this the largest throat size to be simulated was set at 21 mm.
The results of simulation and actual running have errors themselves as
simulation only takes in ideal condition to calculate. As it was known that there will
be difference in simulation and experiment, the 19 mm throat diameter is the best
model to predict at this stage as the model for validation is designed next. Should
there an error occurred as predicted, the experiment is expected to put the graph with
more points at the stoichiometric range. If the 20 mm diameter was chosen it will put
the results in more lean limit and not stoichiometric.
It was shown that the 20 mm throat has the most points in the AF ratio range
specified. All the other mixer throat diameters are considered as too lean or rich for
the mixer design. Nevertheless, CFD can only predict the results and the actual
results was expected to be leaner than the simulated results.
The usage of different throat diameter showed variation in AF as seen in
Figure 4.8. The size of the mixer throat has to provide a good range of AF ratio,
which has to be near stoichimetric condition. In the graph a limit of AF of 5 % of
stoichiometric condition was drawn. For all the results a similar trend was shown.
The shape of the graph was exponential and the graph gradually straightened out.
This was due to the venturi shape that gives AF ratio at certain range. A venturi is
known not to give constant AF ratio in engines.
The throat optimisation shows that the larger sized mixer compared to the
initial throat size is a better working model at stoichiometric. The larger throat size
will give a small change in pressure drop.
Two results were found for the simulation done on the throat sizes. The
results were obtained as planned in Chapter three from the use of the CFD software
package. The first is the AF ratio and the second is the pressure drop of the mixer.
59
60
10121416182022242628
010
0020
0030
0040
00
Fig
ure
4.8
Eff
ect o
f th
roat
dia
met
er s
ize
on a
ir f
uel r
atio
5000
6000
7000
8000
9000
AF=
17.2
3A
F= +
5%
AF=
-5%
21
mm
20
mm
19
mm
18
mm
17
mm
16
mm
Eng
ine
spee
d (r
pm)
Thr
oat
Size
AF Ratio
Fuel suction was the most important aspect of the mixer and the main
operation of the mixer was based on the lower suction at the throat. Due to this, the
19 mm throat predicted is having a balance of low pressure drop to suction compared
to other throat diameters simulated. The pressure drop is around 4230 Pa. The small
pressure drop value can be neglected of giving effect to engine operation. The
predicted the last model of 19 mm will be made into the prototype.
A mixer with large pressure drop was not a good design as the possibility of
engine stalling was there due to loss of energy. This was seen in the initial model of
the mixer with 16 mm throat size. Whereas, a low pressure drop will cause reduction
of the suction pressure needed to suck in fuel. Low pressure drop was seen at large
diameters of 21 mm.
In the pressure drop results, the results are as shown in Figure 4.9. All the
graphs produced the same trend, which was the exponential increase in pressure drop
as the engine speed was increased. The pressure drop curve proves the theory from
Bernoulli, which says that as more airspeed was given to a restriction the pressure
drop would increase to the square of the velocity. Apart from this, the pressure drop
closely relates with the efficiency of the mixer. The smaller throat uses more energy
of the flow to overcome the restriction while the larger throat uses lesser energy. The
energy comes from the engine or in other words, the existence of pressure drop
reduces some efficiency from the engine. The mixer will always produce pressure
drop as the restriction is needed to create the suction to induce fuel to the throat.
61
62
0
1000
2000
3000
4000
5000
6000
7000
8000
9000
1000
0
010
0020
0030
0040
0050
0060
0070
0080
0090
00
21
mm
20
mm
19
mm
18
mm
17
mm
16
mm
Thr
oat S
ize
Fig
ure
4.9
Sim
ulat
ion
pres
sure
dro
p du
e to
dif
fere
nt th
roat
siz
e at
all
engi
ne s
peed
Eng
ine
spee
d (r
pm)
Pressure (Pa)
63
4.2 Prototyping the Mixer
Most of the dimension of the mixer was obtained from simulation. These are
can be summarised as below:
1. Symmetrical inlet and outlet angle of 60°
2. The location of the fuel hole at 24 mm from the inlet
3. There will be 12 holes at the throat circumference
4. A 2 mm drill size was chosen to be used to make the holes at the
throat
5. The throat size was finalised at 19 mm from CFD simulations.
After CFD simulation, the optimised model was made into two prototypes.
The prototypes were made using two different materials which are:
1) Perspex prototype
2) Aluminium prototype
For the first prototype, material was chosen to make a transparent mixer.
Figure 4.10 shows the mixer which was fabricated using Perspex. The Perspex
prototype was used to validate the mixing of gases in the mixer.
Figure 4.10 Perspex model for flow testing
64
Using the same dimensions, an aluminium model was made. The properties
for the chosen material are shown in Appendix E. The fabricated mixer was shown in
Figure 4.11 in assembled view and the detailed view of all the components are shown
in Figure 4.12.
Figure 4.11 Assembled view of Aluminium mixer
Throat
Mixeroutlet
CNG inlet
Mixerinlet
Figure 4.12 Components of Aluminium mixer
65
4.3 Validating the Mixer Design
The main concern of the experiments was to get the pressure drop and air fuel
ratio values and compared these values with results of CFD. Three sets of
experiments were conducted with the prototype of the device:-
i) Smoke mixing in Perspex prototype
ii) Air fuel ratio testing of mixer
iii) Pressure drop testing of mixer
4.3.1 Smoke Mixing in Perspex Prototype
This was the first test done on the mixer. The transparent mixer shows the
accumulation of smoke in the mixer throat due to suction at the CNG inlet. Only
three engine speeds were used. They were 1000 rpm, 2000 rpm and 3000 rpm. The
reason for this was that at higher speeds, the smoke could not be seen as it was
diluted in the air stream.
Figure in Appendix C and Figure 4.13 and shows the pictures that were taken
of the smoke and CFD simulation results. From here the difference between the
simulated condition and the experimental results at the applied engine air speed was
analysed.
66
1000 rpm
2000 rpm
3000 rpm
Figure 4.13 Simulation of smoke at 1000 rpm, 2000 rpm and 3000 rpm air speed
Most of the results obtained were in the range near the stoichiometric value.
This was as predicted from the simulation. Some results were scattered in the plot.
From the experimental results, a best line was drawn onto the graph. This was based
on an average of the points obtained from the experiment. This line was compared to
the simulation results.
After the relevant similitude analysis was done to the results obtained from
the gaseous flow meter, the results were compared to simulation results. Overall, the
simulation result gives a richer value compared to experiments for the AF ratio tests.
This was shown in Figure 4.14. From the results, a maximum of 4.79% AF ratio
error was calculated using the mixer. This was obtained by averaging all the errors
obtained from comparing the points of simulation and experiments.
4.3.2 AF Ratio Testing of Mixer
Based on the colour contour, simulation results show a flow pattern as air
flow speed increases. Mixing of fuel was more at high speeds in the mixer outlet; the
light blue colour shows this mixing. In the experiment, the white colour intensity of
the smoke leaving the mixer (mixed flow) was lesser than in the CNG inlet pipes.
This shows that the air and smoke mixing. Some fuel was divided into two by the
rapid air movement causing the flow to be pushed near to the wall. This was true as
the experiment had smoke taking shape from the mixer walls when leaving the
mixer. The simulation validation proves that the software can be used to analyse the
flow field qualitatively.
67
68
141516171819202122
010
0020
0030
0040
0050
0060
0070
0080
0090
00
AF=
-5%
AF=
17.
23
AF=
+5%
E
xper
imen
t
Sim
ulat
ion
Fig
ure
4.14
Exp
erim
ent a
nd s
imul
atio
n re
sults
of
AF
ratio
Eng
ine
Spee
d (r
pm)
AF ratio
69
The difference of simulation to experiment occurred due to ideal conditions
assumed by the simulation. In real condition, some of these ideal conditions would
not occur. One of these assumptions was that the atmospheric pressure at the CNG
inlet hose was not exactly at the atmospheric condition. The reason being the hose
length connecting the gaseous flow meter has given some restriction in the fuel flow.
From the results of AF ratio in Figure 4.14, it shows that the mixer was
capable of giving an overall stoichiometric AF ratio. The 19 mm mixer was slightly
leaner at low speeds. This condition was not seen very critical for the mixer as in real
life the bi-fuel kit was used to compensate the idle condition until 2000 rpm by
adjustments to the pressure regulator (Landirenzo, 2003).
In real operation, the mixer was estimated to operate within the city driving
cycle at conventional engine speeds, which rarely exceeds 6000 rpm. From the graph
in Figure 4.14, at speeds exceeding 6000 rpm, the mixer is providing a rich mixture
for the engine to combust. The operation speed of 6000 rpm can be neglected as the
engine rarely achieves this high speed in normal driving.
4.3.3 Pressure Drop Testing of Mixer
The test results were obtained with the flow test rig using the Aluminium
prototype and the digital manometer. The tests were done from the same flow rates
that were used in the simulation. The overall pressure drop was obtained by taking
the results shown in the digital manometer.
In the experiment, the result does follow the trend of increasing pressure drop
as seen in the simulations. The overall simulation results were giving higher pressure
drop than experimental data as shown in Figure 4.15. A best fit line was again drawn
As seen by the results the points plotted for the experimental reading are
fluctuating more at higher speeds. The reason was that the higher rotation speed of
the blower uses more electric power. The increased electric power usage also
amplifies the small errors producing more fluctuations to the results.
The reason for the difference was that the simulation ideal assumption did not
actually occur at real condition. During the experiment, the blower causes the
pressure drop reading to fluctuate slightly. The fluctuation was seen when the
pressure reading were taken using the manometer. The error was caused by small
fluctuation from the electric power supply, which was unavoidable in a high
electricity usage facility such as the ADC lab.
to see the errors of simulation and experiments. A maximum 4.96% error in average
of all points was found while comparing simulation and experiment results.
70
71
0
1000
2000
3000
4000
5000
6000
010
0020
0030
0040
0050
0060
0070
0080
0090
00
E
xper
imen
t
Sim
ulat
ion
Fig
ure
4.15
Sim
ulat
ions
and
exp
erim
ent p
ress
ure
drop
Eng
ine
spee
d (r
pm)
Pressure (Pa)
CHAPTER 5
CONCLUSION AND RECOMMENDATION
5.1 Conclusion
In the process of designing the CNG mixer, it can be deduced that the design
has reached the objectives set. A venturi burner type of CNG mixer was designed
and fabricated for a 150cc two stroke engine application. A mixer capable of
providing a stoichiometric AF ratio condition and low pressure drop was created.
CFD was a good tool to analyse the mixer and is a form of cost saving. From
the simulation, dimensions of the design were finalised. A conceptual design was
obtained from CFD analysis and developed into the prototype of the final product.
The two validations of pressure drop and AF ratio showed the accuracy of the
simulation results in real life application. Validation of the mixer by comparing the
CFD result and experimental results showed some errors. The errors were minimised
by doing a repetition of the experiments. The overall pressure drop and AF ratio had
recorded errors of 4.96% and 4.79% for the validation of air fuel ratio and the
validation of pressure drop. A less than 5% error was considered as a good prediction
that the simulations has achieved. The experiments proved that the flow did follow
similar trend as in the simulations. With this design process via simulation, the mixer
was found to have a general specification as shown in Table 5.1.
73
Table 5.1 Specification of the mixer designed
Length 60 mm
Inlet and outlet angles 60°
Location of hole 24 mm from inlet
Number of hole 12
Size of hole 2 mm
Size of throat 19 mm
Maximum pressure drop 4400 Pa
AF ratio range 5% 17.23 (Stoichiometric)
5.2 Recommendation
A better design of the mixer would follow a venturi shape. This was because
the shape does provide lesser losses in term of pressure drop. The assumption of
symmetrical inlet and outlet angles was done to minimise the length to fit the mixer
to the engine. A venturi specification can be followed if there are more spaces to fit
the mixer in the two stroke engine.
To get a more accurate test, a direct CNG test can be done by taking the
extreme safety precautions when using the fuel. A design of spark timing and heating
unit can be used to give timing reading and heated water to the bi-fuel conversion kit.
Actual test in engine would be the ultimate test that should be done in order
to validate the mixer further. This validation should be done by proper mounting and
sealing of the mixer to the engine. The air flow and fuel flow to the engine must be
measured directly. Other parameters such as fuel consumption, power and efficiency
compared to petrol usage can also be focused in actual engine test.
REFERENCES
Andreas N. Alexandrou (2001). Principles of Fluid Mechanics. Prentice Hall. New
Jersey.
Baert R. S. G., Beckman D. E., Veen A. (1999). Efficient EGR technology for future
HD diesel engine emission targets. TNO Road Vehicles Research Institute.
SAE 1999-01-0837.
Bryan Willson. (2002). Direct Injection as a Retrofit Strategy for Reducing
Emissions from 2-Stroke Cycle Engines in Asia. Hong Kong.
Ferguson, C.R (2001). Internal Combustion Engines- Applied Thermo-sciences. John
Wiley & Sons. Canada.
Gan L.M., (2003). Design and Development of Two Stroke Engine Using Blower
Mechanism. UTM, Thesis.
Gas Malaysia Sdn. Bhd. (2003). Natural Gas in Malaysia. Gas Malaysia
Heywood J.B (1988). Internal Combustion Engines Fundamentals, Mc Graw Hill
International Edition. Automotive Technologies Series
Jitendra (Jitu) Shah, N.Harshadeep (2001), Urban Pollution from Two Stroke Engine
Vehicles in Asia, Regional Workshop on Reduction of Emissions from 2-3
Wheelers, September 5-7, 2001– Hanoi, Vietnam.
Landirenzo, (2003). TN-SIC CNG Regulators. Installation Manual. Landirenzo
S.p.A. Italy
Lenz, H.P, (1992). Mixture Formation in Spark-Ignition Engines. SAE Inc. New
York.
Luiz Henrique Borges, Carlos Hollnagel and Wilson Muraro. (1996). Development of
Mercedes-Benz Natural Gas Engine M 366 LAG with a Lean Burn System.
SAE Brasil 1996. 962378 E
Maxwell T.T. and Jones J.C. (1995). Alternative Fuels: Emissions, Economics and
Performance. USA Society of Automotive Engineers: SAE Inc.
75
Mardani Ali Sera, Rosli Abu Bakar, Sin Kwan Leong. (2003). CNG Engine
Performance Improvement Strategy through Advanced Intake System.
Universiti Teknologi Malaysia. JSAE 20030229. SAE 2001-01-1937. Japan.
Mikio Furuyama, Bo Yan Xu. (1998). Mixing Flow Phenomena of Natural Gas and
Air in the Mixer of a CNG Vehicle. SAE 981391. Chiba University. Japan.
Mohamed Maurie Bundu. (1998). Investigation of the Performance of A Spark
Ignition Engine with Gaseous Fuels. Dalhouse University. Canada
Poulton M.L. (1994). Alternative Fuels for Road Vehicles. Computational Mechanics
Publications. Southamton. UK and Boston. USA. Pg 99-121.
Rosli Abu Bakar, Azhar Abdul Aziz and Mardani Ali Sera. (2002a). Effect of Air
Fuel Mixer Design on Engine Performance and Exhaust Emission Of A
CNG Fuelled Vehicles, 2nd World Engineering Congress Sarawak,
Malaysia,22-25 July 2002
Rosli Abu Bakar, Mardani Ali Sera, Sin Kwan Leong. (2002b). Design and
Development of New Compressed Natural Gas (CNG) Engine. IRPA Vot
72351. UTM.
Rosli Abu Bakar, Devarajan Ramasamy, Gan Leong Ming. (2004). Design of
Compressed Natural Gas (CNG) Mixer Using Computational Fluid
Dynamics. 2nd BSME-ASME International Conference on Thermal
Engineering. 2-4 January 2004. Dhaka
Rosli Abu Bakar, Devarajan Ramasamy, Chiew Chen Wee, (2003). Effects of Port
Sizes in Scavenging Process on New Two-Stroke Engine Using Numerical
Analysis. 3rd International Conference on Numerical Analysis in Engineering,
Batam View Beach Resort, 13-15 March.
Sierra Instruments, (1994). Top-Trak Mass Flow Meters. Instruction manual.
California. USA.
Taib Iskandar Mohamad, Mark Jermy, Matthew Harrison, (2003).Direct Injection of
Compressed Natural Gas in Spark Ignition Engines. ICAST 2003.
Willard W. Pulkrabek, (1997). Engineering Fundamentals of the Internal
Combustion Engine. Prentice Hall.
Yeap Beng Hi, Azeman Mustafa, Zulkefli Yaacob. (2002). Computational
Investigation of Air-Fuel Mixing System for Natural Gas Powered
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76
ISBN 983-52-0244-3
Yusoff Ali and Zailani Muhammad (2003). The Issues Promotion of the Use of
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Appendix A
Thesis Gantt Chart
MO
NT
HT
ASK
DIV
ISIO
N
1- J
2- F
3- M
4- A
5- M
6- J
7- J
8- A
9- S
10-
O
11-
N
12-
D
13- J
14-
F
15-
M
16-
A
17-
M
18- J
19- J
20-
A
21-
S
22-
O
23-
N
24-
D
1L
ITE
RA
TU
RE
RE
VIE
W:
Tw
ost
roke
eng
ine,
CN
G a
s fu
el, b
i fue
lkits
,m
ixer
oper
atio
ns, a
nd p
ress
ure
in v
entu
ri.
2M
IXE
R D
ESI
GN
:A
ir r
equi
rem
ents
of
two
stro
kes,
con
cept
mod
el, t
hroa
t siz
ing.
3C
FD
Ana
lysi
s: M
esh
stud
y, f
low
sim
ulat
ion,
air
fue
lrat
io a
naly
sis
and
pres
sure
drop
ana
lysi
s in
CFD
4F
AB
RIC
AT
ION
PR
OC
ESS
: Pr
otot
ype,
tech
nica
l dra
win
gs, m
achi
ning
, fi
nish
ing
and
asse
mbl
y.
5T
EST
RIG
: T
echn
ical
dra
win
gs,
oper
atio
n of
blo
wer
,ada
pter
des
ign
for
mix
er f
ittin
g
6T
EST
ING
:A
ssem
bly
of c
ompo
nent
s,A
Fra
tiote
stin
g, p
ress
ure
drop
test
ing
7R
EP
OR
T W
RIT
ING
: C
ompi
ling
pape
rs, t
ypin
g, e
ditin
g an
d bi
ndin
g.
The
pro
ject
is c
arri
ed o
ut o
n a
two
year
(24
mon
th)
tim
e pe
riod
78
Appendix B
CFD Analysis
80
B1 Inlet and Outlet Angle Analysis
Figure B1 Angle 60 °
Figure B2 Angle 70 °
Figure B4 Angle 90 °
Figure B3 Angle 80 °
81
B2
Hol
e Si
ze a
t cir
cum
fere
nce
sim
ulat
ion
seen
at 8
000
rpm
whi
ch is
the
high
est s
uctio
n th
at w
ill b
e gi
ven
by th
e m
ixer
1mm
1.5
mm
Fig
ure
B5
1 m
m a
nd 1
.5 m
m h
ole
size
82
2 m
m
2.5
mm
Fig
ure
B6
2 m
m a
nd 2
.5 m
m h
ole
size
83
Fig
ure
B7
3 m
m h
ole
size
84
B3
Thr
oat
AF
rat
io s
imul
atio
n
1000
rpm
2000
rpm
Fig
ure
B8
16 m
m th
roat
siz
e at
100
0 rp
m a
nd 2
000
rpm
85
3000
rpm
4000
rpm
Fig
ure
B9
16 m
m th
roat
siz
e at
300
0 rp
m a
nd 4
000
rpm
86
5000
rpm
6000
rpm
Fig
ure
B10
16
mm
thro
at s
ize
at 5
000
rpm
and
600
0 rp
m
87
7000
rpm
8000
rpm
Fig
ure
B11
16
mm
thro
at s
ize
at 7
000
rpm
and
800
0 rp
m
88
1000
rpm
2000
rpm
Fig
ure
B12
17
mm
thro
at s
ize
at 1
000
rpm
and
200
0 rp
m
89
3000
rpm
4000
rpm
Fig
ure
B13
17
mm
thro
at s
ize
at 3
000
rpm
and
400
0 rp
m
90
5000
rpm
6000
rpm
Fig
ure
B14
17
mm
thro
at s
ize
at 5
000
rpm
and
600
0 rp
m
91
7000
rpm
8000
rpm
Fig
ure
B15
17
mm
thro
at s
ize
at 7
000
rpm
and
800
0 rp
m
92
1000
rpm
2000
rpm
Fig
ure
B16
18
mm
thro
at s
ize
at 1
000
rpm
and
200
0 rp
m
93
3000
rpm
4000
rpm
Fig
ure
B17
18
mm
thro
at s
ize
at 3
000
rpm
and
400
0 rp
m
94
5000
rpm
6000
rpm
Fig
ure
B18
18
mm
thro
at s
ize
at 5
000
rpm
and
600
0 rp
m
95
7000
rpm
8000
rpm
Fig
ure
B19
18
mm
thro
at s
ize
at 7
000
rpm
and
800
0 rp
m
96
1000
rpm
2000
rpm
Fig
ure
B20
19
mm
thro
at s
ize
at 1
000
rpm
and
200
0 rp
m
97
3000
rpm
4000
rpm
Fig
ure
B21
19
mm
thro
at s
ize
at 3
000
rpm
and
400
0 rp
m
98
5000
rpm
6000
rpm
Fig
ure
B22
19
mm
thro
at s
ize
at 5
000
rpm
and
600
0 rp
m
99
7000
rpm
8000
rpm
Fig
ure
B23
19
mm
thro
at s
ize
at 7
000
rpm
and
800
0 rp
m
100
1000
rpm
2000
rpm
Fig
ure
B24
20
mm
thro
at s
ize
at 1
000
rpm
and
200
0 rp
m
101
3000
rpm
4000
rpm
Fig
ure
B25
20
mm
thro
at s
ize
at 3
000
rpm
and
400
0 rp
m
102
5000
rpm
6000
rpm
Fig
ure
B26
20
mm
thro
at s
ize
at 5
000
rpm
and
600
0 rp
m
103
7000
rpm
8000
rpm
Fig
ure
B27
20
mm
thro
at s
ize
at 7
000
rpm
and
800
0 rp
m
104
1000
rpm
2000
rpm
Fig
ure
B28
21
mm
thro
at s
ize
at 1
000
rpm
and
200
0 rp
m
105
3000
rpm
4000
rpm
Fig
ure
B29
21
mm
thro
at s
ize
at 3
000
rpm
and
400
0 rp
m
106
5000
rpm
6000
rpm
Fig
ure
B30
21
mm
thro
at s
ize
at 5
000
rpm
and
600
0 rp
m
107
108
8000
rpm
Fig
ure
B31
21
mm
thro
at s
ize
at 7
000
rpm
and
800
0 rp
m
7000
rpm
Appendix C
Apparatus and Experiments
110
Fig
ure
C1
Flow
Tes
t Rig
Ele
ctri
cal D
iagr
am
111
Figure C2 Laminar Flow Element
Figure C3 Digital Manometer (DP-Calc)
112
Figure C4 Sierra Top Trax Flow Meter
Figure C5 Pressure Difference At Inlet And Outlet Setup
113
Figure C6 Digital Manometer used to measure pressure difference, Sierra Flow
meter used to measure flow rate
Figure C7 Gas Connection
114
Figure C8 Flow test rig in suction condition
Figure C9 Flow test rig in blowing condition
115
G
21 3 414
B 5Y
No Part No Part1 3 Stage Pressure
Regulator8 Spark Timing Advance Processor
(STAP)2 Pressure Gauge 9 Battery 12VDC3 Gasoline Solenoid Valve 10 Ignition Key 4 Refueling Valve 11 Gasoline-CNG Fuel Switch 5 High Pressure Pipe 12 Mixer6 CNG Tank 13 Power Valve 7 Engine Ignition Coil 14 Low Pressure Pipe
Color Code R-Red wire
G-Green wire
B-Blue wire
Br-Brown wire
W-White wire
Y- Yellow wire
8
R Br
G
G
R12
13
11
10
9 7
6
-+
STAP
W G
Figure C10 Schematic diagram of bi-fuel system
116
Figure C11 Smoke flow at 1000 rpm
Figure C12 Smoke flow at 2000 rpm
Figure C13 Smoke flow at 3000 rpm
Appendix D
Technical Drawings
A A
38
44
34
61
67
0
10
70
80
A-A
(1
: 1)
DRA
WN
APP
V'D
UN
LESS
OTH
ERW
ISE
SPEC
IFIE
D:
DIM
ENSI
ON
S A
REIN
MIL
LIM
ETER
SFI
NIS
H:
DEB
UR A
ND
BREA
K SH
ARP
EDG
ES
NA
ME
SIG
NA
TURE
DA
TE
MA
TERI
AL:
WEI
GHT
:
UNIV
ERSI
TI TE
KNO
LOG
I MA
LAYS
IA
TITLE
:D
WG
NO
.:SC
ALE
:1:2
SHEE
T 1 O
F1
A4
SURF
AC
E FI
NIS
H:
TOLE
RAN
CES
:(U
NLE
SS S
PEC
IFIE
D)
LIN
EAR:
AN
GUL
AR:
QTY
.:
DEV
ARA
JAN
A/L
RA
MA
SAM
Y05
/01/
2005
Flow
Tes
tRig
Ad
apt
er
FAKU
LTI K
EJUR
UTER
AA
NM
EKA
NIK
AL
PRO
F. M
AD
YA D
R. R
OSL
I
118
720
850
1123
.50
167
93245
3.50
273.
50275
280
212
117
A90
Ben
d P
VC
Pip
e
Pers
pe
x A
da
pte
rD
igita
l Dis
pla
y U
nit
Lam
ina
r Flo
w E
lem
en
t
Ce
ntr
ifug
al B
low
er
645
1397
450
300
220
DET
AIL
ASC
ALE
1 :
5
Mix
er
DO
NO
T SC
ALE
DR
AW
ING
Flo
w T
est
Rig
SHEE
T 1
OF
1
UNLE
SS O
THER
WIS
E SP
ECIF
IED
:SC
ALE
: 1:1
WEI
GH
T:
REV
DW
G.
NO
.
ASIZ
E
TITL
E:N
AM
ED
ATE
Q.A
.
MFG
AP
PR
.
EN
G A
PP
R.
CH
EC
KE
D
DR
AW
N
FIN
ISH
MA
TER
IAL
INTE
RP
RET
GEO
MET
RIC
TOLE
RA
NC
ING
PER
:
DIM
EN
SIO
NS
AR
EIN
MIL
IME
TER
STO
LER
AN
CE
S:F
RA
CTI
ON
AL
AN
GU
LAR
: MA
CH
BE
ND
TW
O P
LAC
E D
EC
IMA
LT
HR
EE
PLA
CE
DE
CIM
AL
PRO
PRIE
TARY
AN
D C
ON
FID
ENTIA
L
THE
INFO
RM
ATI
ON
CO
NTA
INED
IN T
HIS
DR
AW
ING
IS T
HE
SOLE
PR
OP
ERTY
OF
UN
IVER
SITI
TEK
NO
LOG
IMA
LAY
SIA
(U
TM).
AN
YR
EPR
OD
UC
TIO
N IN
PA
RT
OR
AS
AW
HO
LE W
ITH
OU
T TH
E W
RIT
TEN
PER
MIS
SIO
NO
F U
TM IS
PR
OH
IBIT
ED.
DE
VA
RO
SLI
119
A A
CN
G In
let
19 m
mRi
ng
32
18.0
6
32
45.9
4
60
A-A
(1 :
1)
Mix
er O
utle
tM
ixer
Inle
t
DR
AW
N
AP
PV
'D
UN
LESS
OTH
ERW
ISE
SPEC
IFIE
D:
DIM
ENSI
ON
S A
RE
INM
ILLI
MET
ERS
FIN
ISH
:D
EBU
R A
ND
BREA
K S
HA
RP
ED
GES
NA
ME
SIG
NA
TUR
ED
ATE
MA
TER
IAL:
WEI
GH
T:
UN
IVER
SITI
TEK
NO
LOG
IMA
LAYS
IA
TITL
E:D
WG
NO
.:S
CA
LE:1
:2S
HEE
T 1
OF
1
A4
SUR
FAC
E FI
NIS
H:
TOLE
RA
NC
ES:
(UN
LESS
SP
ECIF
IED
) L
INEA
R:
AN
GU
LAR
:
QTY
.:
DEV
AR
AJA
N R
AM
ASA
MY
1/9
/200
4
Alu
min
ium
Mix
er
FAKU
LTI K
EJUR
UTER
AAN
MEK
ANIK
ALA
SSO
C P
ROF
RO
SLI A
BU
BAKA
R
120
A A
37
32
30°
4
2
29
36
024
9.80
18.06
A-A
(2
: 1)
DRA
WN
APP
V'D
UNLE
SS O
THER
WIS
ESP
ECIF
IED
:D
IMEN
SIO
NS
ARE
IN M
ILLI
MET
ERS
FIN
ISH
:D
EBUR
AN
DBR
EAK
SHA
RP E
DG
ES
NA
ME
SIG
NA
TURE
DA
TE
MA
TERI
AL:
WEI
GH
T:
UNIV
ERSI
TI TE
KNO
LOG
IMA
LAYS
IA
TITLE
:D
WG
NO
.:SC
ALE
:1:1
SHEE
T1
OF
1
A4
SURF
AC
E FI
NIS
H:
TOLE
RAN
CES
:(U
NLE
SS S
PEC
IFIE
D)
LIN
EAR:
AN
GUL
AR:
QTY
.:
DEV
ARA
JAN
RA
MA
SAM
Y1/
9/20
04
Alu
min
ium
Mix
er In
let
FAKU
LTI K
EJUR
UTER
AA
N M
EKA
NIK
AL
ASS
OC
PRO
FRO
SLI A
BU B
AKA
R
121
A A
30°
32
36
37
29
36
44
M10
0
9.94
16
20
27
36.14
41.9443.9445.94
A-A
(2
: 1)
M36
Thr
ead
Leng
th 5
mm
DR
AW
N
AP
PV'
D
UNLE
SS O
THER
WIS
E SP
ECIF
IED
:D
IMEN
SIO
NS
AR
EIN
MIL
LIM
ETER
SFI
NIS
H:
DEB
UR
AN
D
BREA
K S
HA
RP
EDG
ES
NA
ME
SIG
NA
TUR
ED
ATE
MA
TER
IAL:
WEI
GH
T:
UNIV
ERSI
TI T
EKNO
LOG
I MA
LAYS
IA
TITL
E:D
WG
NO
.:S
CA
LE:1
:1S
HEE
T 1
OF
1
A4
SUR
FAC
E FI
NIS
H:
TOLE
RA
NC
ES:
(UN
LESS
SPE
CIF
IED
) L
INEA
R:
AN
GU
LAR
:
QTY
.:
DEV
AR
AJA
N R
AM
ASA
MY
1/9
/200
4
Alu
min
ium
Mix
er O
utle
t
FAKU
LTI K
EJUR
UTER
AAN
MEK
ANIK
ALA
SSO
C P
ROF
ROSL
I ABU
BA
KAR
122
A A
30°
19
29
3.50°
12 X
2
Thr
u
0
7.94
5
25
A-A
(2
: 1)
DR
AW
N
AP
PV
'D
UN
LESS
OTH
ERW
ISE
SPEC
IFIE
D:
DIM
ENSI
ON
S A
RE
INM
ILLI
MET
ERS
FIN
ISH
:D
EBU
R A
ND
BREA
K S
HA
RP
EDG
ES
NA
ME
SIG
NA
TUR
ED
ATE
MA
TER
IAL:
WEI
GH
T:
UN
IVER
SITI
TEKN
OLO
GI M
ALA
YSIA
TITL
E:D
WG
NO
.:S
CA
LE:1
:1S
HEE
T 1
OF
1
A4
SUR
FAC
E FI
NIS
H:
TOLE
RA
NC
ES:
(UN
LESS
SP
ECIF
IED
) L
INEA
R:
AN
GU
LAR
:
QTY
.:
DEV
AR
AJA
N R
AM
ASA
MY
1/9
/200
4
Alu
min
ium
19 m
m R
ing
FAKU
LTI K
EJUR
UTER
AAN
MEK
ANIK
ALA
SSO
C P
ROF
RO
SLI A
BU B
AKA
R
123
124
A A
16
18
7
M10
0
22
32
44
48
A-A
(2
: 1)
M10
Thre
adLe
ngth
4 m
m
DR
AW
N
APP
V'D
UNLE
SS O
THER
WIS
E SP
ECIF
IED
:D
IMEN
SIO
NS
AR
EIN
MIL
LIM
ETER
SFI
NIS
H:
DEB
UR
AN
DBR
EAK
SHA
RP
EDG
ES
NA
ME
SIG
NA
TUR
ED
ATE
MA
TER
IAL:
WEI
GH
T:
UN
IVER
SITI
TEKN
OLO
GIM
ALA
YSIA
TITL
E:D
WG
NO
.:S
CA
LE:1
:1S
HEE
T 1
OF
1
A4
SUR
FAC
E FI
NIS
H:
TOLE
RA
NC
ES:
(UN
LESS
SPE
CIF
IED
)LI
NEA
R:
AN
GU
LAR
:
QTY
.:
DEV
AR
AJA
N R
AM
ASA
MY
1/9
/200
4
Alu
min
ium
CN
G In
let
FAKU
LTI K
EJUR
UTER
AAN
MEK
ANIK
ALA
SSO
C P
ROF
ROSL
I ABU
BA
KAR
Appendix E
Material Selection
126
MatWeb.com, The Online Materials Database
Aluminum 6061-T8
Subcategory: 6000 Series Aluminum Alloy; Aluminum Alloy; Metal; Nonferrous Metal
Key Words: Aluminium 6061-T8; UNS A96061; ISO AlMg1SiCu, AD-33 (Russia); AA6061-T8
Component Wt. %
Al 98
Cr 0.04 - 0.35
Cu 0.15 - 0.4
Component Wt. %
Fe Max 0.7
Mg 0.8 - 1.2
Mn Max 0.15
Component Wt. %
Si 0.4 - 0.8
Ti Max 0.15
Zn Max 0.25
Material Notes:Weldability = A; Stress Corrosion Cracking Resistance = A; General CorrosionResistance = B (A = best; E = worst). General 6061 characteristics and uses: Excellentjoining characteristics, good acceptance of applied coatings. Combines relatively high strength, good workability, and high resistance to corrosion; widely available. The T8 and t9 tempers offer better chipping characteristics over the T6 temper.
Uses: Aircraft fittings, camera lens mounts, couplings, marines fittings and hardware,electrical fittings and connectors, decorative or misc. hardware, hinge pins, magnetoparts, brake pistons, hydraulic pistons, appliance fittings, valves and valve parts.
Most data provided by Alcoa.
Physical Properties Metric English Comments
Density 2.71 g/cc 0.0979 lb/in³
Mechanical Properties
Hardness, Brinell 120 120 500 kg load/10 mm ball
Hardness, Knoop 150 150 Estimated from Brinell
Hardness, Rockwell A 46.8 46.8 Estimated from Brinell
Hardness, Rockwell B 75 75 Estimated from Brinell
Hardness, Vickers 136 136 Estimated from Brinell
Tensile Strength, Ultimate Min 310 MPa Min 45000 psi
Tensile Strength, Yield Min 276 MPa Min 40000 psi
Elongation at Break 8 % 8 %
127
Modulus of Elasticity 69 GPa 10000 ksi Average of Tension andCompression. In
Aluminum alloys, thecompressive modulus is
typically 2% greater than the tensile modulus
Poisson's Ratio 0.33 0.33 Estimated from trends insimilar Al alloys.
Machinability 50 % 50 % 0-100 Scale (A=90; B=70; C=50; D=30;
E=10)
Shear Modulus 26 GPa 3770 ksi Estimated from similarAl alloys.
Shear Strength 185 MPa 26800 psi Estimated from ultimatetensile strength
Electrical Properties
Electrical Resistivity 3.7e-006 ohm-cm 3.7e-006 ohm-cm
Thermal Properties
CTE, linear 20°C 23.6 µm/m-°C 13.1 µin/in-°F average over 20-100°C
CTE, linear 250°C 25.2 µm/m-°C 14 µin/in-°F Estimated from trends insimilar Al alloys. 20-
300°C.
Heat Capacity 0.896 J/g-°C 0.214 BTU/lb-°F
Thermal Conductivity 170 W/m-K 1180 BTU-in/hr-ft²-°F
Melting Point Min 582 °C Min 1080 °F Solidus
Solidus 582 °C 1080 °F
Copyright 1996-2003 by Automation Creations, Inc. The information provided by MatWeb is intended for personal, non-commercial use. The contents, results, and technical data from this site may not be reproduced either electronically,photographically or substantively without permission from Automation Creations, Inc. No warranty, neither expressed nor implied, is given regarding the accuracy of this information. The user assumes all risk and liability in connection with the use ofinformation from MatWeb.
Appendix F
Mesh Independent Analysis
129
Mesh Independent Analysis
Before further simulation was done a mesh analysis was carried out to
determine the most accurate mesh. Firstly, the model with diameter 16 mm is meshed
in the CFD software using a very coarse mesh this is level-3 meshing in the software.
The mesh is then increased to level 4, level 5, level 6 and level 8 as shown in Table
F1 until the pressure drop is almost constant between two points in the mixer as
shown in Figure F2. The result is level 6 mesh was chosen for the simulation with
refinements as shown in Figure F1. This mesh was considered as the results were not
varying and did not take too much CPU time to calculate
Level 3 Level 4
Level 8 Level 6
Figure F1 Mesh Levels analyzed level 3, level 4, level 6 and level 8
130
0
2000
4000
6000
8000
10000
12000
14000
0 1000 2000 3000 4000 5000 6000 7000 8000
Engine Speed, rpm
Pre
ssur
e D
rop,
Pa
Level 3 Level 4 Level 6 Level 8
Figure F2 Pressure drop for simulation diameter 16 at different mesh
Table F1 Number of cells for each level for 16 mm diameter simulation
Level 3 Level 4 Level 6 Level 8
Fluid Cells 9274 25461 63312 188536
Solid Cells 12553 23009 55505 111835
Partial Cells 14063 25778 53105 100515
Total Cells 35890 74248 171922 400886
CPU TIme1596s27min
4865s1hour 30 min
17893s4 hours 50min
63365s17 hour 36 min