THE EFFECTS OF EXHAUST GAS RECIRCULATION (EGR) 
ON THE PERFORMANCE OF DIESEL ENGINE 
 
 
MAROA, Samwel Semakula, B. Tech (Automotive) 
217081370 
 
 
Supervisor: Prof. Freddie L. Inambao 
 
 
Dissertation submitted in fulfilment of the requirement for the degree of 
Master’s in Engineering (M. Sc. Eng.) 
 
 
(MECHANICAL ENGINEERING) 
 
 
School of Engineering, 
University of KwaZulu-Natal, 
Durban South Africa 
 
 
 
As the candidate’s supervisor, I agree/do not agree to the submission of this dissertation. The 
supervisor must sign all copies after deleting which is not applicable. 
 
 
PROF  F.N INAMBAO                                        
Name of Supervisor                                                Signature 
 
 
MAY 2018 
 
  
Declaration 1 –Plagiarism 
 
I, Maroa Samwel Semakula declare that: 
1. The research reported in this thesis, except where otherwise indicated is my original 
research. 
2. This thesis has not been submitted for any degree or examination at any other university. 
3. This thesis does not contain other persons’ data, pictures, graphs or other information, 
unless specifically acknowledged as being sourced from other persons. 
4. This thesis does not contain other persons' writing, unless specifically acknowledged as 
being sourced from other researchers. Where other written sources have been quoted, 
then: 
a) Their words have been re-written but the general information attributed to them 
has been referenced 
b) Where their exact words have been used, then their writing has been placed in 
italics and inside quotation marks, and referenced. 
5. This thesis does not contain text, graphics or tables copied and pasted from the internet, 
unless specifically acknowledged, and the source being detailed in the thesis and in the 
References sections. 
 
 
Signed……………………………… 
 
  
iii 
 
Declaration 2 –Publications 
 
This section presents the articles that form part and/or include the research presented in this 
thesis. 
ISI/DoHET Accredited Journals 
Maroa, S., & Inambao, F. (2017). The Effects of Exhaust Gas Recirculation on the 
Performance and Emission Characteristics of a Diesel Engine–A Critical Review. 
International Journal of Applied Engineering Research, 12(23), 13677-13689. 
http://www.ripublication.com 
 
Maroa, S., & Inambao, F. (2018). Transportation, Pollution and the Environment. 
International Journal of Applied Engineering Research, 13(6), 3187-3199. 
http://www.ripublication.com 
 
Maroa, S., & Inambao, F. (2018). The Formation, Effects and Control of Oxides of Nitrogen 
in Diesel Engines. International Journal of Applied Engineering Research, 13(6), 
3200-3209. http://www.ripublication.com 
 
Maroa, S., & Inambao, F. (2018). Influence of Exhaust Gas Recirculation and Emission 
Characteristics of a Diesel Engine Using Pyrolyzed Waste Plastic Biodiesel and 
Blends. International Journal of Applied Engineering Research, 13(10), 8321-8335. 
http://www.ripublication.com 
 
International and DoHET Accredited Conferences 
Maroa, S.S. and Inambao, F. L. (2017). The Effects of EGR on the Performance and 
Emission Characteristics of a Diesel Engine – A Critical Review Proceedings of the 
15th biennial Botswana Institution of Engineers Conference, Gaborone, Botswana: 
16th-19th October, 2017: Proceedings Being Processed by Conference Organizers. 
 
 
 
  
iv 
 
DEDICATION 
 
This work is dedicated to my late mother Selina Sibora Marwa. 
  
v 
 
ACKNOWLEDGEMENTS 
 
First and far most to the good Lord for His enabling grace and mercies, to my supervisor Prof. 
FN Inambao for his tireless efforts in molding me. It has been a humbling and learning 
experience to work with you. To my editor Dr Richard Steele for his professionalism and 
personal effort to edit this dissertation. To the LAN manager Shaun Savvy for making sure I 
got a bigger screen and other necessary computer applications to enable me to work 
comfortably. To my brother David K. Marwa for his vision, effort and tireless support. To my 
family members especially my sisters Prof. M. B Mwita and Dr I. N Marwa. They have trusted 
and encouraged me in pursuing my goals. The UKZN fraternity for the enabling environment, 
especially my fellow post-graduate students in the Green Energy Solutions Group, who helped 
me to complete this work. Thank you for not getting tired when I came calling for assistance. 
I do not want to forget my spiritual home at Umbilo Seventh-Day Adventist Church. They have 
been an example in faith and in truth. May God continue to bless as you nurture more souls in 
their spiritual journey. There are a number of people who have contributed in one way or the 
other to this work materially or morally. I am not able to mention you individually but I will 
always remain indebted to you. Thank you all. 
 
  
vi 
 
ABSTRACT 
 
The aim of this work was to study the effects of EGR on the performance of a diesel engine 
using waste plastic pyrolysis oil (WPPO) and conventional diesel (CD). WPPO was developed 
through the pyrolysis extraction method. The blends were made up of WPPO and conventional 
diesel mixed in the ratios of WPPO10, WPPO20, WPPO30, WPPO40 and WPPO100. The 
EGR % flow rate chosen was 0 % to 30 % graduated in intervals of 5 %. The lower blend ratios 
of WPPO10 and WPPO20 showed lower values of brake specific fuel consumption (BSFC) 
compared to conventional diesel values and the high blend ratio of WPPO100. The brake 
thermal energy (BTE) showed increased values for lower blend ratios of WPPO10 and 
WPPO20 of 8.35 % and 8.15 % respectively with application of an EGR % flow rate of 15 % 
compared to high blend ratios of WPPO30, WPPO40 and WPPO100. The application of EGR 
% flow rate was observed to cause no significant change in the engine brake power (BP) for 
all the test fuels used. The application of EGR % flow rate in increasing rates reduced exhaust 
gas temperature (EGT), with conventional diesel reporting 440 ⁰C at 5 % EGR flow rate and 
340 ⁰C being the lowest at 30 % EGR flow rate. The application of EGR % flow rate reduced 
the amount of hydrocarbon emissions emitted by the applied test fuels across the board. At 
EGR flows rate of 5 %, 10 %, 15 %, 20 %, 25 % and 30 %, conventional diesel had 43 ppm, 
57 ppm, 70 ppm, 82 ppm and 85 ppm respectively. As the blend ratio increased with increased 
EGR % flow rate there was an increased rate of NOX emissions. At 20 % EGR flow rate, blends 
WPPO10, WPPO2O, WPPO30, WPPO40 and WPPO100 had 591ppm, 645 ppm, 750 ppm, 
778 ppm and 851 ppm respectively compared to at the 10 % EGR flow rate where their values 
were 830 ppm, 971 ppm, 1031 ppm, 1151 ppm and 1116 ppm respectively. There was a 
significant continuous and marginal increase in the percentage of carbon emissions by volume 
as the load increased across all the test fuels irrespective of the EGR % flow rate. At 80 % 
engine load the value for WPPOB100 was 2.0 % up from 1.65 % by volume at part engine 
load, while the value of conventional diesel was 4.1 % at 80 % engine load compared to 2.95 
% by volume at part engine load. The application of EGR % flow rate increased the carbon 
dioxide emission exponentially by almost doubling the values. At 10 % EGR flow rate the 
value of conventional diesel was 3.85 % compared to WPPOB100 at 6.25 %, WPPOB10 at 
4.75 %, WPPOB20 at 4.25 %, WPPOB30 at 3.95 %, and WPPOB40 at 6.65 %.  
 
  
vii 
 
TABLE OF CONTENTS 
 
DECLARATION 1 – PLAGIARISM………………………………………………………. iii 
DECLARATION 2 – PUBLICATIONS……………………………………………………. iv 
DEDICATION………………………………………………………………………………...v 
ACKNOWLEDGEMENTS………………………………………………………………….vi 
ABSTRACT………………………………………………………………………………… vii 
TABLE OF CONTENTS……………………………………………………………………viii 
LIST OF TABLES……………………………………………………………………………xii 
LIST OF FIGURES………………………………………………………………………… xiv 
LIST OF APPENDICES……………………………………………………………………xvii 
ABBREVIATIONS AND ACRONYMS…………………………………………………. xviii 
SUBSCRIPTS SYMBOLS………………………………………………………………… xxii 
GREEK SYMBOLS………………………………………………………………………. xxiv 
GENERAL SYMBOLS…………………………………………………………………… xxv 
CHAPTER ONE……………………………………………………………………………….1 
INTRODUCTION………………………………………………………………………...…...1 
1.1 Introduction and Background………………………………………………………………1 
1.2 Aims and Objectives……………………………………………………………………….6 
1.3 Dissertation Outline……………………………………………………………………….6 
1.4 Research Motivation……………………………………………………………………….8 
References……………………………………………………………………………………10 
CHAPTER TWO………………………………………………………………………...…...12 
LITERATURE REVIEW…………………………………………………………………….12 
2.1 Introduction………………………………………………………………………………12 
viii 
 
2.3 The Effect of EGR On the Role of Exergy Destruction………………………………….22 
2.4 The Effects of EGR On Emission and Combustion Characteristics of Biodiesel……….24 
2.5 Effects of EGR on Performance, Combustion and Emission Characteristics of Diesel 
Engines (DI/IDI) …………………………………………………………………………….32 
2.6 Effects of EGR On CNG and GTL On Performance Combustion and Emissions……….41 
2.7 Effects of EGR on Temperature and Emissions………………………………………….43 
2.8 Effects of EGR On Performance, Combustion and Emissions on Biodiesel Fuels………50 
2.9 The Effects of EGR On Particle Formation and Oil Degradation…………………………60 
2.10 Effects of Emission Gases on Public Health…………………………………………….62 
References……………………………………………………………………………………67 
CHAPTER THREE………………………………………………………………………….73 
METHODOLOGY…………………………………………………………………………...73 
3.1 Introduction………………………………………………………………………………73 
3.2 Numerical Equations Model……………………………………………………………...73 
3.3 Definitions of Numerical Equations………………………………………………………73 
3.3.1 First Law and Conservation of Mass……………………………………………………73 
3.3.2 The Second Law and Conservation of Species…………………………………………74 
3.3.3 The Conservation of Energy……………………………………………………………74 
3.3.4 Air/Fuel Ratio………………………………………………………………………….75 
3.3.5 Brake Thermal Efficiency (BTE)……………………………………………………….75 
3.3.6 Brake Power (BP)………………………………………………………………………76 
3.3.7 Brake Specific Fuel Consumption (BSFC)…………………………………………….76 
3.3.8 The Volumetric Efficiency (VE)……………………………………………………….77 
3.4 The EGR Valve Model Equations……………………………………………………….77 
3.4.1 Modelling the EGR Valve Using the Orifice Equation ……………………………….81 
ix 
 
3.4.2 The Throttle Valve Equation……………………………………………………………83 
3.4.3 Fresh Air Fraction………………………………………………………………………84 
3.5 The EGR Cooler and The Exchange of Heat in the Exhaust Manifold……………………85 
3.5.1 The Axial Heat Transfers in the Exhaust Manifold…………………………………….89 
3.5.2 The Radial Heat Exchange in the Exhaust Manifold……………………………………91 
3.5.3 The Calculation of Coefficients of Heat Transfer………………………………………92 
References……………………………………………………………………………………94 
CHAPTER FOUR……………………………………………………………………………95 
THE EFFECTS OF EXHAUST GAS RECIRCULATION ON THE PERFORMANCE AND 
EMISSION CHARACTERISTICS OF A DIESEL ENGINE – A CRITICAL REVIEW……95 
Journal article………………………………………………………………………………...95 
CHAPTER FIVE……………………………………………………………………………108 
TRANSPORTATION, POLLUTION AND THE ENVIRONMENT………………………108 
Journal article……………………………………………………………………………….108 
CHAPTER SIX…………………………………………………………………………….122 
THE FORMATION, EFFECTS AND CONTROL OF OXIDES OF NITROGEN IN DIESEL 
ENGINES………………………………………………………………………………….122 
Journal article………………………………………………………………………………122 
CHAPTER SEVEN…………………………………………………………………………133 
Influence of Exhaust Gas Recirculation and Emission Characteristics of a Diesel Engine Using 
Pyrolyzed Waste Plastic Biodiesel and Blends……………………………………………...133 
Journal article……………………………………………………………………………….133 
CHAPTER EIGHT………………………………………………………………………….149 
THE EXPERIMENTAL SET-UP AND EQUIPMENT……………………………………149 
8.1 Introduction…………………………………………………………………………….149 
x 
 
8.3 Experimental Apparatus and Equipment……………………………………………….149 
8.3 Waste Plastic Preparation and the Conversion Process………………………………...150 
8.4 Physical Properties of Waste Plastic Pyrolysis Oil (WPPO) Sample……………………152 
8.5 Experimental Procedure…………………………………………………………………154 
8.6 Performance Measurements…………………………………………………………….156 
8.7 Emission Measurements………………………………………………………………...158 
8.8 Analysis of Error and Percentage Uncertainties…………………………………………161 
References……………………………………………………………………………...…...163 
CHAPTER NINE…………………………………………………………………………...164 
RESULTS AND DISCUSSION…………………………………………………………….164 
9.1 Brake Specific Fuel Consumption (BSFC)……………………………………………...164 
9.2 Brake Thermal Efficiency (BTE)…………………………………………………...…...165 
9.3 Brake Power (BP)……………………………………………………………………….167 
9.4 Exhaust Gas Temperature (EGT)……………………………………………………….169 
9.5 NOX Emissions………………………………………………………………………….170 
9.6 Hydrocarbon (HC) Emissions………………………………………………………….172 
9.7 Carbon Monoxide (CO) Emission………………………………………………………173 
9.8 Carbon Dioxide (CO2) Emissions ………………………………………………………175 
9.9 Smoke Emissions (Opacity)…………………………………………………………….177 
References……………………………………………………………………………...…...179 
CHAPTER TEN…………………………………………………………………………….181 
CONCLUSION AND RECOMMENDATIONS…………………………………………...181 
10.1 Conclusion…………………………………………………………………………….181 
10.2 Future Recommendations……………………………………………………………...183 
xi 
 
 
LIST OF TABLES 
 
Table 1.1: LTC research strategies…………………………………………………………….5 
Table 2.1: Baseline diesel test fuel and various pentanol blends of test fuels and their properties 
adapted from (L. Li, Wang, Wang, & Liu, 2015) …………………………………………….34 
Table 2.2: Comparison between the mean oil film thickness ( 𝒉𝒉𝒉𝒉𝒉𝒉𝒉𝒉) and the mean carbon 
black diameter (Hu et al., 2013)………………………………………………………………61 
Table 2.3: The main sources of pollution, types of pollutants, the environmental and health 
effects of these pollutants (Kim et al., 2017) …………………………………………………63 
Table 8.1: The engine specifications, position value and their type…………………………150 
Table 8.2: The test fuel properties, their units of measurement, standard methods of testing and 
the values for conventional diesel in comparison to the values of waste plastic pyrolysis 
oil……………………………………………………………………………………………153 
Table 8.3: The chemical composition of the waste plastic oil from a pyrolysis plant process: 
the carbon chains and percentage compositions in the test fuel, obtained from a GC-MS 
laboratory analysis report……………………………………………………………………154 
Table 8.4: BSFC values in relation to engine load and various test fuel blends of WPPO and 
conventional diesel (CD)……………………………………………………………………156 
Table 8.5: BSFC kg/kW. hr. values under the effects of EGR % flow rate, full engine load with 
different fuel blends of WPPO and conventional diesel (CD)………………………………156 
Table 8.6: BTE % values, at full engine load with blends of WPPO and conventional diesel 
(CD)……………………………...………………………………………………………….157 
Table 8.7: BTE % values, under load with EGR % flow rate application with blends of WPPO 
and conventional diesel (CD)……………………………………………………………….157 
Table 8.8: BP values under engine load with blends of WPPO and conventional diesel 
(CD)…………………………………………………………………………………………157 
xii 
 
Table 8.9: Temperature values for exhaust gas under different WPPO blends and conventional 
diesel (CD) with the application of EGR % flow rate at full engine load………………….158 
Table 8.10: NOX emissions (ppm) values versus engine load with different WPPO blended 
fuels and conventional diesel (CD) under the influence of EGR % flow rate application….158 
Table 8.11: NOX emissions (ppm) values versus full engine load with different blends of WPPO 
and conventional diesel (CD)……………………………………………………………….159 
Table 8.12: UHC emissions (ppm) values under engine load with different fuel blends of 
WPPO and conventional diesel (CD) under the effects of EGR % flow rate application….159 
Table 8.13: CO emissions % values versus engine load with the different fuel blends of WPPO 
and conventional diesel CD) ……………………………………………………………….159 
Table 8.14: CO emissions % values versus EGR % flow rate with different fuel blends of 
WPPO and conventional diesel (CD) ………………………………………………………160 
Table 8.15: CO2 emissions % values for various blends of WPPO and conventional diesel (CD) 
under engine load……………………………………………………………………………160 
Table 8.16: CO2 % values for various blends of WPPO and conventional diesel fuel under 
engine load with the application of EGR % flow rate………………………………………160 
Table 8.17: Smoke emissions or opacity % values at full engine load, with different blends of 
WPPO and conventional diesel (CD) under the effects of EGR % flow rate application……160 
Table 8.18: Instruments used for measurements and data collection, their measuring range, 
accuracies and percentages of inaccuracies, as calculated from Equation 2……………….162 
 
 
  
xiii 
 
LIST OF FIGURES 
 
Figure 1.1: Variation of world energy consumption percentage shares for each source of 
energy [21] ........................................................................................................................... 9 
Figure 2.1: Variation of NOX emissions with oxygen mass fraction in hot and cool EGR 
modes [2] ........................................................................................................................... 13 
Figure 2.2: EGR loop – high pressure loop (HPL) .................................................................. 18 
Figure 2.3: EGR loop – low pressure loop (LPL) [14] ............................................................ 18 
Figure 2.4: The LSB schematic set-up system and the main components ............................... 22 
Figure 2.5: The geometrical parameters of the LSB unit as used in the experimental set-up . 22 
Figure 2.6: Variation of temperature versus flame radius length of lifted and attached flame 24 
Figure 2.7: Variation of temperaure versus flame radius length and flame progression of lifted 
and attached flame ............................................................................................................. 24 
Figure 2.8: Variation of the temperature range for the two types flame progression regimes 
shown in Figures 2.7 and 2.8. ............................................................................................ 24 
Figure 2.9: UHC vs Load ......................................................................................................... 29 
Figure 2.10: Variation of exhaust gas temperature with brake mean effective pressure under 
different diesel test fuel with application of EGR % rate flow [3] .................................... 32 
Figure 2.11: Variation of burnt fraction mass with crank angle at different varying engine 
load conditions [47] ........................................................................................................... 33 
Figure 2.12: Variation of in-cylinder pressure with crank angle change in relation to heat 
release rate [49] .................................................................................................................. 35 
Figure 2.13: ISFC VS IMEP .................................................................................................... 37 
Figure 2.14: ITE VS IMEP ...................................................................................................... 37 
Figure 2.15: BSFC VS EGR % flow rate ................................................................................ 38 
Figure 2.16: Figure 17 SFC VS Brake Power.......................................................................... 39 
Figure 2.17: Figure 18 BTE VS Brake Power ......................................................................... 39 
Figure 2.18: ITE %, CO (%) and HC (ppm) VS EGR % flow rate application ...................... 41 
Figure 2.19: Carbon Dioxide Vs Load ..................................................................................... 48 
Figure 2.20: EGT VS Load ...................................................................................................... 49 
Figure 2.21: EGT VS diesel and palm biodiesel with and without EGR flow rate application
 ............................................................................................................................................ 51 
Figure 2.22: BSFC vs different test fuels with and without application of EGR .................... 51 
xiv 
 
Figure 2.23: ROPR VS Crank Angle ....................................................................................... 52 
Figure 2.24: Equivalence ratio VS Engine Full Load (diesel fuel and compressed natural gas)
 ............................................................................................................................................ 53 
Figure 2.25: EGR schematic for the mass fraction method of calculating the EGR % flow rate 
in a system [72] .................................................................................................................. 56 
Figure 2.26: NOX VS EGR % with different blends of pentanol and diesel ........................... 57 
Figure 2.27: BTE VS EGR % flow rates ................................................................................. 57 
Figure 2.28: Figure 29 BTE VS Brake Power ......................................................................... 58 
Figure 2.29: HRR VS Crank Angle ......................................................................................... 59 
Figure 2.30: Effects of EGR on (a)specific emissions and (b) mass fractions of carbonaceous 
components [79] ................................................................................................................. 62 
Figure 3.1: The EGR schematic loop nomenclature ................................................................ 78 
Figure 3.2: The EGR cooler heat balance definitions and nomenclature ................................ 86 
Figure 3.3: Axial and radial components of the heat transfer model of the exhaust pipe and its 
nomenclature ...................................................................................................................... 89 
Figure 3.4: Axial heat resistance model for the exhaust pipe with its nomenclature .............. 90 
Figure 3.5: Modeled radial heat resistance of the exhaust pipe ............................................... 91 
Figure 8.1: Experimental test engine schematic set-up diagram ........................................... 149 
Figure 8.2: schematic of the lab set-up of the fabricated and modified exhaust recirculation 
mechanism to conform to the engine experimental set-up in Figure 8.1 ......................... 150 
Figure 8.3: Pyrolysis flow chart ............................................................................................. 152 
Figure 8.4: The distillate samples from the waste plastic pyrolysis oil samples ................... 153 
Figure 9.1: BSFC versus EGR % rate flow under full engine load conditions ..................... 164 
Figure 9.2: Variation of BTE versus engine load % .............................................................. 166 
Figure 8.3: BTE % versus EGR % flow rate ......................................................................... 166 
Figure 9.4: Variation of engine brake power versus varying engine load % ......................... 168 
Figure 9.5: Variation of EGT ⁰C versus EGR % flow rate .................................................... 169 
Figure 9.6: Variations of NOX emissions (ppm) versus EGR % flow rate ............................ 170 
Figure 9.7: Variations of NOX emissions (ppm) versus varying engine load % ................... 171 
Figure 9.8: Unburnt hydrocarbons emissions versus EGR % flow rate ................................ 172 
Figure 9.9: CO emission % versus varying engine load ........................................................ 174 
Figure 9.10: Variation of CO VS EGR % flow rate application ........................................... 174 
Figure 9.11: Variation of CO2 % emissions versus engine load %, with different types of fuel 
blends of WPPO and conventional diesel ........................................................................ 176 
xv 
 
Figure 9.12: Variations of CO2 % versus EGR % flow rate .................................................. 177 
Figure 9.13: Variation of smoke emissions or opacity % versus EGR % flow rate .............. 178 
 
 
  
xvi 
 
ABBREVIATIONS AND ACRONYMS 
 
AFR   Air Fuel Ratio 
ASTM  American Society of Testing Materials 
BP  Brake Power 
BRICS  Brazil Russia India China and South Africa 
BSEC  Brake Specific Energy Consumption 
BSFC  Brake Specific Fuel Consumption 
BSHC   Brake Specific Heat Capacity 
BTE   Brake Thermal Energy 
CA  Crank Angle 
CI   Compression Ignition 
CN  Cetane Number 
CNG  Compressed natural gas 
CO2   Carbon Dioxide 
COMESA Common Markets for Eastern and Southern Africa 
CBD  Corn Biodiesel 
CI  Compression Ignition 
CR  Compression Ratio 
CSBD  Cotton Seed Biodiesel 
CV   Calorific value 
DI  Direct Injection 
EC  Elemental Carbon 
EAC  Eastern African Community 
ECOWAS Economic Community of Western African States 
xvii 
 
EGR    Exhaust Gas Recirculation 
EGT  Exhaust Gas Temperature 
EU  European Union 
ER  Equivalence Ratio 
GCV  Gross Calorific Value 
GTL  Gas to Liquid 
HD  Heavy Duty 
HCCI   Homogeneous Charge Compression Ignition 
HP  Horse Power 
HRR  Heat Release Rate 
HVO  Hydro-Treated Vegetable Oil 
IMEP   Indicated Mean Effective Pressure 
Irr  Irreversibility Ratio 
ISFC   Indicated Specific Fuel Consumption 
ITE  Indicated Thermal Efficiency 
Kg/kW-h kilogram per kilowatt hour 
Kg/m3  Kilogram per meter cube 
Kj   Kilojoules 
kW  kilowatt 
L  The Recess Length 
LD  Light Duty 
LPL  Low Pressure Loop 
LSB  Low Swirl Burner  
LTC  Low Temperature Combustion 
xviii 
 
MK   Modulated Kinetics 
𝑛𝑛  Number of Moles 
𝑛𝑛𝑓𝑓  Fuel Molar Quantity 
nm  Nanometre 
NLFS  Normalized Laminar Flame Speed 
NOX    Nitrogen Oxides 
PBD  Petroleum Based Diesel 
PCCI  Pre-Mixed Charge Compression Ignition 
PFI  Port Fuel Injection 
PPM  Parts Per Million 
PHCBD Preheated Corn Biodiesel 
REGR  Reformed Exhaust Gas Recirculation 
RPM  Revolutions Per Minute 
SADEC South-African Development Community 
SFC  Specific Fuel Consumption 
SHR   Specific Heat Ratio 
SI  Spark Ignition 
SOC  Start of Combustion 
SOF  Soluble Organic Fraction 
SHC  Specific Heat Capacity 
TRAP  Traffic Related Air Pollution 
UHC   Unburned Hydrocarbon 
ULSD   Ultra Low Sulphur Diesel 
VCR  Variable Compression Ratio 
xix 
 
VOCs  Volatile Organic Compounds 
Vs  Flame Speed 
WCOME Waste Cooking Oil Methyl Ester 
WPPO  Waste Plastic Pyrolysis Oil 
  
xx 
 
SUBSCRIPTS SYMBOLS 
 
a  Air 
amb  Ambient 
ax  Axial 
bgf  Burnt Gas Fraction 
bz  Burnt Zone 
c  Compression 
cd  Resistance Due to Conduction Heat Transfer 
cr  Critical 
cv  Resistance Due to The Convectional Heat Transfer 
e  Engine 
df  Diesel Fuel 
dmf  Dual Fuel Mode 
eg  Exhaust Gases   
exh   Exhaust Manifold 
f  Fuel 
fd  diesel fuel flow rate 
fpo  pyrolysis oil flow rate 
fc  Fresh Charge 
fr  Fraction of Air Mass 
g  force of gravitational acceleration 
i  Ith Species 
j  Jth Species 
im  Intake Manifold 
xxi 
 
in  Input / Inside 
out   Outside as Output 
rad  Radial 
𝑟𝑟𝑟𝑟  Radial Heat Transfer 
rf  Flow Rate 
s  Stroke 
sec  Cross Section/Section 
sh  Shell 
th  Throttle 
v  Kinematic Viscosity of the Air Medium 
w  System Wall 
𝑥𝑥  Exhaust Molar Gas Quantity 
y  Inlet Intake Charge Quantity 
z  Remainder of the Intake Charge  
xxii 
 
GREEK SYMBOLS 
 
α  Thermal Diffusivity 
λ    Air Excess Ratio 
ω   Angular Velocity 
𝛄𝛄  Air Fraction Mass 
П   Pressure Ratio 
𝜋𝜋  PI ≈ 3.14 
θ  Angle of Emissivity 
ρ  Density 
𝜑𝜑    Effective Angle of Throttle Opening 
𝛽𝛽   EGR Gases Temperature Mean Reciprocal 
𝜇𝜇  Dynamic Viscosity 
𝜂𝜂  Efficiency 
 
 
 
  
xxiii 
 
GENERAL SYMBOLS 
 
A  Area 
𝐴𝐴𝑒𝑒𝑓𝑓𝑓𝑓  Effective Flow Area 
𝐴𝐴𝑡𝑡ℎ  Throttle Valve Effective Area 
𝐴𝐴𝑠𝑠𝑒𝑒𝑠𝑠  Cross Sectional Area of Gaseous Exchange 
𝐴𝐴𝑠𝑠ℎ  Area of The Shell of the Cylinder 
∈𝐶𝐶   Compression Ratio 
ℇ𝑒𝑒𝑒𝑒𝑒𝑒   EGR Cooler Efficiency 
c  Coefficient of discharge 
𝐶𝐶𝐶𝐶𝐶𝐶𝜑𝜑0 Initial Effective Angle of Opening of the Throttle Valve 
𝐶𝐶𝐶𝐶𝐶𝐶𝜑𝜑  Final Effective Angle of Opening 
𝑐𝑐𝑖𝑖  Empirical Constant 
𝑐𝑐𝑝𝑝  Specific Heat at Constant Pressure 
cr  Critical Pressure 
𝑐𝑐𝑣𝑣  Specific Heat at Constant Volume 
𝑥𝑥𝑒𝑒𝑒𝑒𝑒𝑒  EGR Rate 
𝜂𝜂𝑣𝑣𝑣𝑣𝑣𝑣  Engine Volumetric Efficiency 
d  Diameter 
𝑑𝑑𝑒𝑒𝑒𝑒𝑒𝑒  Diameter of The EGR Valve 
𝑑𝑑𝑣𝑣𝑜𝑜𝑡𝑡  Diameter of The Exhaust Manifold 
D  Distance 
𝐷𝐷𝑝𝑝  Particles Diameter 
Gr  Grashofs Number 
xxiv 
 
h  Heat Transfer Coefficient 
ℎ𝑎𝑎𝑎𝑎  Axial Heat Transfer Coefficient 
ℎ𝑠𝑠𝑣𝑣  Heat Transfer Coefficient Due to Convection Heat Transfer 
ℎ𝑒𝑒𝑎𝑎𝑟𝑟  Heat Transfer Coefficient Due to Radial Heat Transfer 
ℎ𝑒𝑒𝑒𝑒𝑒𝑒  Specific Heat Transfer of The EGR Gases 
H  Specific Enthalpy 
k  Specific Heat Ratio 
𝑘𝑘𝑎𝑎𝑎𝑎𝑎𝑎  Specific Heat at Constant Pressure and Volume in The System  
𝑘𝑘  Thermal Conductivity 
L  Length 
M  Mass 
ṁ  Mass Flow 
ṁ𝑎𝑎  Fresh Air Charge 
ṁ𝑒𝑒𝑒𝑒𝑒𝑒  Mass Flow Rate of the EGR % Gas 
N  Rotational Speed 
Nu  Nusselt Number 
P  Pressure 
𝜌𝜌𝑒𝑒𝑒𝑒𝑒𝑒  Density of The EGR Gases 
Pr  Prandtl Number 
Q  Heat Flux / Heat 
𝑄𝑄𝑎𝑎𝑎𝑎  Heat of Axial Direction 
𝑄𝑄𝑒𝑒𝑎𝑎  Heat from The Exhaust Manifold 
𝑄𝑄𝑒𝑒𝑎𝑎𝑟𝑟  Heat of Radial Direction 
R  Specific Universal Gas Constant 
xxv 
 
Ra  Raleigh Number 
Re  The Reynolds Number 
T  Temperature 
T  Torque 
𝑇𝑇𝑎𝑎𝑎𝑎𝑎𝑎  Ambient Temperature 
𝑇𝑇𝑒𝑒  Temperature of The Engine 
𝑇𝑇𝑒𝑒𝑒𝑒𝑒𝑒  Engine Temperature 
𝑇𝑇𝑒𝑒𝑎𝑎  Exhaust Manifold Temperature 
𝑇𝑇𝑒𝑒𝑒𝑒𝑒𝑒  Temperatures of The EGR Gases 
𝑇𝑇(𝐿𝐿)  Temperature as A Function of Length 
𝑇𝑇𝑤𝑤  Wall Temperature 
𝑇𝑇𝑎𝑎  Temperature as A Function of X 
𝑇𝑇∞   Temperature of The Exiting Gases 
u  Control Signal Input 
V  Volume 
𝑣𝑣∞  Velocity of The Exiting Exhaust Gases 
Ψ  Flow Function 
𝜇𝜇𝑒𝑒𝑒𝑒𝑒𝑒  Dynamic Viscosity 
𝛺𝛺𝑖𝑖  Dimensionless Integral of Order Unity 
𝑟𝑟  Conductive Heat Transfer Resistance 
𝑟𝑟𝑎𝑎𝑎𝑎  Axial Heat Resistance Within the System 
𝑢𝑢  Control Signal / Input 
𝛄𝛄  Fresh Air Fraction Mass 
𝜉𝜉  Stephan-Boltzman Constant 
xxvi 
 
𝜃𝜃  Emissivity Angle 
𝜐𝜐  Velocity 
𝑊𝑊𝑎𝑎𝑤𝑤  System Flow 
𝑤𝑤𝑒𝑒𝑒𝑒𝑒𝑒  The Inlet Intake Manifold System Flow to Engine Cylinders 
Ẏ𝑗𝑗𝑖𝑖   Stoichiometric Coefficient 
 
xxvii 
 
CHAPTER ONE 
INTRODUCTION 
1.1 Introduction and Background 
The world has witnessed phenomenal growth, particularly in the middle class, of most 
emerging and developing economies. However, failure of most public transport systems and 
infrastructure in developing and emerging nations has increased the number of personalized 
transport vehicles on the road. This development has tended to increase the popularity of diesel 
propelled vehicles especially in commercial and personal transport. Consequently, there is a 
high level of demand on conventional sources of fuel energy. Although there have been 
tremendous achievements in technology and advances in automobile manufacturing and 
technology, there is still a need to continue with research to reduce the volume of fuel used and 
pollution emitted. This is necessary to reduce global warming which has seen phenomenal 
growth in the 21st century. It is imperative, therefore, for automobile and engine manufacturing 
industries to incorporate new studies related to reduction in emissions and energy wastage from 
engine exhaust systems and make an effort to improve on and harness exhaust gas recirculation 
(EGR) system gases [1].  
The diesel engine can be defined as a type of heat engine that uses internal combustion 
processes to convert fuel energy into useful mechanical energy [2]. More efficiency is 
demanded in most engine management systems of diesel engines as a way of reducing energy 
losses, particularly those being wasted through the exhaust system which accounts for between 
20 % to 25 % of total energy losses. There have been major strides in reduction of this energy 
waste through re-tapping of the exhaust gases through techniques such as EGR systems. 
Recycling of waste gate gases improves volumetric efficiency, brake thermal efficiency (BTE), 
brake specific fuel consumption (BSFC) and reduction of particulate matter (PM) emissions 
[3]. 
In order to reduce emissions and particulate matter (PM) effluents from automobile diesel, 
particularly nitrogen oxide (NOX), carbon monoxide (CO), and hydrocarbons (HCs), EGR as 
a technique has gained widespread use due to its efficiency and ease of application. EGR 
involves the recirculation of a portion of the exhaust gas along with the intake of a fresh air 
charge into the combustion chamber of the diesel engine for combustion together with the fuel 
injected in the normal power cycle of the diesel engine. EGR as an emission control system 
1 
 
allows significant reduction in NOX gases in almost all types of diesel engines, including light 
duty (LD), medium duty and heavy duty (HD) engine applications. EGR is also suitable for 
primary engines as low as two stroke such as those used in marine operations and applications.  
The use of EGR systems in diesel engines has a primary function of diluting oxygen in the inlet 
intake fresh air charge, by displacing the combustion products of carbon dioxide (CO2) and 
hydrogen dioxide (H2O). These products are known to be inert and so suppress the combustion 
process rate, due to their endothermic dissociation. EGR applications have other merits besides 
increasing efficiency and reduction of emissions [1], including that EGR imparts knock 
resistance and reduces the need for high load fuel enrichment in closed cycle diesel engines. In 
spark ignition (SI) engines, EGR systems enable the vaporization of liquid fuels thus enabling 
improved ignition quality. Cooled EGR reduces the EGR flow rate inlet intake charge 
temperature leading to reduced peak in-cylinder temperatures [1]. Moreover, EGR systems 
lead to reduction in lubricating oil consumption, increased fuel injection pressures, increased 
use of diesel oxidation catalysts, and increased intake manifold boost pressure.  
[4] reports that despite the high thermal efficiency in fuels provided by diesel engines, diesel 
engines have the disadvantage of emitting NOx and PM which are considered to be the main 
pollutants and toxic gases, contributing to air pollution in major urban centers and the 
atmosphere in general. Other disadvantages include increased unburned hydrocarbon (UHCs) 
and soluble organic fractions (SOFs), increased particulate matter (PM), increased 
condensation in UHC and PM deportation in cooler fouling (i.e. through degraded heat transfer 
and higher flow resistance). NOX and PM emissions are currently strictly legislated by 
environmental protection agencies in almost all countries, making diesel propelled engines and 
machines more expensive and less appealing. However, EGR systems greatly reduce NOX 
emissions of diesel engines making a compelling case for this technology to be incorporated 
into the design and manufacturing of diesel engines.  
[5] emphasizes that since NOX emission is a product of temperature, especially in-cylinder 
temperatures above 2000 k, any method that can control instantaneous temperature in the 
combustion chamber to below 2000 k will ultimately reduce the formation of NOX. This is the 
main reason behind the use of EGR as a technique for the reduction of NOX. In the recent past 
a number of studies have been conducted on combustion strategies that impact positively on 
the overall reduction of toxic gas emissions associated with the use of diesel engines, especially 
NOX, CO2, CO, PM, and soot emissions.  
2 
 
There are three general methods of reducing NOX, CO2, CO, PM and soot emissions: the 
alternation of fuel method, the alternation of the combustion process and the after treatment of 
the exhaust gases [6]. The first two employed widely today because of the low cost of 
application although these have the disadvantages of higher fuel economy (high BSFC) and 
low BTE. Azad et al. (2016) propose the combination of low temperature combustion (LTC) 
and shear stress transport (SST) k-ω model as the best strategy for biodiesel combustion control 
in compression ignition (CI) engines. The combustion of biodiesel under LTC strategies have 
been shown to reduce NOX emission gases by a factor of 2 of mass volume magnitude and soot 
emissions by a factor of 10 compared to conventional diesel fuel as reported by [7, 8]. 
Thus, combustion modelling is a vital field due to its role in conserving energy as well as 
providing in-depth knowledge concerning complex combustion process control strategies. 
There are three types of LTC modelling widely in use today, namely, homogeneous charge 
compression ignition (HCCI), pre-mixed charge compression ignition (PCCI), and reactivity 
controlled compression ignition (RCCI). HCCI is a pioneer in the field of diesel combustion 
control strategies; it is a change from the conventional combustion process aimed at creating a 
homogeneous air fuel mixture before ignition in CI engines, which is similar to combustion 
operation in SI engines. The HCCI uses a two-stage heat release combustion system for diesel 
fuel. In stage one it has a low temperature kinetic reaction, controlled and dominated by low 
temperature chemistry. The second stage, which is the main heat release phase, is dominated 
or controlled by the oxidation of the CO gas [6]. HCCI mixture preparation plays a major role 
in reducing soot and NOX emissions. There are two types of charge preparation strategies 
employed, namely, external mixture formation and in-cylinder direct injection (DI), according 
to [9]. External mixture formation leads to reduced volumetric efficiency while the in-cylinder 
DI leads to dilution of the lubricating oil. Various techniques are employed with highly volatile 
fluids such as gasoline and the alcohols, such as port fuel injection (PFI), wide-open throttle 
and fumigation. However, for low volatility fuels like diesel, vaporizers are generally deployed. 
There are a number of researchers who have used this technique and experimented with 
different types of fuels, for example, [10, 11]. 
Pre-mixed charge compression ignition (PCCI) is a combustion control strategy which evolved 
from the HCCI model of combustion due to the limitations that were exhibited by the HCCI 
model, including noisy combustion and violent uncontrolled combustion. This strategy sought 
to bring better control over the start of combustion (SOC) phase. PCCI utilizes early premixing 
of the air fuel mixture followed by a late injection of fuel during the compression stroke, thus 
3 
 
governing how and when the SOC should begin [6]. This makes it possible to utilize very high 
levels of EGR % flow rates. This factor increases ignition delay duration thereby offering better 
air-fuel mixing and offering less fuel-rich pockets in the LTC areas, hence the reduction in 
NOX emissions and soot emissions [12]. Among the advantages offered by PCCI over HCCI 
are ignition delay which is achievable through enhancement of charge motion [6] especially 
for mixtures which are not homogeneous, reduced compression ratio (CR), higher injection 
pressure and use of very high EGR % flow rate applications. This system allows fuel to be 
injected into the combustion chamber in three main ways, namely, advanced early pilot 
injection, PFI, and late direct fuel injection. The first two systems have the disadvantage of 
increasing HC and CO emissions from the residue fuel on the cylinder walls, which is corrected 
by use of narrow spray injector nozzles and application of high EGR % flow rate, thus 
switching the occurring combustion so that it looks similar to conventional combustion 
especially during low engine loads.  
Reactivity controlled compression ignition (RCCI) [13] is a more recent combustion strategy 
development. It uses multiple fuels with different reactivity rates and properties as a 
combustion fuel. Fuel is injected at predetermined intervals thus governing the reactivity of the 
injected fuels in the combustion cylinder for the desired combustion and magnitude. This 
allows the onset of combustion at different periods and at different rates and intensity. In RCCI 
the combustion is staged to start from local high reactivity fuel areas to low fuel reactivity areas 
[14]. The staging of the combustion process leads to significant expansion of the premixed 
combustion. Hence the high thermal efficiency, low in-cylinder pressure rise rate and low 
emissions, even when the engine is subjected to high load conditions of up to 16 Mpa of 
indicated mean effective pressure (IMEP) [15].  
Another combustion model worth mentioning is the one reported by [16, 17] i.e., low swirl 
burner (LSB). It uses premixed flame technology especially for leaner flame stabilization, 
where its main principle of operation is in propagating turbulent premixed flames. The standing 
flame front [16] in the LSB is produced through weakening the swirl motion of the premixed 
mixture thus providing a divergent down stream flow of the premixed mixture which produces 
a low temperature flame. These conditions make it possible for low levels of NOX emissions 
to be achieved. 
The use of biodiesel as an alternative fuel because of its physio-chemical fuel properties [6] 
does not replace diesel. This is due to the difference in the combustion process between the 
4 
 
two fuels, as they too produce toxic gases. Although biodiesels produce toxic gases, this can 
be mitigated through adoption of good combustion strategies to meet the current demand for 
leaner and more fuel-efficient engines. This is as demanded from national and global 
environmental protection agencies. Recently developed biodiesel combustion control strategies 
have been developed by [18, 19]. However, it should be mentioned here that there is no single 
and perfect model that can produce a solution to the many combustion emission challenges 
experienced in modern day diesel internal combustion engines.  
Table 1 shows various research groups and their strategies regarding LTC [16]. 
Table 1.1: LTC research strategies 
Acronym Meaning Location 
ATAC Active thermo-atmospheric combustion Nippon Clean Engine 
Research Institute 
TS Toyota-Soken combustion Toyota/Soken 
CIHC Compression ignited homogeneous charge University of Wisconsin 
Madison 
HCCI Homogeneous charge compression ignition SwRI 
ARC Active radical combustion Honda 
PREDIC Pre-mixed diesel combustion New Ace 
HiMICS Homogeneous intelligent multi-injection system Hino 
UNIBUS Uniform bulky combustion system Toyota 
MULDIC Multi stage diesel combustion New Ace 
PCCI Pre-mixed charge compression ignition Mitsubishi 
MK Modulated kinetics Nissan 
NADI Narrow angle direct injection Institute Francais du Petrole 
  
The importance of EGR in gas emission and control cannot be over-emphasized, as it affects 
the formation of deposition particles in the diesel engine. The particles formed as a result of 
5 
 
application of EGR % flow rate have far reaching implications on human health and the 
environment at large. The distribution of particles emitted from diesel engine can be 
categorized into two major areas: the Nano scale group which is 𝐷𝐷𝑝𝑝  <  50 𝑛𝑛𝑛𝑛 often referred 
to as the particle nucleation mode, and the Nano scale group of 50 𝑛𝑛𝑛𝑛 <  𝐷𝐷𝑝𝑝  <  1000 𝑛𝑛𝑛𝑛 
known as the accumulation mode [9]. The former includes the volatile organic compounds 
(VOCs) and some solid carbon metal compounds while the latter includes carbonaceous soot 
agglomerates which have a significant effect on the residue particle mass, which can be 
expressed mathematically in equation form [9] as: 
3
𝑀𝑀𝑃𝑃 = 𝜌𝜌
𝜋𝜋𝑑𝑑
𝑝𝑝           Equation 1 6
Where 𝑀𝑀𝑝𝑝 is the dependence particle mass, 𝜌𝜌𝑝𝑝 is the particle density, 𝑑𝑑 is the particle diameter, 
𝜋𝜋 is pi = 3.14. 
The finer range particles of PM2.5 are more dangerous than the larger range particles PM10 [9]. 
The finer particles are more prevalent in diesel particulate exhaust emissions, especially LD 
diesel vehicle engines, indicating that there is a correlation between particulate emission size 
and the cubic capacity of a diesel engine. However, [20] reported lower solid particle emissions 
from HD diesel engines compared to the emissions from LD diesel engines, a factor they 
attributed to three main factors: higher EGR rates, rapid combustion in HD diesel engines and 
higher PM limits of emissions associated with LD diesel engines. Therefore, the authors 
concluded that emission particles from diesel engines are not directly proportional to engine 
cubic capacity. [20] further reported that with increase in the use of EGR % flow rate from 0 
% to 19 %, the particle diameter increased by 70 %, which can be credited to the thermal effect 
of EGR % flow rate. This was due to the dilution effect which plays a dominant role in 
producing most primary emission particles. They concluded that the geometrical size of 
primary particle diameter formation is affected by the variation of the EGR % flow rate. 
1.2 Aims and Objectives 
The aim of this dissertation work was to investigate the effects of EGR on diesel engine 
performance. This work had set three main objectives that were pursued in answering this aim, 
namely:  
1. To investigate the effects of EGR on the performance of a diesel engine.  
6 
 
2. To investigate development of an alternative fuel and assess how it affects diesel engine 
performance and emissions control under the influence of EGR.  
3. To investigate the effects of EGR on performance and emission characteristics of a diesel 
engine. 
1.3 Dissertation Outline 
This dissertation work sought to achieve the aims and objectives presented above in nine 
chapters, as summarized below.  
Chapter 1 presents an introduction and background to the study, outlining the need to study 
exhaust gas recirculation as a technique of for emission control and its effects on performance. 
This chapter provides a background on the main research question that is linked to the aims 
and objectives of this study. The layout of this dissertation, its scope of study, research 
motivation and contribution to the field of science and technology especially renewable energy 
are outlined. 
Chapter 2 is the literature review. This chapter deals with previous work by various authors 
conducted in the same field of study. Various topics of importance in EGR emission control 
strategy, development in alternative fuels and recent development in combustion control 
strategies and their wider implications for diesel engines are comprehensively reviewed and 
discussed. This review consists of proceedings from conferences, journal articles, published 
books, online articles and publications and dissertations from masters and thesis from doctoral 
levels. 
Chapter 3 deals with the methodology that was employed in meeting the aim and objectives 
of this work. Two main approaches are proposed, namely, the numerical and experimental 
methods. However, in this chapter only the numerical approach is attempted with mathematical 
models presented to help build a firm foundation for deepening understanding of the aim and 
objectives of this work. This chapter attempts to link numerical models with the experimental 
procedure set-up for study. 
Chapter 4 is a review publication entitled “The Effects of Exhaust Gas Recirculation on The 
Performance and Emission Characteristics of a Diesel Engine – A Critical Review”. In this 
chapter various topical issues linked to EGR such as performance, combustion control strategy, 
emission characteristics of NOX, CO, CO2, UHC, PM, and smoke emissions, are discussed. 
7 
 
Chapter 5 is a publication entitled “Transportation, Pollution and the Environment”. This 
chapter discusses the importance of modern day forms of transport and their effects on the 
environment as a result of vehicular emissions. These emissions result from the combustion 
characteristics of diesel engines, a primary fuel in the both commercial and personal transport. 
Chapter 6 is a publication entitled “The Formation, Effects and Control of Oxides of Nitrogen 
in Diesel Engines”. This chapter covers topical issues on how oxides of nitrogen are formed 
within diesel engines, the sources of NOX, types of formation routes and interaction 
mechanisms with one another and the effects of NOX emissions on the environment and human 
public health. 
Chapter 7 covers the experimental set-up, the apparatus and the procedure for conducting the 
experiment. It outlines in detail experimental equipment and specification with percentage 
error of their measurement and vital modifications for the EGR circuit. In addition, this chapter 
outlines the procedures experimental material preparation, data collection and the sequence in 
which the experimental procedure was executed.  
Chapter 8 presents the results and discussion as observed and reported from the experimental 
study work described in Chapter 7. It breaks down the results into comprehensive detailed 
individual components, through diagrammatic graphic figures. Each result is analyzed and 
discussed independently, according to the result obtained and conclusions deduced. This 
chapter provides a systematic sequence of results discussion dealing with performance 
parameters of BSFC, BTE, BP and EGT with emission components of NOX, CO, CO2, UHC 
and smoke opacity.  
Chapter 9 discusses the conclusions and future recommendations from the experimental 
results obtained in Chapter 8 and discusses the research findings in the light of the research 
gaps presented in the literature review in Chapter 2.  
1.4 Research Motivation 
The impact of global warming is growing due to the increasing atmospheric air pollution in our 
major urban centers and cities. This development is associated with the expansion of urban 
transport systems, industrial development and the increase in world population (projected to 
grow by 2 billion people annually) [21]. Environmental degradation caused by pollution 
threatens food security. There is increasingly stringent emission control legislation being 
passed in individual countries, emerging economies, the G7, and economic blocks like the 
8 
 
COMESA, EAC, ECOWAS, European Union (EU), SADC, BRICS. These challenges call for 
research in renewable energy and alternative fuels that can reduce emissions and pollution, 
allowing time for adjustments and wider applications in the transport sector. The transport 
sector is a pivotal GDP economic factor in the production chain, without which other equally 
important sectors of the economy cannot be sustained.  
Figure 1 shows the past and predicted future variation of world energy consumption with 
percentage shares for each source of energy [21]. 
 
Figure 1.1: Variation of world energy consumption percentage shares for each source of energy [21] 
 
The development of alternative fuels from recycling technology offers a reprieve for the 
primary sources of energy and of biodiesel alternative fuels some of which rely heavily on food 
plants and other food sources. Recycling technology mitigates the challenges and 
disadvantages identified and associated with petroleum-based fossil fuels, which are leading 
contributors to global warming and atmospheric air pollution. The above factors are a profound 
motivation for this work and study and have driven my focus, and will continue to do so, with 
a view to contributing through this work to make our world a better place. 
  
9 
 
References 
 
[1] Manieniyan V and S. S, "“Experimental Analysis of Exhaust gas Recirculation on 
Diesel Engines Operating with Biodiesel," International journal of Engineering and 
Technology (IJET), vol. 3, pp. 129-134, 2013. 
[2] R. Senthilkumar, K. Ramadoss, and R. Manimaran, "Experimental investigation of 
performance and emission characteristics by different Exhaust gas recirculation 
methods used in diesel engine," Global Journal of Research In Engineering, vol. 13, 
2013. 
[3] J. Hussain, K. Palaniradja, N. Alagumurthi, and R. Manimaran, "Effect of exhaust gas 
recirculation (EGR) on performance and emission characteristics of a three cylinder 
direct injection compression ignition engine," Alexandria Engineering Journal, vol. 51, 
pp. 241-247, 2012. 
[4] D. Agarwal, S. K. Singh, and A. K. Agarwal, "Effect of Exhaust Gas Recirculation 
(EGR) on performance, emissions, deposits and durability of a constant speed 
compression ignition engine," Applied Energy, vol. 88, pp. 2900-2907, 2011. 
[5] K. S. Rao, "Effect of EGR on Diesel Engine Performance and Exhaust Emission 
Running with Cotton Seed Biodiesel," international journal of mechanical and 
mechatronic engineering, vol. 16, pp. 64-69, 2016. 
[6] A. Azad, M. Rasul, M. Khan, S. C. Sharma, and M. Bhuiya, "Recent development of 
biodiesel combustion strategies and modelling for compression ignition engines," 
Renewable and Sustainable Energy Reviews, vol. 56, pp. 1068-1086, 2016. 
[7] M. Arbab, H. Masjuki, M. Varman, M. Kalam, S. Imtenan, and H. Sajjad, "Fuel 
properties, engine performance and emission characteristic of common biodiesels as a 
renewable and sustainable source of fuel," Renewable and Sustainable Energy Reviews, 
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[8] A. Azad, M. Rasul, M. Khan, A. Omri, M. Bhuiya, and M. Hazrat, "Modelling of 
renewable energy economy in Australia," Energy Procedia, vol. 61, pp. 1902-1906, 
2014. 
[9] J. Thangaraja and C. Kannan, "Effect of exhaust gas recirculation on advanced diesel 
combustion and alternate fuels-A review," Applied Energy, vol. 180, pp. 169-184, 2016. 
[10] S. S. Nathan, J. Mallikarjuna, and A. Ramesh, "Effects of charge temperature and 
exhaust gas re-circulation on combustion and emission characteristics of an acetylene 
fuelled HCCI engine," Fuel, vol. 89, pp. 515-521, 2010. 
[11] M. M. Ibrahim and A. Ramesh, "Experimental investigations on a hydrogen diesel 
homogeneous charge compression ignition engine with exhaust gas recirculation," 
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[12] M. P. Musculus, P. C. Miles, and L. M. Pickett, "Conceptual models for partially 
premixed low-temperature diesel combustion," Progress in Energy and Combustion 
Science, vol. 39, pp. 246-283, 2013. 
[13] S. Imtenan, M. Varman, H. Masjuki, M. Kalam, H. Sajjad, M. Arbab, et al., "Impact of 
low temperature combustion attaining strategies on diesel engine emissions for diesel 
and biodiesels: a review," Energy Conversion and Management, vol. 80, pp. 329-356, 
2014. 
[14] D. Splitter, S. Kokjohn, K. Rein, R. Hanson, S. Sanders, and R. D. Reitz, "An optical 
investigation of ignition processes in fuel reactivity controlled PCCI combustion," SAE 
International Journal of Engines, vol. 3, pp. 142-162, 2010. 
10 
 
[15] K. Pohlkamp and R. D. Reitz, "Reactivity controlled compression ignition (RCCI) in a 
single-cylinder air-cooled HSDI diesel engine," SAE Technical Paper 0148-7191, 
2012. 
[16] S. I. Pishbin, M. Ghazikhani, and S. M. R. M. Razavi, "Experimental study on the 
effects of flame regime on the exergy destruction in premixed low swirl combustion," 
International Journal of Exergy, vol. 17, pp. 267-286, 2015. 
[17] R. Cheng, D. Littlejohn, W. A. Nazeer, and K. Smith, "Laboratory studies of the flow 
field characteristics of low-swirl injectors for adaptation to fuel-flexible turbines," 
Journal of Engineering for Gas Turbines and Power, vol. 130, p. 021501, 2008. 
[18] L. Coniglio, H. Bennadji, P. A. Glaude, O. Herbinet, and F. Billaud, "Combustion 
chemical kinetics of biodiesel and related compounds (methyl and ethyl esters): 
Experiments and modeling–Advances and future refinements," Progress in Energy and 
Combustion Science, vol. 39, pp. 340-382, 2013. 
[19] S. J. Klippenstein, V. S. Pande, and D. G. Truhlar, "Chemical kinetics and mechanisms 
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American Chemical Society, vol. 136, pp. 528-546, 2014. 
[20] J. Zhu, K. O. Lee, A. Yozgatligil, and M. Y. Choi, "Effects of engine operating 
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[21] A. Hegab, A. La Rocca, and P. Shayler, "Towards keeping diesel fuel supply and 
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11 
 
CHAPTER TWO 
LITERATURE REVIEW 
2.1 Introduction 
There are a number of studies and research publications that have been written on the effects 
of exhaust gas recirculation (EGR) on diesel engines in relation to their performance and 
emission characteristics. Emissions is a pivotal factor in internal combustion engines because 
of the stringent emission control regulations that keep on changing day by day. This review 
work will look at the development of EGR as a control measure for NOX emission reduction 
and how it affects the parameters of diesel engine performance in terms of both the dependent 
and independent variables. It is imperative for both the old and new emission control techniques 
to be harnessed, as there is no one solution that can solve the global emission problem 
associated with diesel internal combustion engines. Since we are dealing with diesel internal 
combustion engines performance and emission characteristics, this review will look critically 
at the developments in relation to combustion control strategies. This work will then look at 
developments in alternative fuel which have been gaining momentum as the reserves of our 
primary sources of energy become increasingly depleted, and the concerns about pollution 
increase. Additionally, this review work will examine the impact of EGR on biodiesel fuels 
whose use is gaining widespread acceptance and is being researched by various groups around 
the world. This is important as waste to energy solutions are gaining momentum in energy 
studies and development, to replace earlier biodiesels which were primarily from agricultural 
food plants. This is intended to address and answer the problem of food security especially in 
developing countries. Therefore, several works on recycled waste energy have been considered 
here and reviewed. Finally, this review will look at the global impact of the emissions from 
diesel propelled automobiles on human health and what the field has been doing to address the 
problem and the environmental pollution causing it. 
2.2 EGR, Combustion and Development in Control Strategies  
The future of the diesel engine considering the current context demands modification in 
combustion strategies coupled with development of high quality fuels not limited to only 
conventional diesel, to meet increasingly stringent emission control legislation. It would be 
unfair to this review if modern day combustion strategies were not discussed and reviewed to 
reveal how this area has been progressing over time, and how these studies have impacted on 
the performance and emission characteristics of the diesel engine. EGR is considered as one of 
12 
 
the most effective techniques for controlling NOX and PM emissions, mainly due to the many 
advantages it offers, which include lower flame temperatures and decrease in oxygen 
concentration in the combustion mixture of the combustion chamber.  
Zheng et al. (2004) reported that the main challenge is to minimize the pollutants by 
manipulation of the thermodynamic fuel properties and oxygen concentration of the cylinder 
charge, through limiting and minimizing degradation of power and deficiencies of the diesel 
engine. [1] says that under conventional operations the diesel engine burns 99.5% of injected 
fuel, but when operating in LTC environment, this efficiency is decreased to around 95 %, 
resulting in misfiring brought about by the lower boundaries of LTC and the lack of sufficient 
fuel for flame propagation. [1] further suggests that to mitigate LTC it is imperative that intake 
temperatures and intake pressure be controlled as these are reported to reduce LTC NOX and 
PM by 24 % and 50 % respectively; when intake boost pressure is increased by 1 bar, it reduces 
HC and PM by 50 % and 75 % respectively. 
The EGR system can be classified into three main categories according to [12]. 
• Hot EGR, where the circulated exhaust gas is directly delivered into the intake system 
when hot, resulting in very high inlet intake temperatures. 
• Partly Cooled EGR, where the circulated gas temperatures are kept just above the dew 
point.  
• Fully Cooled EGR, where the circulated gas is fully cooled before it is mixed with the 
intake fresh charge, either using water cooled heat exchangers or intercooling 
mechanisms. 
Figure 2.1 shows the variation of NOX emissions with oxygen mass fraction in hot and cool 
EGR modes. 
 
Figure 2.1: Variation of NOX emissions with oxygen mass fraction in hot and cool EGR modes [2] 
 
13 
 
Increasing the EGR ratio changes the shape of the heat release rate (HRR) especially during 
pre-mixed combustion, which greatly suppresses NOX formation as reported by [3, 4]. When 
compared to petrol engines, diesel engines almost remain unaffected by the application of a 
high EGR % flow rate. This is due to the fact that EGR affects the stoichiometric conditions of 
SI engines by directly reducing the volumetric efficiency i.e. the total number or amount of 
oxygen that can be combusted with fuel [1].  
Due to the increase in the rate of EGR in diesel engines the following effects on performance 
and combustion are observed [4]: 
Dilution Effect 
EGR application, as aforementioned, decreases the oxygen concentration in the combustion 
cylinder, therefore the fuel spray has to travel, diffuse and mix with oxygen to form a balanced 
stoichiometric mixture suitable for combustion, [3, 5, 6]. The localized regions within the 
combustion chamber not only contain a stoichiometric mixture or a balanced mixture but also 
contain other additional gases, namely, CO2, H2O and N2 [4]. These combustion components 
absorb the heat energy released during combustion, thereby lowering the temperature and 
consequently affecting the kinetic movement of NOX elements and thus the formation of NOX. 
Ignition Delay 
The presence of diluent elements of CO2 and H2O increase ignition delay (combustion 
retardation) by changing or altering the location of the SOC, shifting it towards the end of the 
expansion stroke, further reducing the exposure of the combustion mixture to high temperatures 
and thus forming less NOX [3, 7]. 
Chemical Effects 
During the high temperature combustion process recirculated CO2, H2O and water vapor 
disassociate, modifying in the process the combustion temperature and the formation of NOX. 
Due to the reduction in the flame temperature this leads to decreased NOX emissions [1, 3]. 
Thermal Effects 
Thermal effects result from the higher heat capacity of CO2 and H2O in the EGR gases as 
compared to the heat specific ratio of the incoming charge i.e. the intake charge which has CO2 
and N2. This phenomenon is responsible for the overall increase in the heat capacity of the in-
cylinder mixture thereby reducing the flame temperatures. [8] conducted various experiments 
14 
 
and detailed research on irreversibility in the combustion process and found out that the 
chemical reactions and transport phenomena are responsible for irreversibility in the 
combustion process. This work was achieved by measuring the role of exergy destruction in 
the combustion process exchange owned by the internal thermal energy exchange. Exergy is 
defined as energy availability and defined further by [9] as the “maximum possible useful work 
that can be obtained from energy in a given environment”, a discussion derived from the second 
law of thermodynamics in the analysis and examination of engine performance responses under 
different operating conditions and parameters. The second law of thermodynamics states that 
“No heat engine can produce a network output by exchanging heat with a single fixed 
temperature region” [10] and is expressed in equation form as: 
𝑄𝑄𝑎𝑎𝑑𝑑𝑑𝑑 𝑎𝑎𝑛𝑛𝑑𝑑 𝑄𝑄𝑟𝑟𝑟𝑟𝑟𝑟  ≠ 0         Equation 1 
Where 𝑄𝑄𝑎𝑎𝑑𝑑𝑑𝑑 is the heat added into the system, 𝑄𝑄𝑟𝑟𝑟𝑟𝑟𝑟 is the heat being ejected by the system. 
Another area worthy of mention in connection with this review and the scope of this study, is 
the EGR ratio which has been expressed by many scholars and researchers in different forms 
and mathematical equations, for example Equation 2 and 3 [11] and expressed as: 
𝐸𝐸𝐸𝐸𝐸𝐸 % =  (𝐶𝐶𝑜𝑜2−𝐸𝐸𝐸𝐸𝐸𝐸−𝐶𝐶𝑂𝑂2−𝑎𝑎𝑎𝑎𝑎𝑎)        Equation 2 
(𝐶𝐶𝑜𝑜2−𝐸𝐸𝐸𝐸𝐸𝐸−𝐶𝐶𝑂𝑂2−𝑎𝑎𝑎𝑎𝑎𝑎)
Where 𝐶𝐶𝐶𝐶2 is the recirculated carbon dioxide fraction, 𝐶𝐶𝐶𝐶2 is the carbon dioxide in the exhaust 
system, 𝐸𝐸𝐸𝐸𝐸𝐸 % is the exhaust gas recirculated mass flow rate percentage. 
From Equation 3 [11] define the percentage rate of EGR as “the mass percentage of the 
recirculated exhaust gas (MEGR) in the total intake mixture (Min).” This suggests the use of the 
recirculated CO2 fraction by comparing CO2 concentration in the exhaust (CO2-exh) and CO2 
concentration in the intake manifold of the engine (CO2-int) to determine the EGR flow rate. 
After defining the EGR % flow rate the authors further found that the use of EGR increases the 
specific heat capacity (SHC) and decreases oxygen availability. This directly reduces the peak 
cylinder pressure and leads to the conclusion that decreased combustion temperature affects 
peak heat release ratios if and when exhaust gas recirculation is applied. Regarding the intake 
mass flow rate, they observed that an increase of EGR % flow rate reduces the intake mass 
flow rate as compared to when the EGR % flow rate is zero. This affects the volumetric 
efficiency as less air is admitted to the engine cylinders. However, as EGR % flow rate 
increases they observed a decrease in fuel efficiency, which they attributed to the application 
of the EGR % flow rate.  
15 
 
The second method is derived from the second equation of calculating exhaust gas recirculation 
proposed by [12], where they represent the percentage of recycled gases by an EGR ratio (the 
mass ratio of recycled gases as compared to the whole mixture intake into the engine) by 
comparing it to CO2 in the fresh air intake which is assumed here to contain negligible amounts 
of CO2 and the EGR CO2 which contains vast amounts of CO2, as the rate of EGR % flow rate 
increases. This is captured in Equation 3 from [12]: 
 𝐸𝐸𝐸𝐸𝐸𝐸 𝐸𝐸𝑅𝑅𝑅𝑅𝑅𝑅𝐶𝐶 = 𝐼𝐼𝐼𝐼𝐼𝐼𝐼𝐼𝐼𝐼𝐼𝐼 𝐶𝐶𝑂𝑂2 𝐶𝐶𝑜𝑜𝐶𝐶𝐶𝐶𝑟𝑟𝑎𝑎𝑟𝑟𝑎𝑎𝑎𝑎𝐶𝐶𝑜𝑜𝐶𝐶       Equation 3 
𝐼𝐼𝐸𝐸𝐸𝐸𝐼𝐼𝐸𝐸𝐸𝐸𝐼𝐼 𝐶𝐶𝑂𝑂2 𝐶𝐶𝑜𝑜𝐶𝐶𝐶𝐶𝑟𝑟𝐶𝐶𝑎𝑎𝑟𝑟𝑎𝑎𝑎𝑎𝐶𝐶𝑜𝑜𝐶𝐶
The EGR ratio in Equation 3 is calculated or measured from the amounts of CO2 in the intake 
charge and the figure is compared to the amount of CO2 in the exhaust gases. It should be noted 
that the concentration of the engine intake CO2 is lower than that from the cylinder charge, but 
the residue amount is boosted to dilute the fresh charge intake of the engine thereby providing 
prompt EGR, which normally falls outside the concentration of the intake CO2 defined above 
and as shown in the linear Equations 4 and 5 [12] as: 
𝑥𝑥𝑂𝑂2 = 0.21 (1 − 𝑥𝑥𝐶𝐶𝑂𝑂2)        Equation 4 
𝑥𝑥𝐼𝐼2 = 0.79(1 − 𝑥𝑥𝐶𝐶𝑂𝑂2)        Equation 5 
Where 𝑥𝑥𝑂𝑂2 is the exhaust molar concentration of oxygen, 𝑥𝑥𝐼𝐼2is the exhaust molar concentration 
of nitrogen, 𝑥𝑥𝐶𝐶𝑂𝑂2 is the molar concentration in the exhaust system. 
In a simulation experiment of the Equation 4, [12] observed that as the molar concentration of 
CO2 increases, the molar concentration of O2 and N2 mixtures decrease linearly as in Equation 
4. A simulated EGR operation was used to explain the prompt EGR phenomena and the 
thermodynamic and dilution effect of EGR. However, since EGR affects the equilibrium state 
of the entire system, it makes feedback control conservation necessary. Therefore, in order to 
estimate the EGR rate the air intake was measured using a MAF sensor, by making an 
assumption that the cylinder mass flow rate and that of the EGR % flow rate are to be conserved 
and therefore assumed to be as in Equation 6:  
Ṁ𝐼𝐼𝐸𝐸𝐸𝐸  = Ṁ𝐼𝐼𝐼𝐼𝐼𝐼  −  Ṁ𝑀𝑀𝐼𝐼𝑀𝑀        Equation 6 
Where Ṁ𝐼𝐼𝐸𝐸𝐸𝐸 is the mass flow rate of the EGR gases, Ṁ𝐼𝐼𝐼𝐼𝐼𝐼  is the inlet intake mass flow, Ṁ𝑀𝑀𝐼𝐼𝑀𝑀 
is the mass flow rate as measured by the mass air flow sensor. 
16 
 
But the mass flow rate of the cylinder is dependent on a number of factors such as EGR gases 
temperatures, engine block temperature, after cooling temperature and boost pressure, which 
are outside the monitoring capacity of the ECM. General engine operating stabilities are 
affected by the widespread use of EGR operation through the following key areas as noted by 
[11]: 
• EGR loop differential pressure 
• EGR cooler cooling efficiency 
• In-cylinder combustion efficiency 
• Exhaust temperatures  
• Intake temperatures 
Although EGR is a vital method in the control of NOX it is still important to say that it has 
application limits, which lead to degradation of efficiency of the energy derived from the fuel 
and the mechanical durability of the diesel engine components. Among the measures [11] 
recommend is exhaust gas recirculating cooling, because of its proven capacity to increase the 
intake inlet charge density. This is equivalent to the intercooling boosting of the mass air flow 
rate of the admitted charge. This was identical to the findings of [13]. In addition to EGR 
cooling they additionally recommend the adoption of an EGR fuel reforming catalytic rich 
combustor into the EGR loop to generate gaseous fuels. These gaseous fuels are intended to 
produce hydrogen gas and carbon monoxide in a controlled amount through the process of 
diesel reformation. Fuel reforming shares the advantages of homogeneous charge CI engine 
systems, and can be expressed in equation form to give a mathematical expression as in 
Equation 7:  
λ gas ≥ 1.35          Equation 7 
Where 𝜆𝜆 𝑔𝑔𝑎𝑎𝑔𝑔 is the air fuel ratio, in this case of the gas fuel. 
[1] discusses a number of control strategies to be employed in mitigating the effects of 
emissions, one of which is EGR. The authors point out that the application of EGR alters the 
normal combustion process in several ways: firstly, by altering the physical properties of the 
charge mixture; secondly by altering the air fuel ratio (AFR); and thirdly by altering the start 
of ignition thereby altering the heat release pattern.  
17 
 
 
Figure 2.2: EGR loop – high pressure loop (HPL) 
Figure 2.3: EGR loop – low pressure loop (LPL) [14] 
 
From Figure 2.2 above one can see that if the EGR system gases are taken from the exhaust 
upstream of the turbine directly into the intake manifold where pressure 𝑃𝑃3 − 𝑃𝑃2 > 0, the 
delivery arrangement pipe is shorter, hence the name short part (HPL) EGR, where the pressure 
𝑃𝑃3  > 𝑃𝑃2 where in this case it is the intake pressure. The long part arrangement, known as low 
pressure loop (LPL) EGR, is when 𝑃𝑃3 − 𝑃𝑃2 < 0 as shown in Figure 2.3. The LPL has the 
disadvantage of ravaging aluminium made components like the compressor and the plumbing 
components, while the HPL has the demerit of low boost pressure requiring a turbocharger to 
be matched as reported by [15]. This allows the turbocharger to increase the boost pressure 
with the addition of a venturi device as an auxiliary unit [6, 13] to help in expanding the feasible 
range of boosting without affecting fuel economy (pumping losses increase as boost pressure 
drops) [16]. 
While discussing novel combustion strategies it is imperative to mention a very important 
factor in combustion studies, the factor of the dilution ratio (DR), derived and defined by [6] 
as in Equation 8 as:  
𝐷𝐷𝐸𝐸 % = 100 ∗ 21−𝐸𝐸𝑜𝑜2−𝑖𝑖𝑖𝑖        Equation 8 
21
Where DR % is the dilution ratio percentage, 𝑋𝑋𝑜𝑜2−𝐶𝐶𝐶𝐶 is the O2 concentration as measured by 
the output tools in the inlet manifold, especially for the short part HPL EGR system which is 
mostly used with commercial diesel engines. 
18 
 
Dilution ratio plays a dominant role in the formation of NOX emissions especially when LTC 
modelling is used. [14] add that the EGR effect of dilution, in addition to the aforementioned, 
plays a very influential role in the reduction of NOX both in advanced concepts of combustion 
and conventional concepts of combustion. However, in relation to the emissions of soot and 
homogeneity of the combustion mixtures, it is ignition delay (combustion retardation) that 
becomes the dominant factor in advanced combustion control modelling. Among the problems 
the authors found to be associated with the HCCI include poor cold starting, higher HC/CO 
emissions, limited operation range, difficulty in combustion phase control, control over auto 
ignition and preparation of the charge mixture.  
The pre-mixed ratio (PR) as a control strategy is normally associated with the PCCI combustion 
strategy that has been adapted by a number of studies notable among them [17, 18] and defined 
in Equation 9 as: 
 𝑟𝑟 = 𝑄𝑄𝑃𝑃𝑃𝑃           Equation 9 𝑄𝑄𝑎𝑎
Where the 𝑟𝑟𝑝𝑝  is the pre-mixed ratio, 𝑄𝑄𝑃𝑃  is the ratio of energy of the premixed fuel, 𝑄𝑄𝑎𝑎  is the total 
energy 
But:  
𝑄𝑄  = 𝑀𝑀𝑝𝑝𝑃𝑃           Equation 10  ℎ𝑢𝑢𝑝𝑝
𝑄𝑄𝑎𝑎 = 𝑀𝑀𝑝𝑝ℎ𝑢𝑢𝑝𝑝 + 𝑀𝑀𝑑𝑑ℎ𝑢𝑢𝑑𝑑        Equation 11 
Where 𝑀𝑀𝑝𝑝 is the pre-mixed fuel mass, 𝑀𝑀𝑑𝑑 is the diesel fuel mass, 𝑝𝑝 denotes pre-mixed diesel 
fuel, ℎ𝑢𝑢𝑝𝑝 is the lower heating value of the premixed diesel fuel mass. 
[19] in “Fuel Injection and EGR Control Strategy on Smooth Switching Of CI/HCCI Mode in 
a Diesel Engine” conducted an experiment by varying the cam phase and tapping EGR gases 
as a heating or vaporizing mechanism of the fuel to harmonize the mixture charge, while using 
fuel injection to administer HCCI with an electronically controlled fuel injection system. The 
authors used HCCI combustion modelling to study the effects of a premixed ratio in relation to 
performance and emission characteristics, when different EGR % flow rates were applied. 
They observed a 78 % drop in NOX and lowered smoke and soot emissions of 40 %, when 
premixed ratios of 80 % and 30 % EGR respectively were applied. An increased pre-mixed 
ratio led to an increase in IMEP and BSFC, but fared poorly regarding HC and CO emissions. 
19 
 
[20] in “Recent Development of Biodiesel Combustion Strategies and Modelling for 
Compression Ignition Engines” reported recent developments in combustion strategies for 
biodiesel engines, as biodiesel fuels have the same physio-chemical characteristics often shown 
and exhibited by conventional fossil fuels save for the combustion characteristics. They 
concluded that the biodiesel combustion process under LTC combustion strategies reduces 
NOX significantly due to reduction in the peak combustion temperature with the application of 
EGR % flow rate. However, there was increase in PM emission resulting from the unburnt 
hydrocarbons and the fuel impinged on the cylinder walls. These problems are associated with 
the difference in the fuel properties of the biodiesel fuel mixture. Other conclusions they 
reported are that biodiesel fuels had lower BTE and higher fuel consumption, therefore, poor 
BSFC, besides increasing CO and CO2 emissions.  
[21] in “Control of Combustion Process in an HCCI-DI Combustion Engine Using Dual 
Injection Strategy with EGR” conducted an experiment using internal and external EGR modes 
to realize fuel vaporization and in-cylinder temperature reduction of 50 %. EGR from the 
external mode was used in the compression stroke. The result indicated a significant increase 
in CO/HC emissions as the mixture seemed to suffer from the effect of almost total incomplete 
combustion. 
[22] in their experimental study “Control of Combustion Process in an HCCI-DI Combustion 
Engine using Dual Injection Strategy with EGR” compared HCCI and conventional diesel 
engine performance and emission characteristics. When 30 % EGR flow rate was applied, they 
noted that the combustion of HCCI produced ULN (ultra-low NOX) emissions, but produced 
no meaningful reduction in the emissions of soot and SOF. 
[23] in “Combustion Characteristics of Diesel HCCI Engine: An Experimental Investigation 
using External Mixture Formation Technique” used a twin piston DI diesel engine to conduct 
a study by means of the modification of one cylinder to use HCCI premixed mixture fuel and 
the other to operate on conventional diesel. During their study they varied their mixture 
concentration from 𝜆𝜆 = 4.96 (as the lean mixture level) or the misfiring limit and 𝜆𝜆 = 2.56 
(considered as the rich mixture level) or the knock limit. When a 20 % EGR flow rate was 
applied the results indicated increased combustion efficiency, especially when the value of 𝜆𝜆 ≤
3.70 with medium range engine loads being maintained. Moreover, when the engine load was 
increased to high levels it was observed that the values of lambda (λ) recorded at 𝜆𝜆 ≤  3.0 and 
20 
 
produced engine knock and a very rough unpleasant running combustion experience from the 
experimental engine. 
[24] in “Effects of Port Fuel Injection (PFI) of N-Butanol and EGR on Combustion and 
Emissions of a Direct Injection Diesel Engine” studied the effects of PFI on n-butanol with 
application of EGR % flow rate and how this influences emission and combustion 
characteristics in a diesel engine. In their study and experiment, they equipped their engine 
with independent inlet and outlet systems and varied the EGR % flow rate which they divided 
into three sections of 0 %, 15 % and 45 % and the premixed fuel of the n-butanol which they 
divided into four premixed ratios of 0 %, 20 %, 38 % and 47 % by volume. 
They observed that as the rate of EGR % flow increased to 45 % there was a significant 
decrease by 97 % of the total NOX emissions, although there was an exponential increase in 
soot emissions and SOFs. However, when they combined the use of the PFI and n-butanol of 
47 % while maintaining the EGR % flow rate at 45 %, the emissions of NOx and those of soot 
were drastically reduced by 88 % and 17 % respectively, thus achieving the traditional basic 
fundamental problem of the NOX versus soot trade-off point. 
[25] in their study on “Combustion and Emission Characteristics of a DME (Dimethyl Ether)-
Diesel Dual Fuel Premixed Charge Compression Ignition Engine with EGR (Exhaust Gas 
Recirculation)” studied the effects of external EGR and dimethyl ether (DME) on the premixed 
ratio and on the performance and emission characteristics of a two cylinder naturally aspirated 
diesel engine. During their experiment they varied EGR % from 0 % to 27 % while keeping 
the premixed ratio at between 0 % to 30 %. Their results indicated that an increase in EGR % 
mass flow rate reduced NOX emissions significantly, but increased CO2 and HC emissions 
considerably. However, a very good result was obtained with higher PR % and higher EGR % 
flow rates especially in regard to NOX and smoke emissions. 
[26] in “Potential of HCCI Concepts for DI Diesel Engines” investigated late combustion 
strategies of injection of HCCI combustion using the modulated kinetics type of combustion 
as comparatively experimented by [27], and observed that reduced NOX levels could be 
achieved with the application of a high EGR % flow rate of 40 % although the smoke level 
emissions increased. Retardation and late injection timing after top dead center (ATDC), the 
authors reported, led to a decrease in fuel efficiency, thus increasing brake specific fuel 
consumption (BSFC) with a considerable lineal increase in HC emissions. In order to mitigate 
this problem, the in-cylinder flow to the mixture gases was strengthened through the 
21 
 
optimization of the geometrical design of the intake port, thus allowing them to achieve their 
objectives of reducing NOX using a high EGR flow rate of 40 %.  
2.3 The Effect of EGR On the Role of Exergy Destruction 
[28] in “Effect of Exhaust Gas Recirculation (EGR) on the Performance Characteristics of a 
Direct Injection Multi Cylinders Diesel Engine” found that there is a relationship between pre-
mixed turbulent combustion with the second law of thermodynamics and lost exergy in 
irreversibility and entropy generation. The authors employed the low swirl combustion (LSC) 
technique for flame stabilization. Exergy analysis is defined by [29] as a technique that uses 
the conservation of mass and the conservation of energy in combination with the second law 
of thermodynamics for design analysis and improvement of energy and other systems. 
In their experimental set-up they used a swirler, which consisted of an outer annular swirled 
section and a center channel for flow retardation (this is a key component) with an attachment 
of a perforated plate with 16 vanes that aids the center channel to control the flow rate that is 
allocated to each plate. Another key component in this design is the LSB swirler recess distance 
which controls the time of interaction between the swirled and unswirled mixture flows 
parameters with an effect on flow divergence which affects the performance of the LSB 
component and its operation (Figures 2.4 and 2.5). 
 
Figure 2.4: The LSB schematic set-up system and the main components 
 
 
Figure 2.5: The geometrical parameters of the LSB unit as used in the experimental set-
up  
22 
 
 
Figure 2.5 shows the mechanical phenomenon characteristics of geometric, kinematic, and 
dynamic aspects which affect the swirler combustion performance according to [30]. These 
factors are as follows: the burner radius(𝐸𝐸𝑏𝑏), unswirled core radius (𝐸𝐸𝐶𝐶 ), performance rated 
plate radius, the vane angle (𝛼𝛼), vane numbers, blockage of perforated plate, recess length (𝐿𝐿), 
which are summed up as the geometrical parameters and are considered to be very important. 
After the set-up study and experimentation, the following conclusions were reached. The 
exergy analysis conducted on the LSB running on NG fuel resulted in two distinguished flame 
regimes, the lifted flame and the attached flame. When the normalized laminar flame speed 
(Vs) velocity of 10.3 and 17.1 was reached it was observed that a lifted flame type came up, 
while when the NLFS velocity of 31.9 and 42.9 was reached the flame observed was the 
attached flame type. Thirdly the exergy of heat transfers between the flame and the combustor 
walls in the lifted flame type was observed to be greater and more significant compared to the 
values of the attached flame type, although they observed via the exhaust gases released that 
the exergy transfer had minimal increment in the attached flame regime when the same was 
compared to the lifted flame regime.  
The value of the irreversibility ratio (Irr) for the diffusion flames was found to be very large, 
while the irreversibility of the premixed combustion was 35.5 %. However, when the flame 
regimes were changed the authors reported that the irreversibility decreased as the flames were 
being altered by a value of 11 %. The Irr was observed to receive significant contribution from 
the flame regime local temperature gradient, because temperature gradients are stronger when 
the values of Vs are higher.  
Figure 2.6 and Figure 2.7 [28] show the relationship between flame temperature and the 
combustor, with Figure 2.9 indicating the temperature range of the combustor during the 
experiment. The lifted flame showed a better heat transfer of 14 % more between the 
combustor’s combustion chamber and the surroundings environment compared to the attached 
flame. 
  
23 
 
 
Figure 2.6: Variation of temperature versus flame radius length of lifted and attached flame  
 
 
Figure 2.7: Variation of temperature versus flame radius length and flame progression of lifted and 
attached flame 
 
 
Figure 2.8: Variation of the temperature range for the two types flame progression regimes shown in 
Figures 2.7 and 2.8. 
 
24 
 
2.4 The Effects of EGR On Emission and Combustion Characteristics of Biodiesel 
In mentioning and discussing the combustion models it is obvious that the development of 
alternative fuels is of paramount importance. There are several methods that have been 
proposed for processing alternative fuels also called biodiesels or biofuels, the common 
methods being transesterification, hydrogenation, pyrolysis (used in plastics) and dilution. The 
use of alternative fuel for emission control has been increasing, particularly with the use of the 
EGR control technique to control and reduce NOX, HC, CO, and soot emissions. A number of 
authors have carried research in this area and this review will present some of the studies and 
experimental work conducted and the current direction in research and development.  
[31] in “An Experimental Investigation on a DI Diesel Engine using Waste Plastic Oil with 
Exhaust Gas Recirculation” showed that plastic oil produced the most NOX emissions but 
showed reduced NOX emissions with increasing EGR of 20 % onwards. However, in the 
smoke, CO and HC levels of emissions there was no noticeable change. The authors concluded 
that with the application of increasing EGR % flow rate levels, plastic oil can be used as an 
alternative biodiesel fuel. 
[32] in the “Effects of Exhaust Gas Recirculation on the Performance and Emission 
Characteristics of a Diesel Engine Fueled with Waste Cooking Oil Methyl Ester” used a single 
cylinder four stroke water cooled DI compression ignition engine, with a dynamometer, 
coupled to an output shaft. Modifications were carried out to allow the experiment to be carried 
out. In order to collect data a digital panel was installed, while an exhaust gas analyzer was 
used to investigate emission characteristics of the various gases in this study, namely, CO, CO2, 
HC, NOX and O2. The test fuel blends they used were BO = (normal diesel) blended with waste 
cooking oil methyl ester (WCOME) where B10 = 10 % WCOME + 90 % diesel, B20 % 
WCOME + 80 % diesel by volume, B30 % WCOME + 70 % diesel by volume. The EGR % 
mass flow rate ranged from 0 % to 20 % graduated in intervals of 5 %. All the tests were 
conducted at constant engine speed of 1500 rpm. The following parameters were investigated 
and their variations: BSFC, BTE, EGT and emission characteristics of NOX, HC and CO. 
During their study they observed that the BTE remained unaffected by the application of EGR 
% flow rate that was above 15 %, but tended to decrease as the EGR % mass flow rate increased 
beyond 15 % EGR. This drop is explained from the fact that the availability of oxygen was 
reduced during combustion with the application of a EGR of 15 %. The brake specific fuel 
consumption (BSFC) was observed to be independent of the rate of EGR % flow rate, 
25 
 
especially at lower rates of EGR % flow, but it increased as the rate of EGR % flow rate 
increased above 15 %. This phenomenon is explained by the fact that there was formation of a 
rich mixture due to the dilution effect in the combustion chamber, this being identical to the 
findings of [33]. 
The exhaust gas temperature (EGT) was observed to be lowered if the engine was running on 
cooled EGR compared to when the engine exhaust gases were not cooled. The cooling of the 
EGR gases affects the heat specific ratio of the intake air mixture, by causing less oxygen in 
the combustion chamber. The NOX emission findings indicate that for all the fuel blends used 
during the study, the NOX emissions decreased with the increase in EGR % flow rate. This can 
be explained by two factors: reduction in the availability of oxygen during combustion, and the 
lower combustion temperatures (NOX is temperature dependent for its formation so will 
decrease as flame temperatures drop). The HC and CO emission showed an increase with the 
increase of the EGR % flow rate above 15 %, as less oxygen becomes available for combustion, 
resulting in incomplete combustion and more HC and PM emissions.  
[34] in “Experimental Analysis of Exhaust Gas Recirculation on Diesel Engines Operating with 
Biodiesel” used a four stroke, water cooled engine, single cylinder DI, rated at 5.2 kW power 
with 1500 rpm rated speed. The experiment was conducted in two phases. The first phase 
investigated the performance, combustion and emission of the diesel fuel with biodiesel blends 
of B20 where the EGR application was zero. The second phase of the experiment was 
conducted using the same equipment and test rig with fuel blend of B20 but under EGR % flow 
rate application. The range of distribution for EGR % flow rate was 5 %, 10 %, 20 %, while 
the load was distributed into 20 %, 40 %, 60 % 80 % and 100 %.  
The tests and study found that the BSFC for the test fuel without EGR under full engine load 
was seen to be 0.2779 kg/kW-hr and 0.2794 kg/kW-hr for the biodiesel respectively. The values 
of the BSFC at full engine load and EGR % flow rates of 5 %, 10 %, 15 % and 20 %, were 
0.2853 kg/kW-hr, 0.2796 kg/kW-hr, 0.2832 kg/kW-hr and 0.3050 kg/kW-hr respectively, 
whereas for higher EGR % flow rate of 20 % the BSFC increased for both diesel and biodiesel. 
This was attributed to the high calorific values (CVs), high viscosity, high density and boiling 
point. The BTE for the biodiesel was found to be comparable to that of diesel with or without 
EGR % flow rate.  
The variations between the hydrocarbons and the brake power (BP) results and conclusions 
indicate that with an increase in EGR % flow rate levels, the HC emissions also increased only 
26 
 
for the biodiesel, attributed to the presence of high oxygen content in most biodiesels, thus 
compensating for the lack of oxygen to facilitate complete combustion thereby increasing BTE. 
The values for the NOX during the study were found to be 736 ppm for diesel fuel running 
without EGR in a full load environment and 796 ppm for biodiesel at full load without EGR. 
However, with the application of EGR it was noticed that the levels of NOX reduced to 157 
ppm for diesel and 158 ppm for biodiesel respectively, but a reduction in BTE was established 
with large increments in smoke density. 
Carbon monoxide (CO) in relation to diesel and biodiesel with various levels of EGR % flow 
rate and load condition was noted to vary from 0.14 % (by volume) for the diesel to 0.18 % (by 
volume) for the biodiesel all at full engine load conditions. The authors concluded that as the 
EGR % flow rate increased so did the CO emission increase, caused primarily by the lack of 
oxygen during the combustion process. The smoke density was generally found to be lower or 
decreased for biodiesel compared to diesel fuel and this was observed for all engine loads. As 
the EGR % flow rate increased there was an observed increase in the smoke density. This is 
explained by the fact that EGR affects the oxygen availability leading to incomplete 
combustion and the formation of more PM. 
The EGT was observed to increase with increase in load, being more noticeable for biodiesel 
than for diesels at all engine applicable loads and conditions. This increase in temperature 
during this experiment was attributed to higher oxygen content in all biodiesel fuels particularly 
those used in the study. This study finding was identical to [35]. The cylinder pressure at full 
engine loads with EGR at zero was observed to be comparable to biodiesel, arising from the 
fact that there was a good mixing from the biodiesel mixture under high load, because EGR 
gases act like heat absorbing agents. Combustion analysis revealed significant HRRs, with or 
without the application of EGR % flow rate. 
[36] in “Effect of EGR on Performance and Emission Characteristics of a Dual Fuel Engine 
Fueled with CNG and JOME” studied the dual fuel engine mode which has been identified as 
one of the most promising combustion techniques available today. The experimental set-up 
used a Kirloskar engine model CAF 8 rated at 5.9 kW/h @1500 rpm. In order for the 
homogeneous mixture of compressed natural gas (CNG) and air to be introduced into the 
cylinder, a venture type gas mixture was used. The fuel injection system pressure was not 
changed but remained at 200 bar and injection timing at 23˚ BTDC. The engine was then 
coupled to a dynamometer, with other necessary modifications done to enable the engine to be 
27 
 
used during the experiment. In this experiment two different types of fuels were used 
simultaneously, one as a primary fuel and the other fuel as a pilot fuel. A mixture of the two 
was then introduced into the combustion chamber during the intake stroke, at different phases. 
The homogeneous mixture could not ignite automatically due to the high self-ignition 
temperature, thus requiring injection of the pilot fuel to ignite the homogeneous mixture. This 
then was a hybrid ignition model of SI and CI engine types. 
After testing and evaluation, the BTE was observed to be dependent on the combustion quality 
produced in the combustion chamber. The BTE for conventional diesel was shown to be 
relatively higher than the dual fuel, especially in low engine load environment, explained by 
the high self-ignition temperatures exhibited by the CNG in addition to the its poor flame 
propagation qualities. There was better utilization of the CNG fuel due to the ignition delay as 
the engine load was increased, tending the fuel-air mixture towards the richer side, which 
increased the combustion temperature in the cylinder. At high engine loads the authors reported 
an improved BTE equaling that normally associated with conventional diesel fuel propelled 
engines. When EGR % flow rate was employed, the dual fuel showed minimal increment at 
lower engine loads, but at higher loads with application of EGR % flow rate increasing, the 
BTE decreased. The application of EGR made the mixture of the dual fuel richer thus leading 
to poor utilization hence decreased values for the BTE. This outcome was identical to the study 
of [37, 38].  
The brake specific energy consumption (BSEC) or brake specific fuel consumption (BSFC) 
during the study was found to be higher for conventional diesel fuel compared to the dual fuel 
at low and intermediate engine loads. This is explained by the relatively poor utilization of the 
gaseous fuel in the dual fuel mode at low engine loads up to intermediate engine loads. 
However, as the engine load increased there was better utilization of the high CV of the dual 
fuel CNG, thus the BSEC increased compared to conventional diesel. The application of EGR 
showed improved BSEC at low engine loads as reburning of the unburnt hydrocarbons 
occurred, but as the load was increased the BSEC decreased, explained by the richer mixture 
and drop in combustion temperatures. 
Figure 2.9 shows the variation of UHC with engine load for diesel fuel, CNG and blends of 
Jatropa oil methyl ester (JOME) operating under EGR [36]. 
28 
 
 
Figure 2.9: UHC vs Load 
 
The emissions of UHC or the un-oxidized particles of fuel showed that at low loads the dual 
fuel performed poorly in utilization of fuel, but as the load increased, the UHC for the dual fuel 
was noted to be higher and increasing s compared to conventional diesel fuel for all the 
remaining load conditions tested. This is explained by the overlapping valves between the 
exhaust and the inlet valves respectively and the flammability limit of the CNG. The results of 
this study were identical to those of [39, 40].  
UHC emissions are noted to increase with increasing engine load, but at high engine load the 
dual UHC emission continued to decrease, a phenomenon also identified and observed by [41] 
as in Figure 2.9. [36] reported that the use of EGR with the dual fuel mode reduced the volume 
of UHC emissions especially during low and intermediate engine loads. This is explained from 
the unavailability of O2 due to the displacement caused by the re-introduction of the exhaust 
gases back to the combustion cylinder, thus burning the UHC, hence reducing the unburnt 
hydrocarbon (UHC) emissions. However, higher engine loads mean less availability of O2 
resulting in incomplete combustion and poor oxidation of the induced charge, hence an increase 
in the UHC emissions for the dual fuel with the application of EGR % flow rate. 
The NOX emissions gases, which depend on combustion temperature and the availability of O2 
for their formation, indicated an increase for conventional diesel fuel compared to the dual fuel 
at low and intermediate engine loads. This factor is explained by the high in-cylinder 
temperatures associated with conventional diesel combustion, which is an identical result to 
the study by [42]. The dual fuel did not show any difference from the conventional diesel fuel 
29 
 
results mentioned above. However because biodiesels contain more O2 content (in this case 
JOME), there was a considerable increase in NOX emissions, being an identical outcome to the 
studies conducted by [43]. The application of the EGR % flow rate with the dual fuel mode 
was seen to reduce the NOX, emissions for all the engine load conditions and environment, 
being identical to the study findings of [38]. The application of EGR % flow rate dilutes the 
inlet intake mixture, thus increasing its SHC hence reducing in-cylinder temperature, therefore 
leading to the suppression and reduction of NOX emissions. However, it should be remembered 
that increasing the EGR % flow rate reduces NOX emissions but also decreases BTE. 
The CO emissions during the study on the dual fuel showed very poor fuel utilization and 
incomplete combustion, explained by the low in-cylinder combustion temperature and the 
limited availability of O2. Carbon emissions from dual fuel are higher than from conventional 
diesel fuel for all the engine operating loads except full engine load. The application of EGR 
% flow rate enhanced complete combustion, hence the decrease in CO emissions at low and 
intermediate loads. However, at high engine loads the richer mixture reduces the O2 
concentration in the combustion chamber thus reducing CO emissions for the dual fuel mode 
application. Smoke emissions were observed to increase with increasing engine load for all the 
operating experimental modes. This was explained by poor utilization of fuel and incomplete 
combustion during the dual fuel mode operation. The application of EGR increased smoke 
emissions for all the engine loads that were tested, due to the decreased levels of O2 in the 
combustion chamber. 
[44] In the “Effects of Varying EGR Rates in the Performance and NOX Emission of an In-
Direct Diesel Injection (IDI) Diesel Engine Fueled With JB100, JB80, JB60, JB40, JB20 And 
Diesel” used a single cylinder engine, water cooled in direct diesel injection (IDI) diesel engine, 
connected to an electrical loading system.  
After conducting the study, they observed that at 15 % EGR flow rate, blends of JB20 at 25 % 
EGR rate flow, JB40 at 15 % EGR flow rate, JB60 at 20 % EGR flow rate, JB80 at 40 % EGR 
flow rate, and JB100 at 5 % EGR flow rate. Their NOX emissions effectively reduced by 10.1 
%, 11.94 %, 13.4 %, 15.2 %, 19.8 % and 24.8 % respectively. However, soot emission and 
PM, CO, and CO2 increased. The biodiesel blends all showed high levels of NOX emission 
explained by the fact that biodiesel blends contain higher amounts of oxygen and can be seen 
at all full loads in the diesel blends with values of 882 ppm, 848 ppm, 806 ppm, 775 ppm, 737 
ppm respectively, whereas for diesel only the value is 643 ppm.  
30 
 
The BTE was observed to be lower for all the JB blends compared to diesel due to the lower 
CV of the biodiesel but the viscosity index was seen to be very high. During the study diesel 
and the JB20 blend of fuel showed a better BTE value among all the blends used in the study 
when 5 % EGR flow rate was applied, while the JB90 blend of fuel showed the maximum 
reduction of NOX percentage but with decreased BTE of 1.25 %. The decrease in BTE during 
this study was attributed to less availability of oxygen in the fresh charge which is an effect 
from the EGR application, where the addition of exhaust gases resulted in lower flame 
temperatures and velocity leading to an increase in soot and PM. However, JB60 test fuel with 
a 20 % EGR flow rate reduced NOX emissions to 19.81 % with a slight increase in the BTE of 
0.77%, attributed to the re-burning of the UHCs entering together with the EGR % flow rate 
and helping to complete the combustion of the fuel.  
[45] in “Performance Analysis of a Single Cylinder Diesel Engine Using Jatropha Oil with 
EGR” investigated the effects of biodiesel from Jatropha oil on the performance and emissions 
characteristics of a diesel engine using a single cylinder TV1 Kirloskar model engine, four 
stroke, water cooled and with a rated power of 3.68 kW at 1500 rpm. 
After testing and experimentation the observation was made that the BTE decreased with an 
increase in EGR flow rate from 5 % to 10 %. The BSFC for all the blends of the Jatropha 
increased compared to conventional diesel with an EGR flow rate of 5 % to10 %. The B20 
blend of Jatropha oil showed the most improved result for BSFC and total fuel consumption. 
When the blends of B10, B20 and B30 were used without the application of EGR % flow rate, 
there was an increase in BSFC which was not clearly noticed with EGR flow rate application. 
Therefore, the authors concluded that CO emissions increased as the rate of EGR is increased, 
while the NOX emissions decreased with application of EGR in all the Jatropha blends. 
[46] in “Emission Reduction Using Hydro Treated Vegetable Oil (HVO) With Miller Timing 
and EGR In Diesel Combustion” studied the effects of EGR on hydro-treated vegetable oil 
(HVO) in reduction of emissions in relation to combustion parameters in a diesel engine. The 
Miller cycle test included three phases of intake valve closing timing and an advanced (SOC) 
system, which showed a considerable reduction in the NOX and the smoke emissions. Their 
study reported a marked reduction in the emissions of NOX and smoke, thus concluding that 
compared to conventional diesel EN590, the hydro treated vegetable oil (HVO) fuel blend can 
allow the use of higher EGR % flow rates without compromising on PM, UHC and smoke 
emissions but with an exponential reduction in NOX emission 
31 
 
2.5 Effects of EGR on Performance, Combustion and Emission Characteristics of 
Diesel Engines (DI/IDI) 
[3] in the “Effect of Exhaust Gas Recirculation (EGR) on the Performance Characteristics of a 
Direct Injection Multi Cylinders Diesel Engine” used a direct DI, four-cylinder Fiat engine, 
coupled to a dynamometer and thermocouple with nickel chromium (Ni-Cr) and nickel alumide 
(Ni-Al) elements, air and fuel flow meters, together with engine speed tachometers. The test 
fuel was Iraqi conventional diesel CN = 46.8 under varying EGR % flow rates of 5 %,10 %, 
20 %, and 30 %. All engine load was maintained at variable speeds of 1250 rpm to 3000 rpm.  
EGR % flow rate at increased rates increased the BSFC, while at low engine load ranges it was 
found to be high compared to when medium engine load ranges were used. This was attributed 
to the drop in combustion chamber temperature. However, reducing the temperature increased 
the fuel delay period, thus reducing engine peak pressure resulting in lower engine outputs and 
increased fuel consumption. [3] observed that at low engine loads diesel engines operate on 
lean mixture and their volumetric efficiency is very high, but increasing the engine load 
consumes more fuel in order to meet the load demands as the engine motor speed is increased. 
However, in increasing the load and engine speed the EGT increases too, while increasing the 
EGR % flow rate reduces the temperatures, as can be seen in Figure 2.10.  
 
Figure 2.10: Variation of exhaust gas temperature with brake mean effective pressure under different diesel 
test fuel with application of EGR % rate flow [3] 
 
[47] in “Effects of Compression Ratio and EGR on Performance, Combustion and Emissions 
of DI Injection Diesel Engine” Studied the effects of cooled EGR on the performance, 
combustion and emission of a variable compression ignition diesel engine with various 
compression ratios under different engine loads with varying EGR % flow rates. The output 
32 
 
parameters investigated included indicated specific fuel consumption (ISFC), BTE, HRR and 
mass fraction of the burnt mixture and the elements composing the gases in the exhaust 
emissions using a single cylinder DI engine, variable compression, high speed, with rated 
power of 5 HP at1500 rpm.  
During their experimental study the speed was maintained constant using different CRs and it 
was reported that irrespective of the CR used, the volumetric efficiency of the engine remained 
the same, while the mechanical efficiency decreased as the engine CR values of 15 and 16 were 
reached. This factor is explained as being due to the increased loss of friction power. When the 
CR value was 19 the BTE reached a maximum of 17 %. The increase in CR was observed to 
decrease the ISFC irrespective of the engine load applied due to two factors, namely, the 
ignition delay and the long combustion duration period. The HRR at a recirculation of 5 % and 
a 1/4 engine load condition was noted to occur where crank angle duration was 37 ⁰ CA of the 
total crankshaft angle rotation. The increase in the engine load to 1/2 saw a minimal increase 
in the combustion duration period from 39 ⁰ CA to 42 ⁰ CA @ 3/4 engine load and 46 ⁰ CA @ 
full engine load respectively. These increments in the combustion duration time increased the 
HRR, as there was complete combustion in the combustion chamber. 
. 
Figure 2.11: Variation of burnt fraction mass with crank angle at different varying engine load conditions 
[47] 
 
The burned fraction mass in relation to the crank angle was reported to have no combustion at 
360 ⁰ CA but at 361 ⁰ CA, and the whole combustion duration period lasted until the 393 ⁰ CA 
33 
 
mark making the duration last for a period of about 32⁰ CA. However, as the engine load was 
increased they observed an increase in the combustion duration period, thus achieving a 
burning rate of 70 % of the total mass fraction input in the combustion chamber. This they 
reported happened within the first 20 ⁰ CA rotation with the remaining balance being burned 
in the remaining combustion period duration cycle of 32 ⁰ CA as in Figure 2.11. 
The emission characteristics of NOX and smoke opacity in relation to the CR under different 
engine varying loads showed that an increase in the engine load increased the NOX emissions 
as more heat was released as more gases in the combustion chamber pocket disassociated. 
However, with the application of the EGR % flow rate at increasing mode the NOX emissions 
reduced. This is due to low peak temperature and complete combustion of the inlet intake and 
recirculated EGR gases which are reburned, hence the reduction. At CR 15 % and EGR % flow 
rate of 10 % there was a reduction of 26.5 % and at full engine load a reduction of 62.5 %. The 
smoke opacity showed continuous increase as the engine load was increased in all the test 
levels, a factor the authors linked to the increased amount of fuel injected to the combustion 
chamber. However, the smoke opacity in all the CR levels tested showed a decrease from a 
high of 17 % to a low of 4 %. 
[48] in their study “Combustion and Emissions of Compression Ignition in a Direct Injection 
Diesel Engine Fueled with Pentanol” selected from a four-cylinder engine a single cylinder as 
the test cylinder while maintaining the other three cylinders as experiment control cylinders 
rated at 4000 rpm. The electronic control module (ECU) was modified to provide flexibility in 
manipulating injection parameters. The tests were carried out in normal engine operating 
temperatures of 85 ⁰C. The test fuels used were conventional diesel as the control fuel or the 
baseline fuel, with three different biodiesel fuels employed and pentanol, as seen in Table 2.1. 
  
34 
 
Table 2.1: Baseline diesel test fuel and various pentanol blends of test fuels and their properties adapted 
from [48] 
Property Unit Diesel Pentanol Biodiesel D70B30 D70P30 D40P30 
Viscosity mm/s 41.27 289 7.159 5.03 3.2 3.9 
@ 20 ⁰C 
Low MJ/kg 42.68 35.66 38.29 41.38 40.42 39.09 
heating 
value 
Oxygen % 0 18.15 11.1 3.3 5.5 8.8 
content 
Latent KJ/kg 270 308 258 - - - 
heating 
@ 25 ⁰C 
Cetane - 56.5 20-25 61.7 - - - 
number 
Density Kg/m3 830.4 815 871.4 842.7 825.8 838.1 
@ 20 ⁰C 
Surface 10-3 27.5 24.7 30.3 28.2 26.2 26.9 
tension Nm-1 
@ 20 ⁰C  
Boiling ⁰C T10-223 138 T10-323 T10-230 T10-135 T10-135 
point T50-266 T50-326 T50-280 T50-240 T50-264 
T90-311 T90-335 T90-314 T90-298 T90-310 
 
 
The results on the combustion characteristics of the test fuels demonstrated that any addition 
of pentanol increased atomization and offered easy mixing of the fuel vapor with inlet air intake 
charge at the same testing conditions, thus leading to higher HRR values during the first stage 
of the combustion phase process as demonstrated by Figure 2.12. This is defined as the CA 
interval between the heat release position at CA 5 % and CA 90 % during the main combustion 
phase process. Thus, with the increase in cylinder temperature after the first stage of the heat 
release there is a reduction and weakening of the chemical effect ignition delay, while the 
physical delay due to low viscosity and the high viscosity of the test fuel additive pentanol play 
35 
 
a significant role on the auto ignition of the mixture. Hence, there was no significant difference 
among the four test fuels based on the ignition delay mentioned above, though each had a 
different cetane number (CN). 
 
Figure 2.12: Variation of in-cylinder pressure with crank angle change in relation to heat release rate [49] 
 
From Figure 2.12 it is evident that the combustion process phase is shortened with the 
introduction of pentanol blends as there is marked improvement in the mixing rate of the inlet 
intake air and fuel blends during the test with smaller values of HRR below CA50 as compared 
to the conventional diesel and the D70B30 pentanol blend of fuel.  
Turning to the performance of the pentanol fuels in terms of emission characteristics, the 
authors report that when supercharging was applied to the test engine there was considerable 
reduction in the emissions of NOX, CO2, CO, THC and soot emissions for all the test fuels that 
were used as compared to when the engine was naturally inducted. However, as the load 
increased there was a minimal increase in the NOX emissions, but a decrease was observed as 
the load continued to increase. This is due to the fact that NOX formation is temperature and 
equivalence ratio (ER) dependent and does very well in the reduction of NOX because of its 
higher latent heat of vaporization (meaning lower combustion temperature). The low CN 
results in longer ignition delay, hence increase in the burned fraction mass of fuel during 
premixing with the inlet intake air mass. It should be mentioned here that the only factor that 
is attributed to the decrease of NOX in the pentanol based fuels during low engine load 
conditions is the latent heat of vaporization. The high flame temperatures and high levels of 
the oxygen content in the pentanol fuels lead to increased NOX emissions being identical to the 
study of [50]. 
36 
 
The indicated thermal efficiency (ITE) was reported to decrease with an increase in the engine 
load, explained by the high ER values, longer combustion durations experienced, and the 
higher fuel efficiency as exhibited by the blended fuels of pentanol. The corrected ISFC showed 
an increase as the engine load increased, thus providing a reversal of the aforementioned 
observation for ITE as indicated by Figure 2.13. It should be mentioned here regarding the 
ISFC, that the D70B30 and conventional diesel fuel produced a result of minimal difference in 
terms of the values generated in relation to engine loads in all the test conditions shown in 
Figure 2.14, leading to the conclusion that any addition of an alcohol to any biodiesel does not 
affect the thermal efficiency. 
Figure 2.13 shows the relationship between ISFC and IMEP of the various pentanol fuel blends 
and the diesel test fuel [49]. 
 
Figure 2.13: ISFC VS IMEP 
 
Figure 2.14 shows the variation of the ITE and IMEP in relation to diesel and pentanol blends 
[48]. 
 
Figure 2.14: ITE VS IMEP 
37 
 
The soot emissions show a similar behavior to the NOX emissions characteristics, although 
there are some slight variations. As the engine IMEP is increased from 0.5 Mpa to 0.09 Mpa 
with more fuel being injected into the combustion chamber, the authors report a reduction in 
the duration of the ignition delay leading to a more diffuse combustion and increase in the soot 
emissions. This can be attributed to the fact that the ER increases with high engine loads. All 
the oxygenated blends of pentanol showed decreased and suppressed soot emissions in all the 
engine operating and testing conditions. This was identical to the study conducted by [51].  
On the emissions of HCs, they observed that as the engine load increases, there is a marked 
reduction in the emissions of HCs in all the test fuels used during the study. This factor is linked 
to the high in-cylinder temperature as the engine load increased. Other factors which they 
observed as contributing to the decrease in the HC emissions include low CN and the latent 
heat of vaporization exhibited in the D70P30 pentanol fuel blend, longer ignition delay, broad 
lean combustion zones, high oxygen content and the presence of some biodiesels which have 
high CNs among some of the pentanol fuel blends used in the experiment. 
[52] in “Combined Effects of EGR and Varying Inlet Pressure on Performance and Emission 
of Diesel Engine” investigated the combined effect of EGR and varying inlet intake pressure 
on the emission and performance of a diesel engine using a single cylinder, DI, water cooled 
diesel engine with a rated power of 5.0 HP or 3.72 kW at 1500 rpm. Necessary modifications 
were mounted to allow the experiment, including a manually operated EGR control valve 
system and asbestos insulation for the exhaust pipe line in order to maintain and reduce heat 
loss. 
After testing and study the authors observed and concluded that increasing the inlet intake air 
pressure and EGR % flow rate improved overall engine performance, with an increase in the 
BTE and decreased BSFC. Increasing the intake inlet air pressure and applying varying EGR 
% flow rates was observed to reduce NOX emissions due to decreased EGT temperatures. 
Figure 2.15 shows the variation of BSFC at varying EGR rates with engine load at 50 % under 
the influence of intake inlet pressure [52]. 
38 
 
 
Figure 2.15: BSFC VS EGR % flow rate 
 
The increase in the intake inlet air pressure and the increase in the rate of EGR % flow rate was 
noticed to decrease the BSFC as the density of air increased and more combustion occurred 
due to availability of oxygen with the increased inlet air flow as shown in Figure 2.15.  
[16] in “Experimental Investigation of Performance and Emission Characteristics by Different 
Exhaust Gas Recirculation Methods used in Diesel Engine” studied the effects of hot EGR and 
cold EGR on the performance and emission characteristics of a DI diesel engine. Their test 
engine was a single cylinder, DI, water cooled CI diesel engine, with a rated capacity power of 
6.5 HP or 4.836 kW at 1500 rpm. Various modifications were made to allow study. 
Figure 2.16 shows the variation in relationship between SFC and BP with different types of 
EGR % flow rates with diesel test fuel [16]. 
 
Figure 2.16: Figure 17 SFC VS Brake Power 
 
39 
 
Fig 2.17 shows the variation between the BTE and BP with varying EGR % types and rates 
with diesel fuel [16].  
 
Figure 2.17: Figure 18 BTE VS Brake Power 
 
The tests for performance and emissions were conducted without EGR and with EGR of 
different flow rates of 10 %, 15 %, 20 % and finally 25 % at 1500 rpm and varying engine 
loads. The specific fuel consumption (SFC) was noted to be high during all engine loads tested 
with or without EGR. At 10 % EGR rate flow the BTE equalled the values similar to when 
there was 0 % application of EGR. However, the indicated thermal efficiency when cold EGR 
was applied during the experiment showed very good results as compared to when hot EGR 
was used. The NOX emissions from cold EGR indicated a significant decrease as compared to 
the hot EGR gases application. At 10 % cold EGR the emissions of CO2 were seen to be very 
high as compared to the hot EGR, increasing at part load but decreasing with higher engine 
loads as shown in Figures 2.16 and 2.17.  
[53] in “ Combustion and Emission Characteristics of High N-Butanol and Diesel Engine and 
EGR Impact” studied the effects of EGR on n-butanol in relation to combustion and emission. 
The experiment was conducted using a modified four stroke, water cooled engine, six-cylinder 
research engine with a rated power of 167 kW at 2500 rpm. In order to reduce cylinder 
interaction one cylinder was isolated and allowed independent intake and exhaust systems, with 
a common rail injection system controlled electronically by an open circuit control unit.  
The tests fuels chosen were conventional diesel as the base fuel, N-butanol as the alternative 
fuel in mixtures of 0 % and 40 % by volume as blends Bu00 and Bu40 corresponding to the 
percentages respectively of the blended mixtures. After their study and experimentation, the 
40 
 
authors concluded that the Bu40 blend of fuel showed a higher engine cylinder pressure, with 
longer delays in ignition and higher rates of combustion compared to the blend of fuel of Bu00. 
However, with the application of EGR flow rate of more than 30 % the cylinder pressure, 
ignition delay and rate of combustion decreased. There was no significant change in the 
emission of the HC, CO and the indicated thermal energy until very high EGR % flow rates 
were applied. However, the Bu40 blend showed change for the named factors in lower EGR % 
flow rates than other blends that were under consideration. Figure 2.18 shows that the 
relationship as aforementioned and discussed by [53]. However, as the EGR % flow rate was 
increased to higher levels the authors noticed a decreased ITE for Bu40 blend of fuel as 
compared to the Bu00 blend of fuel, but with the application of EGR % flow rate there was a 
decrease in the emissions of NOX although with an increase in soot emissions. The authors also 
reported a significant decrease in NOX emissions for the Bu40 blend mixture, with minimum 
increment in soot emissions, besides achieving very good high levels of ITE. 
Figure 2.18 shows the variation of the ITE, CO % and HC in relation to the varying EGR % 
flow rates [53].  
 
Figure 2.18: ITE %, CO (%) and HC (ppm) VS EGR % flow rate application 
The emissions of CO, the authors observed, depended on the production and the oxidation 
processes. However, the Bu40 was observed to have lower CO formation for the early 
combustion phase, but for the late phase of combustion process Bu40 showed increased and 
high levels of CO emissions. 
2.6 Effects of EGR On CNG and GTL On Performance Combustion and Emissions  
[5] in the study “Performance, Combustion and Emissions a Diesel Engine Operated with 
Reformed EGR Comparison of Diesel and Gas to Liquid (GTL) Fueling” used a Lister-Pitter 
TR1 engine, single cylinder, DI, air cooled and naturally aspirated. It was connected to an 
electric dynamometer with a motor and load cell. The crank shaft position timing was done at 
41 
 
22° CA (BTDC) in all the tests carried out. The EGR % flow rate was managed manually using 
a gate valve while EGR levels were determined by volume. The rig also included other 
instruments to do measurements for oil and air inlet/exhaust manifold temperature. The test 
fuels used for the test were ultra-low sulphur diesel (ULSD) and the ultra-clean GTL which 
was obtained through a low temperature Fischer-Tropsch (LTFT) process.  
The experiment was conducted in the two engine speed modes of 1200 rpm and 1500 rpm and 
two loads of 25 % (low load) and 50 % (medium load), while the EGR flow rate ratios chosen 
were classified into standard EGR and reformed exhaust recirculation (REGR) both using 
ratios of 0 %, 10 %, 20 %, and 30%. The REGR was a simulation based on H2 and CO in the 
ratios of 25 % EGR rate flow at H2 0 % and CO 75 % while in the second scenario 15 % EGR 
rate flow at H2 10% and CO 75 %. 
The conclusions from this study were as follows: at low engine loads the GTL fuel was found 
to have high BTE values as compared to the BTE values shown by ULSD, but when REGR 
was introduced it was observed that GTL had a higher fuel replacement than ULSD, thus 
showing an improvement in efficiency and reduced emissions. BTE was also found to be load 
dependent, i.e.at low engine loads efficiency decreased with increasing REGR. However, when 
the loads were high it was observed the high flame temperatures and velocity of hydrogen gas 
and the rapid expansion increased the BTE as the REGR was increased. 
Regarding HRR, the high octane number of GTL resulted in less pronounced premixed 
combustion, leading to lower HRR and low in-cylinder pressure, leading to a reduction in NOX 
emission where GTL fuels were employed unlike when ULSD fuel was used in the experiment. 
Whereas the combustion pattern was affected by the REGR, the use of the 30 % REGR brought 
efficient gas fuel combustibility and an increase in cylinder pressure together with the ROHR. 
The authors noted that an increase of REGR resulted into a decrease of both NOX and smoke 
emissions. However, at higher engine loads, higher REGR resulted in a decrease in smoke 
emission but compromised the NOX emissions. Lastly when the combination was used for both 
REGR and GTL fuel, the results showed a marked improvement in the reduction of emissions. 
The 30 % REGR when combined with the GTL fuel produced a 60 % decrease in the smoke 
emission, while the NOX emission showed a 75 % decrease as compared to when ULSD fuel 
was used at 0 % REGR and low engine loads.  
[38] in “Combustion and Emission Characteristics of a Natural Gas Fueled Diesel Engine with 
EGR,” used a Lister-Petter PH1W single cylinder, naturally aspirated, four stroke, water 
42 
 
cooled, and high speed DI, with a bowl-in-piston combustion design. The engine was modified 
to suit dual fuel operation with natural gas fuel as the main fuel and diesel as the secondary or 
pilot fuel. Other modifications were done to enable data acquisition and computation, including 
EGR % flow piping arrangement, and thermocouples for temperature measurement for the 
exhaust system. 
The results report that the cylinder peak pressure reduced with the application of a dual fuel 
strategy, as the application of EGR % flow rate increased. The dual fuel exhibited longer 
ignition delay than conventional diesel fuel with the introduction of EGR % flow rate observed 
to greatly contribute to increases in ignition delay. The dual fuel strategy suffered significantly 
from low thermal efficiency compared to conventional diesel fuel at medium range engine 
loads, but high engine loads demonstrated high values of BTE. However, the application of the 
EGR % flow rate changed BTE of the dual fuel depending particularly on if the load applied 
increased, thus showing a general increase in BTE compared to conventional diesel at high 
engine loads. Therefore, it was concluded and reported that the dual fuel is a factor in reducing 
NOX emissions with application of EGR % flow rate in diesel engines.  
HC and CO emissions for the dual fuel combination returned more emissions than conventional 
diesel fuel especially during intermediate load, but with increased load the HC and CO 
emissions reduced. The application of EGR % flow rate didn’t show any significant reduction 
during the use of dual fuel diesel for HC and CO emission. This is identical to the study and 
conclusions observed by [54]. However, the values of CO2 were seen to be lower during the 
use of dual fuel diesel than during the use of conventional diesel fuel at all engine loads. The 
use of EGR % flow rate increased CO2 emissions even though the values continued to remain 
low compared to conventional diesel fuel. 
2.7 Effects of EGR on Temperature and Emissions 
[55] in “Effects of Diluent Admissions and Intake Air Temperature in Exhaust Gas 
Recirculation on the Emissions of an Indirect Injection Dual Fuel Engine Direct Injection” used 
a Ricardo E 6 modified to accommodate gaseous fuel and normal or conventional diesel use. 
The engine was single cylinder, four stroke, water cooled, with a maximum rated power of 9 
Kw at 3000 rpm. Modifications were done to the engine for the purpose of testing and study. 
Following testing and experimentation it was observed that the introduction of CO2, through 
the intake manifold resulted in a decrease in NOX emissions but showed a relative increase in 
43 
 
UHCs. This is a chemical effect that is associated with the use of the EGR application. The 
authors reported that by continuing to increase the inlet charge temperature there was an 
increase in UHC emissions, mainly due to a decrease in ignition delay in relation to the rise in 
the intake charge temperature. By increasing the intake charge temperature, the combustion 
characteristics were improved due to the decrease in the ignition delay. This improves and 
increases the BTE and brake horse power parameters, leading to a further decrease in the UHC 
and CO emissions.  
[56] in their study “Effects of EGR on Exhaust Gas Temperature and Exhaust Gas Opacity in 
Compression Ignition Engine” carried out an experiment study on a two cylinder, DI engine, 
with a 9.3 kW rated power at 1500 rpm. The engine was air cooled, running on diesel fuel.  
The authors found that brake specific fuel consumption was fairly independent of the EGR % 
flow rate. In other words, increased EGR % didn’t show a considerable change of BSFC. Other 
observations made included that opacity of the exhaust gas increased with the increase in the 
EGR % flow rate and as such the EGT reduced the combustion temperature with the application 
of EGR. Since NOX is a temperature dependent component, reduced combustion chamber 
temperatures leads to NOX reduction. 
[57] in “Effects of EGR on Diesel Premixed-Burn Equivalence Ratio” investigated the effects 
of the ER on premixed burn mixture in diesel engine combustion in relation to the application 
of various rates of EGR % flow rate. PM emissions reduction is related to an increase in ignition 
delay particularly if the delay exceeds injection duration and good fuel mixing, and if cooled 
EGR is used. This allows more mixing time for the air-fuel before ignition decreasing the ER 
of the igniting fuel, thus preventing soot formation where the 𝜙𝜙 <  2. 
[58] in their study on “Combined Effect of EGR and Inlet Air Pre-Heating on Engine 
Performance in Diesel Engines” used a four stroke diesel engine S195, single engine, with a 
maximum speed of 2000 rpm, DI, water cooled and naturally aspirated for their study. The fuel 
used for this study was conventional diesel with a density of 0.80 g/cc at 25 ° C. Other 
modifications to suit the experiment were done to provide measurement readings, particularly 
the preheating mechanism for inlet intake air – to maximize the heat loss the inlet intake air 
passage was insulated by plaster. 
The authors found that the emissions of NOX from the test engine increased as the engine speed 
increased in both the phases of the experiment. i.e. when the inlet intake air preheating was 
applied without EGR application and vice versa. They attributed this to the decrease in the 
44 
 
duration of ignition delay and combustion chamber premixing time of the inlet intake charge. 
In addition, they observed that the application of EGR reduced the in-cylinder combustion peak 
temperature due to the higher CO2 heat specific ratio. This result is identical to other results in 
other studies, particularly those where the EGR % was limited to between 20 % to 25 % by 
[59]. In terms of the relationship between the effects of EGR on inlet intake air preheating and 
NOX emissions, the authors observed that on increasing the inlet intake air temperature the 
emissions of NOX and CO decreased. This factor was linked to the ignition delay and premixed 
combustion mixture, two factors whose time is shortened by the introduction of EGR % flow 
rate. There are other identical studies consistent with this study by [60].  
On the relationship between the effects of the adiabatic flame temperature on the emissions of 
NOX, particularly when conventional diesel fuel is applied, the authors observed that as the ER 
increased (hitting a maximum ratio of 0.9), the adiabatic flame temperature also increased. 
However, the ER begins to decrease as it hits this maximum value. The application of the inlet 
intake air charge preheating at a controlled temperature of 55 ° C and without application of 
EGR % flow rate, lowered CO emissions to almost a minimum, explained by the fact that inlet 
intake air preheating provides CO oxidation. However, when EGR flow rate of 25 % was 
applied the CO emissions increased significantly. This concurs with the findings of a study 
conducted by [54], where it was concluded that the application of EGR doesn’t have similar 
effects on the different species of emission characteristics. 
The BTE was found to be affected by the inlet intake air preheating in combination with the 
application of 25 % EGR flow rate, but at varying engine speeds and constant load the authors 
found that the BTE was higher for the preheated inlet intake air as compared to when preheating 
was not applied. This was seen to be due to higher amounts of heat energy that was being 
transferred from the exhaust gas to the inlet intake air. This allowed for better burning of the 
fuel in the combustion chamber, therefore minimizing the overall heat loss by the system. In 
addition, they observed that the BTE increased with the increase in engine speed both at low 
and intermediate speeds, but at higher speeds they noted a decrease in the BTE, a factor which 
they attributed to shorter mixing time of the air fuel which is normally associated with higher 
engine speeds. 
[54] in “The Influence of Exhaust Gas Recirculation on Combustion and Emissions of N-
Heptane and Natural Gas Fueled Homogeneous Charge Compression Ignition (HCCI)” 
conducted a study at the University of Alberta fuel and combustion emission research facility. 
45 
 
They used a Waukesha CFR single cylinder research engine, water cooled, coupled to a 
dynamometer. The speed was maintained at 800 rpm at wide throttle. A temperature controlled 
air preheating system was installed using a 2.4 kW heater motor, also included were two port 
fuel injectors controlled electronically via an alternative fuel system (AFS) sparrow module. 
The EGR rate was manually controlled, the fuels used for this experiment were n-heptane and 
natural gas either as blends or used individually. 
During their study the authors observed that the recycled gases from the exhaust contained tri-
atomic molecules of H2O and CO2. This species increases the SHC of the combustion mixture 
and reduces the overall specific heat ratio leading to the mixture acting like a thermal heat 
absorbing sink, thereby decreasing the after-compression temperature. The other observation 
they made with regard to EGR % flow rate application was the reduction in peak in-cylinder 
charge temperature due to delayed auto ignition of the combustion charge. Applying EGR % 
flow rate affected the peak in-cylinder pressure by reducing auto ignition. This is explained by 
the fact that the recycled exhaust gases in the form of EGR act as a thermal sink as mentioned 
earlier, thus reducing the charge temperature during combustion. The pressure rise during the 
combustion becomes reduced which becomes a means of avoiding engine knock. 
Heat transfer was noticed to decrease with the application of EGR, since the heat transfer 
coefficient is proportional to pressure and inversely proportional to temperature. Therefore, the 
heat transfer coefficient did not vary much with the application of EGR, since by increasing 
EGR % flow rate both the pressure and temperature decreased. The gross HRR was seen to be 
affected by the application of EGR flow rate because recycled exhaust gases substitute part of 
the fresh air intake charge, thus affecting the in-cylinder ER and the combustion process in 
general. However, considering that EGR rate application can cause delay in SOC this prolongs 
the burning and exposes the elements of combustion to more time for combustion, hence 
increasing the IMEP limit and operating range. 
The effects of EGR on combustion timing was demonstrated to increase the specific heat ratio 
of the combustion elements as the cylinder inlet charge temperature decreased, thus delaying 
the ignition. This occurrence introduces retardation of SOC and increases the combustion 
duration. Defined as the indicated work output per unit of engine swept volume, the IMEP was 
seen to be less affected by the application of the EGR % flow rate which is identical with other 
studies done in this area by [61]. Regarding ISFC, the authors found out that the retardation of 
SOC (ignition delay) reduced ISFC, whereas when the EGR flow rate was applied it produced 
46 
 
the opposite effect i.e. it increased the ISFC. This factor can be explained by the fact that 
introduction of EGR% flow rate removed the availability of oxygen thus causing incomplete 
combustion in the combustion process thereby decreasing efficiency. 
The formation of NOX was limited due to the low charge temperature and that of the air mass 
flow rate. This is explained from two factors; firstly, applying EGR % rate flow depresses peak 
in-cylinder temperature thus reducing NOX emission formation and, secondly, the application 
of EGR % flow rate depresses the availability of O2 concentration in the combustion chamber 
due to EGR flow rate dilution effect. However, the increase of EGR rate was noticed to increase 
HC and CO emissions, as the EGR flow rate continued to increase, which can be explained by 
the unavailability of O2 and the reduced inlet intake charge temperature which results into 
incomplete combustion due to poor mixing and oxidation of the combustion mixture.  
[35] in the “Effects of EGR on Performance and Emission Characteristics of a Diesel Piloted 
Biogas Engine” used a single cylinder engine four stroke engine, water cooled, naturally 
aspirated constant speed CI engine. The engine was coupled to a hydraulic dynamometer for 
load application. Modifications were made to the engine to enable it to run on dual fuel with 
conventional diesel as a pilot fuel and biogas as the main fuel.  
After conducting tests and the experiment they came out with the following findings on BSFC. 
Through the variation of EGR flow rates and engine load from 0 % to 100 %, the BSFC was 
found to decrease as the rate of EGR % flow was increased obtaining a maximum value when 
the EGR flow rate was at 20 %. The decrease in the BSFC with the application of the EGR % 
rate flow was due to an increase in the inlet intake charge temperature, thus increasing the rate 
of fuel consumption. Regarding BP, they reported that the application of the EGR didn’t 
produce any significant effect on the total output power of the engine (with an application rate 
of 30 % EGR rate flow), while the BTE increased with the increase in the engine load for all 
operating conditions. However, it should be mentioned that the BTE showed some insignificant 
increase before the EGR rate flow passed the 20 % mark, attributed to the reburning of the 
hydrocarbons entering the combustion chamber together with EGR and the increase of inlet 
intake charge temperature as a result of EGR % rate flow application, which causes an increase 
in the rate of combustion. This finding is identical to the findings of [62]. 
Regarding EGT, they concluded that the EGT increases with engine load for all the operating 
conditions, which they attributed to the increase in input energy at high engine loads, as fuel 
consumption increases to accommodate high load and speed. Another observation they made 
47 
 
during this study was that EGT temperature reduced with the increase of EGR % flow rate, 
which they attributed to the unavailability of O2 when EGR % flow rate is applied leading to 
lowered combustion temperatures. This is identical to other studies conducted by [63]. 
The CO emissions variations in relation to the increase in EGR % flow rate was noted to 
increase with an increase in EGR % flow rate but was seen to significantly reduce as the EGR 
% flow rate and engine load increased to 75 % and 100 %. This factor is explained by the lack 
of O2 during combustion due to the dilution effect of EGR % flow rate. On the other hand, HC 
emissions are noticed to increase in all the engine load conditions as the EGR % flow rate is 
increased, explained by the fact that the induced charge (intake air + EGR rate) have a higher 
CO2 content when compared to fresh charge of inlet intake air only. However, the increased 
load across the entire engine load shows a decrease in HC emissions particularly when the dual 
fuel is employed for operation. This study findings are identical to the findings of [64]. Finally, 
on carbon emissions and how they interacted with the EGR % flow rate, the authors noted an 
increase of CO emissions with an increase in engine load, while when an EGR flow rate of 10 
% to 30 % was applied, there was a minimal decrease associated with the substitution of fresh 
charge with the elements of the exhaust gas which decreased the CO2 emissions.  
Figure 2.19 shows the variation of CO2 with varying engine load operating under the influence 
of different EGR flow rates [35]. 
 
Figure 2.19: Carbon Dioxide Vs Load 
 
[65] in “Effect of biodiesel–butanol fuel blends on emissions and performance characteristics 
of a diesel engine” investigated the effects of a butanol-biodiesel blend on emissions and 
performance of a diesel engine, single cylinder, naturally aspirated, indirect injection and water 
48 
 
cooled, using standard diesel (D100) and neat biodiesel B100 at four loads. The blend ratios 
were 5 %, 10 % and 20 % as B95Bu5, B90Bu10, B80Bu20 respectively. 
After testing and evaluation they reported that, compared to diesel, butanol blends showed 
lower EGT and lower NOX emissions. However, CO and UHC showed increased tendencies 
especially for higher blend ratios as compared to lower blend ratios of 5 % and 10 %. The 
butanol blend of B80Bu20 produced the highest emissions values for CO and UHC but lower 
NOX emission values as compared to diesel fuel. 
Figure 2.20 shows the variation of EGT with different fuels of diesel and the blends of butanol 
under varying engine load [65]. 
 
Figure 2.20: EGT VS Load 
 
Regarding EGT, the authors reported diesel fuel to have the highest EGT temperatures of all 
the test fuels used due to its high CV and high CN. The blends of butanol reported lower values 
for EGT as they have a lower CN and lower energy density factors which accounts for the 
difference between the two fuels. The high oxygen content of the butanol blends decreases the 
overall energy content in the blended fuels hence the lower EGT temperature and combustion 
as illustrated in Figure 2.20. 
[63] in the “Effect of Exhaust Gas Recirculation (EGR) Temperature for Various EGR Rates 
on Heavy Duty DI Diesel Engine Performance and Emissions” conducted a study using a 
heavy-duty DI diesel engine, single cylinder engine, with high compression peak pressure. To 
improve the BSFC, advanced timing was employed. The EGR for test study was constantly 
cooled and maintained throughout the experiment. The test engine condition was maintained 
49 
 
at full load with variations of engine speed. The theoretical study was on a 3D multi zone model 
modified to suit the experimental set up under study, to include the effects of EGR % rate flow 
and temperature and for calculation analysis. The boost pressure and injection pressure were 
kept constant at 280 BAR and 140 °C for the EGR gas. 
The study found a reduction in cylinder pressure during combustion and expansion of the 
mixture in the combustion chamber. The SHC increased due to the presence of exhaust gas 
recirculation which reduces the availability of O2 thereby affecting the combustion process rate 
and the disassociation of CO2 and H2O, which further reduces the peak cylinder pressure values 
as EGR % flow rate increases. During part engine load different results were be obtained, as 
the availability of O2 was higher compared to when the engine was at full load. As the EGR 
inlet intake temperature increases with EGR % flow rate, it leads to a delay in SOC which leads 
to increased peak cylinder pressure as EGR % flow rate increases. These findings are identical 
to [35, 63, 66]. 
The BTE result indicated a negative effect of the EGR % flow rate on the BTE, especially at 
low engine speeds and full load; increase in EGR % flow rate showed a significant lineal 
decrease in BTE of 5.5 % at 15 % EGR% flow. The authors found that the theoretical and 
measured values of NOX and soot at engine full load and varying EGR % flow rate showed 
that as the rate of EGR flow increased NOX emissions reduced lineally but soot emissions 
increased exponentially [63], the effects becoming more apparent and critical at low engine 
speeds. 
2.8 Effects of EGR On Performance, Combustion and Emissions on Biodiesel Fuels 
In recent years there has been a strong interest in the alcohols which contain long carbon chains 
such as butanol. The butanol molecular structure contains 4 carbons and good physical and 
thermodynamic properties. The higher alcohols have a greater potential to overcome the 
problem of the low energy density associated with alcohols from the lower segment of the 
molecular structure, thus reducing soot and CO emissions compared to the NOX emissions 
emitted by the conventional diesel fuel. 
[67] in “Study of a Diesel Engine Performance with Exhaust Gas Recirculation (EGR) System 
Fueled with Palm Biodiesel” investigated the effects of EGR on the parameters of SFC, EGT, 
and the emissions of NOX, CO, CO2 and UHC in experimental and simulated study situations. 
The study was conducted using a four-cylinder diesel engine with modified EGR application, 
50 
 
naturally aspirated, air cooled, and with a maximum power of 64.9 kW at 4500 rpm using palm 
biodiesel and mineral diesel.  
Figure 2.21 shows the variation of EGT with different blends of diesel and palm biodiesel [67]. 
 
 
Figure 2.21: EGT VS diesel and palm biodiesel with and without EGR flow rate application 
 
The application of EGR % flow rate mode on the palm biodiesel and mineral diesel fuels during 
the study led to a noted decrease in the EGT by 2.8 % when palm biodiesel was used as 
compared to 1.6 % decrease when mineral diesel was used. These findings are similar and 
identical to [68] and lead to the conclusion that most biodiesel fuels produce lower EGT 
temperature values at varied engine speeds and constant engine loads as illustrated in Figure 
2.21. This conclusion is attributed to the fact that the exhaust gases mix with the inlet intake 
fresh charge which dilutes the methyl ester rich oxygen content in the combustion mixture, 
thereby improving combustion quality and completion. 
Figure 2.22 shows the variation of BSFC and different test fuels of palm biodiesel with 
application of EGR [67]. 
 
51 
 
 
Figure 2.22: BSFC vs different test fuels with and without application of EGR 
 
The effect of EGR on the performance and combustion characteristics of the biodiesel palm 
biodiesel and mineral diesel under normal modes, at full load condition, was that palm biodiesel 
produced higher values of BSFC, as illustrated in Figure 2.22. This increase in BSFC was 
attributed to the drop in the efficiency of the engine and the lower HRR values associated with 
the palm biodiesel, in addition to the unavailability of O2 in the inlet intake air mass flow with 
the application of EGR % flow rate, which mixes the inlet intake air with the EGR gases. In 
general, the authors concluded that the combustion of biodiesel in diesel engines contributes 
lower COUHC, PM and smoke emission while conversely emitting higher NOX. The study 
further shows that palm biodiesel increases the EGT by 5.6 % above the exhaust temperature 
values that are produced when mineral diesel is used. This is attributed to the presence of high 
oxygen content in the palm biodiesel test fuel leading to a higher combustion efficiency and 
higher in-cylinder temperature. This result is identical to the findings of [69].  
Although the results obtained showed that palm biodiesel produced 4.7 % higher NOX 
emissions compared to the mineral diesel with EGR % flow rate application, it should be noted 
that low NOX emission of 5.4 % was achieved when mineral diesel was used compared to 22 
% by volume when palm biodiesel fuel was used. This is explained by the fact that an increase 
in total heat capacity (specific heat capacity) of the exhaust gases and a reduction in in-cylinder 
peak temperature tends to reduce NOX emissions. These findings are identical to the findings 
of [38] as shown Figures 2.23 and 2.24. 
Figure 2.23 shows the variation of the rate of pressure release (ROPR) in relation to the engine 
crank angle (CA) position at varying cylinder pressures [38].  
52 
 
  
Figure 2.23: ROPR VS Crank Angle 
 
Figure 2.24 shows the variation of ER with full engine load under the influence of EGR, with 
diesel fuel and compressed natural gas (CNG) [38]. 
 
 
Figure 2.24: Equivalence ratio VS Engine Full Load (diesel fuel and compressed natural gas) 
 
[33] in the “Effects of EGR (Exhaust Gas Recirculation) on the Performance and Emission of 
a CI Compression Ignition Engine Fueled with a Blend” used a single cylinder engine, four 
stroke water, compression ignition engine, with the test fuel being methanol at a concentration 
of 99 %. The test set up consisted of a test engine, a hydraulic dynamometer, fuel flow meters 
and incorporated various pieces of measuring equipment. 
In their findings, the BTE increased as the load increased. Using 10 % methanol diesel blend 
they found increased the BTE by up to 12.85 %, as compared to 20 % blend of methanol which 
improved the BTE by 26.23 % at 10 % EGR flow rate. (BSFC was reduced in 10 % and 90 % 
methanol and diesel blend, while the 20 % and 80 % had an increase of 4.28 % only on the 
BSFC. On the effects of emissions of CO, they found that the 10 % and 90 % methanol diesel 
53 
 
blend had a 30 % reduction in the emissions of CO, while with the methanol diesel blend of 20 
% and 80 %, the reduction was 40 %. The HC in the exhaust system reduced by 13.04 % in the 
10 % and 90 % methanol diesel blend and by 17.39 % with the 20 % and 80 % methanol diesel 
blend. The NOX reduced by 3.33 % in the 20 % and 80 % methanol diesel blend, and by 5.55 
% with the 10 % and 90 % methanol diesel blend. The exhaust temperature gas reduction was 
observed to reduce by 4°C with 10 % EGR flow rate as compared to 6°C 20 % EGR flow rate 
was applied. 
[70] in the “Effect of EGR on Diesel Engine Performance and Exhaust Emission Running with 
Cotton Seed Biodiesel” used a single engine four stroke, water cooled engine with cotton seed 
biodiesel (CSBD) and PBD (petroleum based diesel) blends as test fuels, where 0 % PBD was 
CSBD10, 20 % PBD was CSBD20 and 30 % PBD was CSBD30. The EGR % flow rate was 
limited to between 0 % to 20 % graduated in steps of 5 %. The engine speed was kept constant 
at 1500 rpm. The set up was to examine the following parameters in relation to the effects of 
EGR % flow rate; BSFC, BTE, EGT and the emission characteristics of NOX, HC and CO.  
The results indicate that the BSFC was slightly independent at lower EGR rates, but increased 
as the rate of EGR flow rate increased to 15 %. This was explained by the fact that less oxygen 
produces a rich mixture formation in the combustion chamber, hence more and increased fuel 
consumption. The BTE remained equally unaffected when low EGR rates and full engine loads 
were applied. However, BTE decreased with increase in EGR flow rate above 15 % in the 
combustion chamber, due to the displacement and replacement of fresh air charge by the EGT 
gases. The EGT partly operated with cooled EGR % flow rate was observed to lower the EGT 
as compared to EGT at normal engine operating condition without partially cooled EGR. This 
means that EGT decreased with increased EGR % flow rate, informed by two factors, namely, 
unavailability of oxygen and the high specific heat ratio of the intake air mixture which 
included exhaust mixture.  
On emission characteristics the authors noticed a significant decrease in NOX for all the test 
fuels used due to reduction in O2 and decrease in flame temperature within the combustion 
chamber. The hydrocarbon emissions showed an increase with the application of EGR % flow 
rate for all the test fuels used, especially when the EGR flow rate was above 15 %. This is 
explained by the incomplete combustion occasioned by the unavailability of O2 resulting in 
poor oxidation of the combustion mixture causing incomplete combustion. The CO emissions 
54 
 
were noted to increase with increase in EGR % flow rate for all the fuels used during the 
experimental study, mainly due the aforementioned O2 phenomenon. 
[71] in “Performance Characteristics of a Diesel Engine Operated on Biodiesel with Exhaust 
Gas Recirculation” studied the effects of EGR on the performance of a diesel engine using 
biodiesel. The diesel engine was a naturally aspirated single cylinder 4 stroke, water cooled, 
DI engine with rated power of 3.7 kW at 1500 rpm. After evaluation and testing the authors 
concluded that the BSFC increased for the pure biodiesel fuel, but decreased for biodiesel 
blends B10 and B20 compared to conventional diesel, a factor they attributed to the high 
viscosity index of biodiesels. With the application of EGR % flow rate they noticed an increase 
in the emissions of HC and CO but with reduced values for the biodiesel and the pure biodiesel 
blended fuels. 
The NOX emissions were observed to decrease with application of increased rate of EGR % 
flow. The authors noticed that biodiesel fuel with a combination of EGR % flow rate produced 
NOX reduction without increase in the fuel penalty. The smoke opacity with application of 
EGR % flow rate was observed to increase soot and PM emissions. This is explained by the 
reduction of EGT by dilution which causes less oxidation thus lowering the combustion 
temperatures. In their final observation they noted that the EGR % flow rates of 10 % to 15 % 
were the best rates relative to the traditional trade-off between NOX, HC and the fuel penalty. 
[72] in their study of “Exhaust Gas Recirculation for Advanced Diesel Combustion Cycles” 
used a four-cylinder Ford common rail diesel engine coupled to an eddy current dynamometer. 
The engine was operated in two modes; single cylinder mode (3 cylinder to 1 cylinder) where 
3 cylinders were operated in a conventional mode and the number 1 cylinder was marked for 
research with independent control of the parameters of intake pressure, exhaust back pressure 
and fuel injection. The second mode involved all the four cylinders with the engine running 
using a turbo charger, air fuel and EGR with no major system alterations. 
The authors proposed two methods widely used presently to measure and quantify the amount 
of EGR % flow rate; the mass-based approach method and the gas concentration approach 
method as illustrated in Figure 2.25. The gas-based method uses the carbon dioxide (CO2) gas 
concentration in the intake side by volume to make determination of the EGR % flow rate 
fraction. Thus, for this method there are two equations that can be used to define the EGR % 
flow rate fraction in terms of the gases of CO2 and O2. The Equations 12 and 13 [72] showing 
the two variations as: 
55 
 
 𝑟𝑟 ≈ [𝐶𝐶𝑂𝑂2]𝐶𝐶𝐶𝐶𝑎𝑎− [𝐶𝐶𝑂𝑂2]𝑎𝑎𝑎𝑎𝑏𝑏 [𝐶𝐶𝑂𝑂2]𝐶𝐶𝐶𝐶𝑎𝑎𝐼𝐼𝐸𝐸𝐸𝐸 [ ] ≈        Equation 12 𝐶𝐶𝑂𝑂2 𝑟𝑟𝑒𝑒ℎ−[𝐶𝐶𝑂𝑂2]𝑎𝑎𝑎𝑎𝑏𝑏 [𝐶𝐶𝑂𝑂2]𝑟𝑟𝑒𝑒ℎ
𝑟𝑟 = [𝑂𝑂2]𝑎𝑎𝐶𝐶𝑟𝑟−[ 𝑂𝑂2]𝑟𝑟𝑒𝑒ℎ𝐼𝐼𝐸𝐸𝐸𝐸         Equation 13 [𝑂𝑂2] 𝑎𝑎𝐶𝐶𝑟𝑟 −[𝑂𝑂2]𝑟𝑟𝑒𝑒ℎ
Where 𝑟𝑟𝐼𝐼𝐸𝐸𝐸𝐸 is the EGR % flow rate, [𝐶𝐶𝐶𝐶2]𝑖𝑖𝑛𝑛𝑖𝑖 is the carbon dioxide gas in the intake 
manifold, [𝐶𝐶𝐶𝐶2]𝑒𝑒𝑥𝑥ℎ is the carbon dioxide gas in the exhaust manifold side, [𝐶𝐶𝐶𝐶2]𝑎𝑎𝑛𝑛𝑎𝑎 is the 
carbon dioxide gas in the ambient or atmosphere, [𝐶𝐶2]𝑎𝑎𝑖𝑖𝑟𝑟 is the oxygen amount in the intake 
charge, [ 𝐶𝐶2]𝑒𝑒𝑥𝑥ℎ is the oxygen in the exhaust manifold. 
Figure 2.25 illustrates the second method which is mass-fraction-based model in a 
mathematical diagram representation with symbols. 
 
Figure 2.25: EGR schematic for the mass fraction method of calculating the EGR % flow rate in a system 
[72] 
 
After testing the models and how they interacted with the EGR % flow rate application the 
following results and conclusions were reached. The impact of intake and in-cylinder charge 
composition was characterized by in-cylinder excess air ratio for the recycled O2 with the EGR 
as a function of the EGR displacement. The effects of EGR on intake pressure and engine load 
showed that dilution of the intake charge with application of EGR could correctly be predicted 
through the intake oxygen concentration as compared to the EGR fraction as it was shown to 
lack additional information to show the effectiveness of the EGR % flow rate application [73].  
During low engine loads and low EGR it was noticed that the effectiveness of EGR decreased 
as EGR gases displaced small amounts of fresh air in the intake system due to the fact that the 
EGR gases contained O2 and N2 elements and vice versa. In other words, increasing the intake 
reduced the intake charge dilution for the same EGR ratio, while increase in EGR % flow rate 
(intake dilution) produced a leaner AFR with increased premixed combustion, which caused 
56 
 
reduced soot emissions, and improved the engine combustion efficiency. It should be 
emphasized here that O2 in the EGR is both a function of the engine load and EGR ratio. This 
is identical to the study findings of [74].  
[68] in the “Effects of Exhaust Gas Recirculation (EGR) on the Performance of a Constant 
Speed Diesel Engine Fueled with Pentanol Blends” investigated the effects of EGR on diesel 
engine performance and emissions using four pentanol blends of 10%, 20 %, 30 % and 45 
%with EGR flow rates of 10%, 20 % and 30%. They set out to reduce high NOX emissions 
prevalent during high engine loads. Their results reported that increasing EGR % flow rates 
reduced NOX emissions by almost 41 % at intermediate engine loads. The full load results 
indicate a drop in the increase of the NOX emissions to 33.7 % as in Figures 27.  
Smoke emissions increased for all the blends of pentanol with application of EGR past the 20 
% flow rate point, leading to a conclusion that reduction of NOX and UHC emissions can be 
achieved using pentanol blends especially with moderated EGR % flow rate of 20 % to 30 %. 
This point had the lowest emission values for CO and UHC respectively. 
Figure 2.26 shows variations of NOX at full engine load with application of EGR % flow rate 
with different pentanol blends and diesel fuel [68]. 
 
 
Figure 2.26: NOX VS EGR % with different blends of pentanol and diesel 
 
Figure 2.27 shows the variation of BTE under the influence of EGR % flow rate with different 
blends of pentanol and diesel fuel [68]. 
 
57 
 
 
Figure 2.27: BTE VS EGR % flow rates 
 
On performance they reported that the BTE and BSFC decreased with increased EGR % flow 
rate. The blend PEN45 showed the most decrease in the BTE OF 7.2 % compared to diesel fuel 
as can be seen in Figure 2.27. 
[75] in “Characteristics of a DI-CI Engine Fueled with Preheated Corn Biodiesel” investigated 
the characteristics of a DI-CI that was made to run on preheated corn biodiesel fuel (PHCBD). 
The following engine parameters were measured and investigated: BSFC, IMEP, BTE, BSEC, 
combustion characteristics of cylinder pressure, HRR, EGT and emission characteristics of 
hydrocarbons, CO, NOX, and HC The study used a DI-CI Kirloskar engine, single cylinder 
naturally aspirated, water cooled and with a constant speed of 1500 rpm.  
There were three types of test fuels used during this experiment, petroleum diesel (PD), corn 
biodiesel (CBD), and the preheated corn biodiesel (PHCBD). The engine load was varied in 
four major steps of 25 %, 50 %, 75 % and 100 %. After the tests and experimentation, the 
authors came up with the following observations and conclusions: The BTE of PD, CBD, and 
PHCBD was found to increase as the power output increased for all the test fuels available 
during the experiment and for all engine load conditions recording 38.1 %, 33.8 % and 36.1 % 
respectively as illustrated in Figure 2.27.  
Figure 2.28 shows the variation of BTE with BP at different loads and with test fuels of PD, 
CBD and PHCBD [75]. 
58 
 
 
Figure 2.28: Figure 29 BTE VS Brake Power 
 
The BSFC for the PHCBD is observed to be lower, but when compared to the other test fuel 
PDs both CBD and PHCBD consumed more fuel to produce the same amount of power per 
unit mass of the BP as the two biodiesels have a lower CV. However, at power output of 2.79 
kW the BSFC of the PD test fuel was equal to the BSFC of the CBDPH test fuel. This was due 
to improved fuel spray characteristics experienced with preheating. On the other hand, the 
BSEC which is defined as the amount of fuel energy required to produce a unit of BP output 
for test fuel PHCBD had a lower BSEC than that of the CBD in all the engine loads. However, 
it should be observed that the BSEC of the test fuel PH CBD was equal to brake specific energy 
consumption of PD due to the high rate of evaporation and good fuel atomization as a result of 
pre-heating of the PHCBD test fuel. 
The IMEP which indicates the available pressure of combustion product in the cylinder was 
observed to increase with increase in power output for all the test fuels applied. The IMEP of 
PHCBD was higher than PD and CBD for all the power output, while the peak cylinder pressure 
for the PHCBD was seen to be higher than both PD and CBD with highest peak pressure 
recorded at 66.6 % when the power output was 2.79 kW. The EGT for the CBD and PHCBD 
test fuels was lower compared to the PD test fuel, although CBD and PHCBD test fuels 
recorded very close temperature variant values. The lower EGT of test fuel PHCBD and CBD 
can be linked to the fuels having lower heating values. 
During combustion with the piston nearing TDC and the crank angle cycles of 360 ⁰ CA to 440 
⁰ CA, the PD and CBD showed almost identical pressure curves, while the PHCBD test fuel 
showed a sloping and gentle pressure curve. Regarding HRR, the authors observed that it was 
negative at the SOC for all the test fuels, but it rapidly expanded at late combustion stages. The 
PHCBD was seen to have the lowest HRR at the SOC than the CBD and PD test fuels but 
continued to increase as the combustion progressed.  
59 
 
The rate of mass fraction burnt in relation to the crank angle position at 2.79 kW for all the test 
fuels indicated to initiate at the 351 ⁰ CA position and complete the combustion cycle at the 
398 ⁰ CA position. It should be mentioned here that even though there are different burning 
rates for all the test fuels, the duration of the combustion was the same for all the test fuels. 
PHCBD test fuel had the lowest burning rate but experienced rapid combustion after the 370 ⁰ 
CA mark as in Figure 30. 
Figure 2.29 shows the variation of HRR in relation to the crank angle (CA) with the engine 
load after top dead center (TDC) position, for various test fuels [75].  
 
  
Figure 2.29: HRR VS Crank Angle 
 
The exhaust emissions of HC, CO, and NOX showed that CO emissions increased with increase 
in the power output load for all the test fuels used, due to the increase in the fuel consumption 
and knock resistance with increase in engine load. The CBD test fuel showed less CO emissions 
compared to the PD test fuel and these emissions reduced further with the use of the PHCBD 
test fuel due to reduced viscosity and density in addition to increased rate of evaporation. 
The HC emissions of corn biodiesel showed that without preheating it had the lowest emissions 
of HC, compared to PD test fuel, due to the high content of O2 in most biodiesel fuels. However, 
as the engine load increased the emissions of HC increased. The NOX emissions for PHCBD, 
PD and CBD test fuels in relation to engine load showed an increase for CBD, but there was 
reduced emission for PHCBD test fuels compared to when there was no preheating. It is 
important to note here that PHCBD showed almost similar values for NOX as the PD test fuel 
values. 
60 
 
2.9 The Effects of EGR On Particle Formation and Oil Degradation 
Though EGR has been adopted as the most effective technique in emission control in diesel 
engines, it also has the disadvantage of in-cylinder soot formation due to the poor oxidation; it 
produces smoke particles with the diameter of most particles being between 0. 1 µm to 0.8 µm 
compared to the lubricant boundary layer thickness of 0.001 µm to 0.005 µm. The increase in 
soot deposition is solely blamed on the wear of engine parts, namely, cylinder liners, piston 
rings, valve train and the bearing inserts both main and con bearing [76]. This is influenced by 
the additive degradation of the lubricating oil, as the chemical composition is altered due to the 
accumulation of the soot in the oil. The main wear mechanisms associated with soot presence 
in the engine lubricating oil are abrasion adhesion and scuffing which are known to damage 
surfaces mechanically as reported by [77]. 
[77] in “The Role of Soot Particles in the Tribological Behavior of Engine Lubricating Oils,” 
studied the tribological behavior of soot particles in the lubricating oil. The study mainly 
focused on the two types of lubricants CDSAE15W-40 as test pilot oil and 150SN as the base 
oil to conduct their experiment. Using carbon black, they simulated soot particles normally 
contained in the oil, while the engine wear was measured by a four ball tribometer. After the 
tests and results they concluded that the anti-wear and anti-friction properties of CDSAE15W-
40 were better than the base oil 150SN, a factor they attributed to the presence of carbon black 
as a dispersant and additive. 
Due to the intensification of the formation of particles as a result of application of EGR % flow 
rate in an engine, control aspects during transient operation need to be addressed for the proper 
implementation of LPL EGR % flow rate. The resistance to wear and frictional properties of 
base oil 150SN could be improved if an additive or dispersant was added. For the CDSAE15W‐
40 formulated lubricant, the friction coefficient decreased to a contamination level of 4 wt. %. 
This phenomenon was ascribed to uniform dispersion of carbon black that occurred in the test 
oil used during the study and experiment. The principal tribological mechanism of the 
simulated engine soot was attributed to the absorption and agglomerate effects of the tested 
lubricants. The wear resistance and frictional properties of the 150SN base oil with 2.4 wt. % 
carbon black contents were improved by the addition of the T154 dispersant. 
Table 2.2 shows the comparison between the mean oil film thickness ( ℎ𝑎𝑎𝐶𝐶𝐶𝐶) and the mean 
carbon black diameter. 
61 
 
 
Table 2.2: Comparison between the mean oil film thickness ( 𝒉𝒉𝒎𝒎𝒎𝒎𝒎𝒎) and the mean carbon black diameter 
[77] 
Property Oil 150SN Oil CD SAE 15W-40 
Temperature ⁰C 40 100 40 100 
Kinematic viscosity (cSt) 32 4.80 110.60 15.02 
Mean oil film thickness (nm) 48.48 13.38 107.10 28.10 
Minimum oil film thickness (nm) 30.42 8.26 68.35 17.58 
Minimum soot diameter (nm) 40 40 40 40 
 
[78] in “Characteristics of Diesel Engine Soot that Lead to Excessive Oil Thickening” studied 
diesel engine soot characteristics which lead to oil thickening. Based on examination and 
analysis of the morphology of used oil from HDD operating with different levels of EGR, the 
authors concluded that severe oil thickening is directly linked to soot agglomeration seen in 
HDD oils due to the ineffectiveness of the oil dispersant to retard the growth of soot particles 
in engine lubrication oil. 
In “Impact of Exhaust Gas Recirculation (EGR) on Soot Reactivity from a Diesel Engine 
Operating at High Load”, [79] studied soot reactivity from a diesel engine operating at high 
engine load of over 75 % (336 Nm) under the influence of 0 %, 10 %, and 30 % EGR flow 
rates. Following testing and analysis they reported increased combustion duration of 39.8 % as 
the EGR flow rate increased from 0 % to 30%. The authors continued to report that there was 
a corresponding change in the oxidative reactivity, soot particle nanostructure and 
carbonaceous components, as the EGR flow rate increased, as can be seen in Figure 2.30. This 
they attributed to deterioration of engine performance as the EGR % flow rate was applied at 
high load with oxygen concentration decreased thus lowering the in-cylinder temperature. 
Finally, the authors reported soot with high soot-elemental carbon (EC) content, high ordered 
graphitic structure and low carbon fraction in relation to the soot particle samples analyzed 
from high EGR % flow rates. They attributed this to longer combustion duration and a decrease 
in the AFR. 
Figure 2.30 shows the variation of specific emission and mass fraction with exhaust gas 
recirculation. 
62 
 
 
Figure 2.30: Effects of EGR on (a)specific emissions and (b) mass fractions of carbonaceous components 
[79] 
 
2.10 Effects of Emission Gases on Public Health  
The modern day mass transport systems which include road, rail, water and air produce a lot 
of air pollution both in the developed countries and the developing nations, with major health 
implications. This is more apparent in highly congested major metropolitan urban centers and 
cities. The pollutants of health concern can be classified into the several source areas of 
production such as TRAP (traffic related air pollution), particulate matter (PM), VOCs, and 
EC [80]. 
In major urban cities and centers tail pipe emission account for approximately 30 % of the total 
PM emissions production by mass. This is in addition to the PM emissions emanating from 
suspended road dust, wear from tires, and brake pad dust [80]. However, it should be mentioned 
that the side effects of TRAP are diverse and numerous despite the human body’s ability to 
cleanse itself, with most of the inhaled TRAP PM never being able to move out of the 
respiratory system completely. The most vulnerable elements of the population are the elderly, 
infants, pregnant women and those that have already been struggling with chronic respiratory 
system diseases. Table 2.3 is a summary of the various sources of and types of pollution and 
their effects on human health and the environment.  
 
 
 
 
63 
 
Table 2.3: The main sources of pollution, types of pollutants, the environmental and health effects of these 
pollutants [80] 
MAIN SOURCE TYPE OF ENVIROMENTAL HEALTH 
POLLUTANT EFFECTS EFFECTS 
    
Industry Sulphur dioxide (SO2) Causes acid rain, which Direct cause of 
pollutes all water sources, cardiovascular 
and corrodes housing complications 
settlements. 
Automobiles and Nitrogen oxide (NOX) Over fertilization and Health 
industry eutrophication due to complications 
excess nitrogen arising from cardio-
vascular diseases 
Industry and Particulate matter Smog leading to poor Penetration of lungs 
Auto-mobiles (PM) visibility and bloodstream 
Automobiles Carbon-monoxide Increases greenhouse Headaches and 
(CO) gases and global warming Fatigue 
Automobile and Volatile organic Smog causing poor air Skin and eye 
industrial compounds (VOCs quality irritation, 
processes carcinogenic, 
application nausea and 
headaches 
Use of fossil Lead (Pb) Leading cause of death in Damages to the 
fuels in transport animal and marine life human nervous 
application system if it 
processes accumulates in the 
blood-stream  
Chemical Ozone Smog, increased food Leading cause of 
reactions of NOX insecurity due to poor respiratory diseases 
and VOCs harvests and affects the 
forest cover due to 
prolonged drought 
 
64 
 
[81] in “Human Health Risk Assessment of Air Emissions from Development of 
Unconventional Natural Gas Resources” report that the process of mining natural gases results 
in direct and fugitive air emissions that are a complex mixture of pollutants from natural gas 
resources as well as from diesel engines tanks which produce water due to condensation.  
Multiple studies that have been conducted reveal that inhalation of toxic gases from petroleum 
hydrocarbons like conventional diesel used in heavy duty and LD engines indicates an 
increased risk of cardiovascular diseases [82], eye irritation, headaches, asthma symptoms, 
acute childhood leukemia, acute myelogenous leukemia, and multiple myeloma [81]. It should 
be mentioned here that toxicity information on some of the toxic gases emitted by diesel 
engines is still limited according to the studies that have been done so far. Although tremendous 
progress has been made in reducing air pollution and thus improvement in public health, there 
is still a need for isolation of the individual pollutants species and source categorization to 
increase specific understanding of the effects produced on human health, making it possible to 
adapt control actions to be developed for different sources and better prioritization of mitigating 
measures.  
The problem of air pollution is becoming a serious global concern and as such more emphasis 
has to be put on the role of the transport sector in impacting the health of the general public, 
considering the negative environmental, social and economic impact arising from the existing 
transport sector planning and execution of policies. There seems to be a bias towards the short-
term public health indicators like accidents and noise exposures, while ignoring the long-term 
effects like pollution from diesel car emissions because their impact on public health is easily 
noticeable or complained about [83, 84]. 
Additional work proposed by Bhalla et al. (2015) and Lelieveld et al. (2015) is the need to 
integrate the emerging understanding of health effects into improved model simulations and 
assessments of human health outcomes allowing better emission mitigation and quantification 
of the health benefits and challenges. Therefore, collaboration between transport and urban 
planning is needed [85] in order to make improvements, structurally and globally between 
modern transport and human health issues. The recent scandal involving Volkswagen motors 
in relation to emission cheating in their diesel operated vehicles sold in the world market has 
brought to the attention of the public the impact of pollution on public health from the private 
and the commercial transport sectors and how commercial interests can override public health 
issues.  
65 
 
Though diesel engines have low carbon emissions and improved fuel efficiency, it has been 
proved that they are a primary cause of major health concerns due to their emission of NOX 
and PM compared to other fuels in use today.  
In “What We Breathe Impacts Our Health: Improving Understanding of the Link Between Air 
Pollution and Health”, [86] state that air pollution contributes to the premature death of millions 
of people annually with increasing air quality problems in developed and developing nations. 
The call for closer cooperation between air pollution scientists and engineers especially 
concerning understanding of the chemical and physical properties of air pollution mixtures and 
their long-term implications. The past 15 years has seen the growth of ambient air pollution 
increasing exponentially, and so the array of acute and chronic health effects associated with 
it. The global burden of disease (GBD) study report ranked exposure to ambient fine PM 
(PM2.5) as the seventh largest factor contributing to global mortality of 2.9 million to 3.3 million 
people in 2013 alone [87], of which 184 000 deaths annually are directly linked to the impact 
of air pollution alone [84]. Using more sophisticated source modelling, [83] estimated that 
about 1/5 of all the premature deaths in UK, USA and Germany occur due to ambient PM2.5 
and ozone (O3) emissions. This accounts for 5 % globally for the figure of 2.9 million to 3.3 
million people whose death is associated with outdoor air pollution. Therefore, there needs to 
be a serious reduction in global emissions from the transport sector, in order to prevent 
increases in health effects and the GBD arising from atmospheric air pollution. 
[86] propose a multi-faceted and multi-disciplinary approach between the various fields of 
health sciences, toxicology, epidemiology, exposure science, clinical medicine and engineering 
and the physical science fields of atmospheric science, physics and chemistry. However, there 
has been a lot of progress made in the categorization of air pollution in recent years due to the 
widespread use of the satellite imaging technology. Among the advances are the aerosol mass 
spectrometer and the proton transfer reaction mass spectrometer [88], as well as the 
introduction of novel methods of analysis method in the study of aerosol components with 
adverse health effects utilizing mobile and wearable monitoring devices, thus making 
monitoring of human exposure in new locations possible and easy. This facilitates simulation 
modelling of the transportation modes and mechanical formation of the main pollutants locally, 
regionally and globally. 
  
66 
 
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72 
 
CHAPTER THREE 
METHODOLOGY 
3.1 Introduction 
This work used two methodological approaches in answering the main research aim and 
objectives of this study work. The first approach was numerical with the development of 
equations in modelling that could objectively help in providing the answers to the aim and 
objectives. The second approach was experimental work, which enable the harmonization of 
the experimental work and the numerical theoretical equation work. The two approaches and 
their findings will be presented and discussed in the Chapters 8 and 9 of this work. 
This work utilized numerical equations to help in the calculation and data collection processes 
of this work especially in relation to the parameters of performance, namely: BTE, BSFC, BP, 
and EGT, and emission characteristics of NOX, CO, CO2, UHC and smoke opacity which are 
all closely interlinked with one another. 
3.2 Numerical Equations Model 
The main control equations used in the numerical equation modelling in this work were taken 
from [1-5]. The first equations deal mostly with the first law of thermodynamics and its wider 
applications in the equations of rate of change of mass in an open system across the system 
boundaries. The first law of thermodynamics states that energy can neither be created or 
destroyed but can be adapted into other forms of energy [6]. There are three types of systems 
applied in the thermodynamics of internal combustion engines, namely, the open system, the 
closed system, and the isolated system. For the purpose of this work the open system was 
utilized and defined according to [7], that it is a system whose boundaries allow for mass 
transfer, heat transfer and work. 
3.3 Definitions of Numerical Equations  
3.3.1 First Law and Conservation of Mass 
The first equation relates to the conservation of mass clearly following the definition of the 
first law as aforementioned, and expressed mathematically as: 
𝒅𝒅𝒎𝒎 = ∑𝒋𝒋ṁ𝒋𝒋          Equation 1 𝒅𝒅𝒅𝒅
73 
 
Where 𝑑𝑑𝑛𝑛/𝑑𝑑𝑖𝑖 is the rate of change of the mass within an open system, ∑𝑟𝑟ṁ𝑟𝑟  is the sum or net 
flux of mass flow rate of the 𝑗𝑗𝑎𝑎ℎ species. 
3.3.2 The Second Law and Conservation of Species 
The second equation tracks and looks at the combustion products and the processes within the 
combustion chamber based on the conserved mass balance of the individual species within the 
combustion process and after combustion, referred to as the conservation of species and 
expressed as: 
Ẏ = ∑ �ṁ𝑗𝑗�  �Ẏ𝑟𝑟 − Ẏ𝐶𝐶𝑐𝑐𝑐𝑐𝑟𝑟 𝑟𝑟 𝐶𝐶 𝐶𝐶 � +  
𝛺𝛺𝑖𝑖𝑊𝑊𝑎𝑎𝑚𝑚       Equation 2 
𝑎𝑎 𝜌𝜌
Where 𝛺𝛺𝐶𝐶 is the dimensionless integral of order unity dependent on the force interaction during 
collision of the 𝑖𝑖𝑎𝑎ℎ  species, Ẏ𝑟𝑟𝐶𝐶  is the stoichiometric coefficient on the product side, 𝜌𝜌 is the 
density of air, ṁ𝑟𝑟 is the mass flow rate of the 𝑗𝑗𝑎𝑎ℎ species, Ẏ
𝐶𝐶𝑐𝑐𝑐𝑐
𝐶𝐶  is the stoichiometric coefficient 
on the reactant side, 𝑊𝑊𝑎𝑎𝑚𝑚 is the system flow representation, 𝑛𝑛 is the total mass within the 
control cylinder. 
3.3.3 The Conservation of Energy 
The third equation concerns the conservation of energy in an open system, in a thermodynamic 
internal combustion engine, where the mass transferred in and out of the system can be given 
as a statement in the form of Equations 3 and 4 as: 
ṁ𝐶𝐶𝐶𝐶 = ṁ𝑜𝑜𝑢𝑢𝑎𝑎 = ∆ṁ𝑠𝑠𝑐𝑐𝑠𝑠𝑎𝑎𝑟𝑟𝑎𝑎        Equation 3 
𝑑𝑑 (𝑎𝑎𝑢𝑢)  =  −𝑝𝑝 𝑑𝑑𝑑𝑑  + 𝑑𝑑𝑄𝑄ℎ𝑎𝑎 + ∑ ṁ
𝑑𝑑𝑎𝑎 𝑑𝑑𝑎𝑎 𝑑𝑑𝑎𝑎 𝑟𝑟 𝑟𝑟
ℎ𝑟𝑟        Equation 4 
Where 𝑑𝑑 (𝑎𝑎𝑢𝑢) is the internal energy, −𝑝𝑝 𝑑𝑑𝑑𝑑 is the work done by piston displacement, 𝑑𝑑𝑄𝑄ℎ𝑎𝑎 is the 
𝑑𝑑𝑎𝑎 𝑑𝑑𝑎𝑎 𝑑𝑑𝑎𝑎
heat transfer within the system, ∑𝑟𝑟ṁ𝑟𝑟ℎ𝑟𝑟 is the enthalpy flux of the system, ṁ𝐶𝐶𝐶𝐶 is the mass flow 
in the inlet intake, ṁ𝑜𝑜𝑢𝑢𝑎𝑎 is the mass of exhaust outlet gases, ∆ṁ𝑠𝑠𝑐𝑐𝑠𝑠𝑎𝑎𝑟𝑟𝑎𝑎 is the change in the mass 
flow rate in the system. 
In the following equations some assumptions have been made concerning the opening and the 
closing times for the exhaust and inlet valves in order for the operational equations that are 
used during the simulation of this work to be simplified, for the cylinder equations concerning 
conservation of energy in the cylinder volume. The assumptions made are that the pressure 
remains constant and uniform during compression stroke and remain so as the charge in the 
74 
 
cylinder moves to the expansion stage. Secondly that the gases in this open thermodynamic 
system will be and continue to be ideal gases. The third assumption is that during propagation 
of the combustion flame in the combustion cylinder the burned and unburned gas regions are 
separated and there is no heat exchange between the burned and the unburned areas referred to 
as combustion zones.  
The following parameters of engine performance were calculated based on the above 
assumptions: BSFC, BTE, BP, EGT and the emission parameters of NOX, CO2, CO, UHC and 
smoke emissions, which were analyzed through measurements by a five channel gas analyzer 
machine and the data collected relayed to a data logging interface machine.  
3.3.4 Air/Fuel Ratio 
The of air/fuel ratio (with symbol λ) is given in the following equation and defined as: 
𝜆𝜆 = (𝐼𝐼/𝑀𝑀) = (𝑎𝑎𝑎𝑎/𝑎𝑎𝑓𝑓)         Equation 5 
(𝐼𝐼/𝑀𝑀)𝑆𝑆 (𝑎𝑎𝑎𝑎/(𝑎𝑎𝑓𝑓)𝑆𝑆
Where 𝜆𝜆 is the lambda symbol for the air/fuel ratio, 𝑅𝑅⁄𝐹𝐹 is the air fuel ratio, (𝑅𝑅⁄𝐹𝐹)𝐸𝐸 is air fuel 
ratio under the stoichiometric conditions, 𝑛𝑛𝑎𝑎 is the inlet air intake mass flow rate, 𝑛𝑛𝑓𝑓 is the 
inlet intake fuel flow rate. 
3.3.5 Brake Thermal Efficiency (BTE) 
This can be defined as the BP of a heated engine as a function of the thermal input from the 
fuel used to propel that engine, by converting the heat from fuel into useful mechanical energy. 
This is achieved by comparing the output of the engine system to the input of the engine system 
i.e. BP as the output dividing it against fuel power as the input as in Equation 6 to obtain BTE 
as: 
 𝜂𝜂 =  Ẇ𝑏𝑏𝑝𝑝𝑏𝑏𝑎𝑎𝑟𝑟          Equation 6 ṁ𝑓𝑓×𝐶𝐶𝐶𝐶
Where 𝜂𝜂𝑏𝑏𝑎𝑎𝑟𝑟 is the BTE, Ẇ𝑏𝑏𝑝𝑝 is the engine BP, ṁ𝑓𝑓 is the conventional diesel fuel flow rate to 
the engine, 𝐶𝐶𝐶𝐶 is the calorific value of the test fuel.  
However, for dual fuel operating mode, the following equation is used to calculate the BTE as 
follows: 
𝜂𝜂𝑏𝑏𝑎𝑎𝑟𝑟 =  
Ẇ𝑏𝑏𝑝𝑝         Equation 7 
ṁ𝑓𝑓𝑓𝑓+𝐶𝐶𝐶𝐶𝑓𝑓𝑓𝑓+ṁ𝑓𝑓𝑝𝑝𝑓𝑓+𝐶𝐶𝐶𝐶𝑓𝑓𝑝𝑝𝑓𝑓
75 
 
Where 𝜂𝜂𝑏𝑏𝑎𝑎𝑟𝑟 is the BTE, Ẇ𝑏𝑏𝑝𝑝 is the engine BP, ṁ𝑓𝑓𝑑𝑑 is the rate of conventional diesel fuel flow 
into the engine, 𝐶𝐶𝐶𝐶𝑑𝑑𝑓𝑓 is the calorific value of the conventional diesel fuel, ṁ𝑓𝑓𝑝𝑝𝑜𝑜 is the WPPO 
fuel rate flow into the engine, 𝐶𝐶𝐶𝐶𝑓𝑓𝑝𝑝𝑜𝑜 is the calorific value of WPPO fuel. 
In order to establish the amount of diesel that is substituted with the use of WPPO fuel blend, 
a diesel substitution calculation is commuted and attempted as in Equation 8 as: 
𝐷𝐷𝐷𝐷 = 𝐷𝐷𝑓𝑓𝑓𝑓−𝐷𝐷𝑓𝑓𝑎𝑎𝑓𝑓 × 100        Equation 8 
𝐷𝐷𝑓𝑓
Where DS is the diesel substitution in percentage form, 𝐷𝐷𝑑𝑑𝑓𝑓 is the conventional diesel fuel 
consumption by the experimental engine in single fuel mode in kg/h, 𝐷𝐷𝑑𝑑𝑎𝑎𝑓𝑓 is the diesel fuel 
consumption in dual fuel mode. 
3.3.6 Brake Power (BP) 
The BP is given by the following equation considering that the probability of the air fuel 
mixture in the combustion chamber was fully burnt as it is assumed to contain the necessary 
amount of air to allow to it burn completely the fuel available and to produce the BP, is 
calculated as a product of the engine torque (𝑅𝑅) and the angular rotational speed (ɷ) of the 
engine crankshaft in Equation 9 as: 
𝑃𝑃𝑏𝑏  =  𝑅𝑅 ∗ ɷ          Equation 9 
Where 𝑃𝑃𝑏𝑏 is the BP theoretically, 𝑅𝑅 is the engine torque, ɷ is the angular engine rotational 
speed as measured from the crankshaft. 
3.3.7 Brake Specific Fuel Consumption (BSFC) 
BSFC is a measure of the amount of fuel that an engine often uses or utilizes to produce a unit 
of power measured in kilowatt hour (kW), and is defined in Equation 10 as: 
𝐵𝐵𝐷𝐷𝐹𝐹𝐶𝐶 = ṁ𝑓𝑓          Equation 10 
𝑃𝑃𝑏𝑏
Where BSFC is the brake specific fuel consumption, ṁ𝑓𝑓is the intake fuel flow rate, 𝑃𝑃𝑏𝑏is the 
BP. 
3.3.8 The Volumetric Efficiency (VE) 
The volumetric efficiency (VE) with symbol (𝜂𝜂), is a measure of how efficiently an engine can 
move the products of the air fuel mixture in and outside the engine cylinder, in other words, 
76 
 
it’s the actual ratio of the quantity or amount of fuel and air that enter the engine cylinder during 
the intake stroke in relation to the actual holding capacity of engine cylinder and as such the 
volumetric efficiency is not a volume ratio but a mass ratio [8], especially in static engine 
conditions and defined in Equation 11 as:  
𝜂𝜂 = 𝐶𝐶𝑎𝑎𝑎𝑎𝐶𝐶           Equation 11  𝜌𝜌𝑎𝑎 𝑉𝑉𝐶𝐶 𝑁𝑁 
Where 𝜂𝜂𝐶𝐶 is the cylinder volumetric efficiency, 𝑁𝑁 is the engine speed, 𝐶𝐶𝐶𝐶 is the displaced 
volume by the piston depending on the number of cylinders, 𝜌𝜌𝑎𝑎 is the air density, 𝑛𝑛𝑎𝑎is the inlet 
intake air flow rate, 𝑛𝑛 is the number of revolutions of the crankshaft per cycle. 
3.4 The EGR Valve Model Equations 
Though EGR as a technology reuses the engine exhaust system gas products in the combustion 
process to reduce the amount of O2 in the inlet intake manifold, to suppress NOX, there is a 
limitation on the amount of exhaust gases that can be recirculated without compromising AFR 
levels and thus depriving the engine of the necessary O2 causing damage and failure. 
There are two types of EGR quantification systems that are in use today and which have been 
adopted by various studies to conduct EGR flow rates for various works as presented in the 
literature reviewed in Chapter 2. The first utilizes the gases in the combustion product and 
hence the name of gas method. The second one utilizes the mass of the combustion products 
hence the name mass method and shares the same definition as the molar method. There are 
two assumptions made in order for molar method to be the same as the mass method of 
quantifying the EGR flow rate. These assumptions are: 𝑖𝑖) 𝑀𝑀𝑐𝑐  =  𝑀𝑀𝑓𝑓 and 𝑖𝑖𝑖𝑖) 𝑛𝑛𝑓𝑓  <  𝑦𝑦, where 
the 𝑀𝑀𝑐𝑐 is the molecular mass of the inlet intake charge, 𝑀𝑀𝑓𝑓 is the molecular mass of the fuel, 
𝑛𝑛𝑓𝑓 is the fuel molar quantity, and 𝑦𝑦 is the inlet intake molar gas quantity. 
During the model development of the EGR only the gaseous components of the combustion 
mixtures are taken into consideration [2] for purposes of simplification of the combustion 
equations. This is despite the fact that the EGR applications include all the components that are 
found in the exhaust manifold system. The combustion components considered are N2, O2, 
CO2, and H2O, while from the exhaust the components are N2, O2, CO2 and H2O vapor, while 
the inlet intake side contains N2 and O2. The equations for the EGR valve model are taken from 
[2, 9] as in Figure 3.1. 
77 
 
  
Figure 3.1: The EGR schematic loop nomenclature 
 
Figure 3.1 illustrates the EGR schematic loop definitions and nomenclature as follows: 1. is 
the direction of EGR gases, 2. is the 𝑥𝑥 subscript representing the exhaust molar gas quantity, 
3. is the direction of inlet gases fresh charge, 4. is the 𝑧𝑧 subscript representing remainder of the 
intake charge, 5. is 𝑛𝑛𝑓𝑓 which is the fuel molar quantity, 6. is the EGR valve, 7. is the Rx molar 
gas ratio, 8. is the subscript 𝑦𝑦 representing the inlet intake molar gas quantity, 9. is the direction 
of the exhaust gases exit, 10. is the engine unit.  
𝑦𝑦 = 𝑛𝑛𝑂𝑂2 +  𝑛𝑛𝐼𝐼2 +  𝑛𝑛𝐶𝐶𝑂𝑂2  + 𝑛𝑛𝐸𝐸2𝑂𝑂       Equation 12 
𝑥𝑥 = 𝑒𝑒𝑂𝑂2 +  𝑒𝑒𝐼𝐼2 +  𝑒𝑒𝐶𝐶𝑂𝑂2 +  𝑒𝑒𝐸𝐸2𝑂𝑂       Equation 13 
Where 𝑛𝑛  and 𝑒𝑒  are the component species of the 𝑖𝑖𝑎𝑎ℎ𝐶𝐶 𝐶𝐶  of the number of moles of the molecular 
components of (O2, N2, CO2, and H2O) of the test engine combustion products respectively, 
leading to the conservation of mass equation of the combustion process as: 
𝑦𝑦𝑀𝑀𝑐𝑐 + 𝑛𝑛𝑓𝑓𝑀𝑀𝑓𝑓 = 𝑥𝑥𝑀𝑀𝑐𝑐         Equation 14 
Where 𝑦𝑦 is the inlet intake molar gas quantity, 𝑥𝑥 is the exhaust molar gas quantity, 𝑀𝑀𝑐𝑐 is the 
molecular mass of the inlet intake charge, 𝑀𝑀𝑓𝑓 is the molecular mass of the cycle fuel, 𝑛𝑛𝑓𝑓 is the 
cycle fuel molar quantity. 
Since the molecular masses are assumed to be equal for the inlet and the combustion process 
products, for the chemical balance of the atomic components of O, N, C, and H the equations 
for the exhaust molar quantities take the form:  
𝑒𝑒 = 𝑛𝑛 + (𝛾𝛾 − 𝛼𝛼 − 𝛽𝛽𝑂𝑂2 𝑂𝑂2 ) 𝑛𝑛𝑓𝑓        Equation 15 2 4
78 
 
𝑒𝑒𝐼𝐼2 = 𝑒𝑒𝐼𝐼2          Equation 16 
 𝑒𝑒𝐶𝐶𝑂𝑂2 = 𝑛𝑛𝐶𝐶𝑂𝑂2 +  𝛼𝛼𝑛𝑛𝑓𝑓         Equation 17 
𝑒𝑒𝐸𝐸2𝑂𝑂 = 𝑛𝑛
𝛾𝛾
𝐸𝐸2𝑂𝑂 +  𝑛𝑛𝑓𝑓         Equation 18 2
Where 𝑒𝑒𝑂𝑂2 is the exhaust engine oxygen molar quantity, 𝑒𝑒𝐼𝐼2 is the exhaust molar quantity of 
nitrogen, 𝑒𝑒𝐶𝐶𝑂𝑂2 is the exhaust molar quantity of carbon dioxide, 𝑒𝑒𝐸𝐸2𝑂𝑂 is the exhaust molar 
quantity of water vapor, 𝑛𝑛𝑓𝑓 is the cycle fuel molar quantity, 𝛾𝛾 is the fresh air fraction mass, 𝛼𝛼 
is the thermal diffusivity of the cooling medium of the exhaust gases in the exhaust manifold, 
𝑛𝑛 is the number of moles from fresh inlet intake charge. 
Manipulating and combining Equations 15, 16, 17 and 18, the exhaust molar gas quantity can 
be expressed as:  
𝑥𝑥 = 𝑦𝑦 + (𝛽𝛽 + 𝛾𝛾)𝑛𝑛𝑓𝑓         Equation 19 4 2
Where 𝑥𝑥 is the subscript for exhaust molar quantity, 𝑦𝑦 is the inlet intake molar gas quantity, 𝛽𝛽 
is the reciprocal of the temperature mean of the EGR gases, 𝑛𝑛𝑓𝑓 is the cycle fuel molar quantity, 
𝛾𝛾 is the fresh air fraction mass. 
Using the definition of the molar EGR ratio which is given by the Equation 20 as: 
𝐸𝐸 =  𝑟𝑟           Equation 20 
𝑒𝑒
Where 𝐸𝐸 is the EGR molar ratio, 𝑥𝑥 is the exhaust molar gas quantity, 𝑟𝑟 is the total moles of the 
exhaust gas recirculated and considering the steady state produced in each engine cycle, in 
order to express the in cylinder charge in terms of the volumetric combination of the air fresh 
and EGR and the remainder of the inlet intake charge denoted as subscript 𝑧𝑧, the equation thus 
becomes as follows: 
𝑦𝑦 = 𝑧𝑧 + 𝐸𝐸𝑒𝑒          Equation 21 
Where 𝑦𝑦 is the inlet intake molar gas quantity, 𝑧𝑧 is the remainder of the inlet intake air charge, 
𝐸𝐸𝑒𝑒 is the total EGR molar ratio quantity of the moles of exhaust gases recirculated per engine 
cycle done. 
The value of R as the universal gas constant can be calculated as follows by manipulating and 
combining Equations 19 and 21 and obtaining:  
79 
 
𝐸𝐸 = 𝑐𝑐−𝑧𝑧𝛾𝛾 𝛽𝛽          Equation 22 𝑐𝑐+�2+ 4�𝐶𝐶𝑓𝑓
Where 𝐸𝐸 is the universal gas constant, 𝑦𝑦 is the inlet intake molar gas quantity, 𝑧𝑧 is the remainder 
of the inlet intake air charge, 𝑛𝑛𝑓𝑓 is the cycle fuel molar quantity, 𝛾𝛾 is the fresh air fraction 
mass, 𝛽𝛽 is the reciprocal of the temperature of the EGR gases mean. 
The 𝑦𝑦 component as a subscript for the inlet intake molar gas quantity can also be derived from 
the engine speed-density relationship based on the conditions of the intake manifold and the in 
cylinder volumetric efficiency in Equation 23 as:  
𝑦𝑦 =  𝜂𝜂 𝑝𝑝𝑖𝑖𝑎𝑎𝑉𝑉𝑠𝑠𝐶𝐶           Equation 23 Ṝ𝐼𝐼𝑖𝑖𝑎𝑎
Where 𝑝𝑝𝐶𝐶𝑎𝑎 is intake manifold pressure, 𝑅𝑅𝐶𝐶𝑎𝑎 is intake manifold temperature, 𝐸𝐸 is the EGR molar 
gas constant, 𝜂𝜂𝐶𝐶 is the cylinder volumetric efficiency, 𝑦𝑦 is the subscript of the intake molar gas 
quantity and 𝐶𝐶𝑠𝑠 is the piston swept volume. 
It’s imperative to mention here that in Equation 22, the EGR ratio has been defined as a function 
of the inlet intake manifold with existing measurable engine conditions including the fuel 
quantity delivered by the injection system and the fuel type used. Consequently, Equation 21 
can be manipulated in combination with other equations notably Equations 15, 16, 17 and 18, 
so that the volumetric concentration individual gas components in the combustion zones can 
be derived and defined. Hence the volumetric concentration for O2 becomes as follows: 
[𝑂𝑂 ] 𝑐𝑐𝑐𝑐+ 𝐸𝐸 �𝛾𝛾− 𝛼𝛼− 𝛽𝛽
 [𝐶𝐶 ] = 2 𝑓𝑓𝑓𝑓 2 4
� 𝐶𝐶f
2 𝐶𝐶𝐶𝐶𝑎𝑎        Equation 24 𝑐𝑐(1−𝐸𝐸)
[𝑂𝑂 ] 𝑐𝑐 +�𝛾𝛾−𝛼𝛼− 𝛽𝛽
[𝐶𝐶 ] 2 𝑖𝑖𝑖𝑖𝑎𝑎 2 4 
� 𝐶𝐶𝑓𝑓
2 𝑟𝑟𝑒𝑒ℎ =        Equation 25 𝑐𝑐 + �𝛾𝛾2+
𝛽𝛽
4� 𝐶𝐶𝑓𝑓
Where [𝐶𝐶2]𝐶𝐶𝐶𝐶𝑎𝑎 is the volumetric oxygen concentration at the inlet intake manifold, [𝐶𝐶2]𝑟𝑟𝑒𝑒ℎ is 
the volumetric oxygen concentration in the exhaust manifold, 𝑦𝑦 is the inlet intake molar gas 
quantity, 𝛾𝛾 is the fresh air fraction mass, 𝑛𝑛𝑓𝑓 is the cycle fuel molar quantity, 𝑧𝑧 is the remainder 
of the inlet intake air charge, [𝐶𝐶2]𝑓𝑓𝑟𝑟𝑐𝑐𝑐𝑐 is the volumetric concentration of fresh inlet oxygen in 
the combustion zone, 𝛽𝛽 is the reciprocal of the temperature of the EGR gases mean, 𝛼𝛼 is the 
thermal diffusivity of the cooling medium of the exhaust gases in the exhaust manifold. 
Since the EGR ratio influences the air excess ratio λ, it can be coupled with the EGR ratio and 
the engine in-cylinder availability of O2 especially in lean burn engines which diesel engines 
80 
 
are normally associated. Practically, λ is simply an indicator of the amount of oxygen present 
that can allow complete combustion relative to the fuel delivered by the system. Thus, as seen 
from Equation 24, the expression can be shown as: 
𝜆𝜆 𝑐𝑐[𝑂𝑂2]𝑖𝑖𝑖𝑖𝑎𝑎𝐶𝐶𝑐𝑐𝑐𝑐 = �𝛼𝛼+𝛽𝛽−𝛾𝛾          Equation 26 4 2� 𝐶𝐶𝑓𝑓
Where 𝜆𝜆𝐶𝐶𝑐𝑐𝑐𝑐 is the engine in cylinder air excess ratio, 𝑦𝑦[𝐶𝐶2]𝐶𝐶𝐶𝐶𝑎𝑎 is the inlet intake volumetric 
oxygen concentration molar gas quantity, 𝛽𝛽 is the mean reciprocal of the temperature of the 
EGR gases, 𝛼𝛼 is the thermal diffusivity, 𝑛𝑛𝑓𝑓 is the cycle fuel molar quantity, 𝛾𝛾 is the fresh air 
fraction mass.  
3.4.1 Modelling the EGR Valve Using the Orifice Equation 
The EGR valve is modelled as a function of the upstream pressure and downstream pressure 
of the EGR valve [10] by adopting the standard orifice flow equations as:  
𝑊𝑊𝑟𝑟𝑒𝑒𝑟𝑟 = 𝐶𝐶𝑅𝑅
𝑝𝑝𝑖𝑖
𝑟𝑟𝑓𝑓𝑓𝑓 Ѱ �
𝑝𝑝𝑗𝑗�  𝑖𝑖𝑖𝑖 𝑝𝑝𝑟𝑟<  𝑝𝑝𝐶𝐶      Equation 27 �𝐸𝐸𝐼𝐼𝑖𝑖 𝑝𝑝𝑖𝑖
𝑊𝑊𝑟𝑟𝑒𝑒𝑟𝑟 = 0 𝑖𝑖𝑖𝑖 𝑝𝑝1 = 𝑝𝑝2         Equation 28 
𝑊𝑊𝑟𝑟𝑒𝑒𝑟𝑟 =  𝐶𝐶𝑅𝑅
𝑝𝑝𝑗𝑗 𝑝𝑝𝑖𝑖
𝑟𝑟𝑓𝑓𝑓𝑓 Ѱ � �  𝑖𝑖𝑖𝑖 𝑝𝑝𝑟𝑟  > 𝑝𝑝𝐶𝐶      Equation 29 �𝐸𝐸𝐼𝐼𝑖𝑖 𝑝𝑝𝑗𝑗
Where the 𝑖𝑖 is the subscript of the upstream flow condition of the EGR cooler exit (upstream 
EGR cooler exit), 𝑗𝑗 is the subscript of the downstream conditions exhibited by the inlet intake 
manifold (downstream = intake manifold), 𝐶𝐶 is the discharge coefficient of the engine valve 
opening, 𝑅𝑅𝑟𝑟𝑓𝑓𝑓𝑓 is the effective flow area as a function the EGR valve relative to its position, Ѱ 
is the pressure ratio factor and 𝑝𝑝 is the pressure of the system and 𝛾𝛾 is the ratio of specific heat 
of the exhaust gas produced by the system or the engine, 𝑊𝑊𝑟𝑟𝑒𝑒𝑟𝑟 is the system flow of gases from 
the inlet intake manifold side to the engine cylinders available, with subscript 𝑘𝑘 which in this 
case has been substituted with subscript 𝑒𝑒𝑔𝑔𝑟𝑟 denoting EGR flow for all equations shown above, 
𝑅𝑅 is the temperature of the upstream gases. 
The following equations are adopted from the work of [10, 11] and give the relationship 
between the system and the pressure ratio correction factor and help in determining modelling 
the boundary conditions that will be available for the system, for the mass flow through the 
standard orifices. For the mass flow through an isothermal orifice assuming no losses during 
81 
 
acceleration of the fluid flowing through the system (in this case the air), and assuming that the 
entire potential energy is completely converted into thermal energy, the equation becomes: 
ṁ = 𝐶𝐶𝑅𝑅 𝑝𝑝𝑖𝑖𝑖𝑖 𝝍𝝍         Equation 30 
�𝐸𝐸𝐼𝐼𝑖𝑖𝑖𝑖
Where ṁ is the mass flow rate, 𝐶𝐶 is the discharge coefficient, 𝑅𝑅 is the effective area of flow, 
𝑃𝑃𝐶𝐶𝐶𝐶 is the pressure of the intake air, 𝐸𝐸 is the specific universal gas constant, 𝑅𝑅𝐶𝐶𝐶𝐶 is the 
temperature of the upstream. 
But if we define the 𝑃𝑃𝐶𝐶𝐶𝐶 and 𝑃𝑃𝑜𝑜𝑢𝑢𝑎𝑎 then this relationship becomes the ratio of pressure upstream 
and the ratio of pressure downstream through the restriction represented by the sign symbol П 
as the pressure ratio and is given in Equation 31 as: 
П = 𝑝𝑝𝑓𝑓𝑢𝑢𝑎𝑎          Equation 31 
𝑝𝑝𝑖𝑖𝑖𝑖
Where П is the pressure ratio, 𝑃𝑃𝑜𝑜𝑢𝑢𝑎𝑎 is the pressure of the ambient (atmospheric), 𝑃𝑃𝐶𝐶𝐶𝐶 is the 
pressure in the system. 
But the flow function with symbol 𝜓𝜓 is given by the following equations with two critical 
existing modelling conditions and thus becomes:  
𝑘𝑘+1
 𝜓𝜓 = �𝑘𝑘[ 2 ]𝑘𝑘−1  𝑖𝑖𝑖𝑖 𝑝𝑝𝑜𝑜𝑢𝑢𝑎𝑎  < 𝑝𝑝𝐶𝐶𝑟𝑟 (condition 𝑖𝑖)     Equation 32 𝑘𝑘+1
1 𝑘𝑘−1
𝜓𝜓 = П 2𝑘𝑘𝑘𝑘 � [1 − П 𝑘𝑘 ] 𝑖𝑖𝑖𝑖 𝑝𝑝𝑜𝑜𝑢𝑢𝑎𝑎 ≥ 𝑝𝑝𝐶𝐶𝑟𝑟 (condition 𝑖𝑖𝑖𝑖)    Equation 33 𝑘𝑘−1
𝑘𝑘
𝑝𝑝𝐶𝐶𝑟𝑟 =  𝑝𝑝𝐶𝐶𝐶𝐶 [
2 ]𝑘𝑘−1         Equation 34 
𝑘𝑘+1
Where 𝑃𝑃𝐶𝐶𝑟𝑟 is the critical pressure, 𝑃𝑃𝐶𝐶𝐶𝐶 is the pressure in the upstream of the system, 𝑃𝑃𝑜𝑜𝑢𝑢𝑎𝑎 is the 
pressure of the downstream of the system, П is the pressure ratio, 𝑘𝑘 is the thermal conductivity. 
However, in order to prevent the flow function with symbol (𝜓𝜓) from developing some kind 
of infinite gradient when the resultant value of the pressure ratio is approaching one or equaling 
one (П=1), the flow function is corrected using Equations 35 and 36 with pre-existing 
conditions as: 
𝜓𝜓 =  1  𝑖𝑖𝑖𝑖 𝑝𝑝 1𝑜𝑜𝑢𝑢𝑎𝑎  <  𝑝𝑝𝐶𝐶𝐶𝐶        Equation 35 √2 2
𝜓𝜓 = �2П [1 − П 𝑖𝑖𝑖𝑖 𝑝𝑝 1𝑜𝑜𝑢𝑢𝑎𝑎  ≥  𝑝𝑝𝐶𝐶𝐶𝐶       Equation 36 2
82 
 
Where 𝑃𝑃𝐶𝐶𝐶𝐶 is the pressure upstream, 𝑝𝑝𝑜𝑜𝑢𝑢𝑎𝑎 is the pressure downstream, 𝑅𝑅𝐶𝐶𝐶𝐶 is the temperature 
upstream, 𝑅𝑅 is the flow area, 𝐸𝐸 is the specific gas constant, 𝜓𝜓 is the flow function, ṁ is the 
mass flow measured in kg/s, 𝑝𝑝𝐶𝐶𝑟𝑟 is critical pressure that is achieved in the most narrowed 
section of the orifice device, 𝑘𝑘 is the 𝑐𝑐𝑝𝑝 / 𝑐𝑐𝑣𝑣 ( where 𝑐𝑐𝑝𝑝 and 𝑐𝑐𝑑𝑑 are the specific heat at constant 
pressure and volume in the system), П is the pressure ratio of the engine produced by the valves.  
Thus, the flow 𝜓𝜓 function is defined and calculated with two main existing assumption 
conditions as follows: 
(𝛾𝛾+1)
𝜓𝜓�𝑝𝑝𝑖𝑖� = { 𝛾𝛾( 2√ )2(𝛾𝛾−1) 𝑖𝑖𝑖𝑖 
𝑝𝑝𝑖𝑖  ≤ ( 2 )𝛾𝛾/(𝛾𝛾−1)     Equation 37 
𝑝𝑝𝑗𝑗 𝛾𝛾+1 𝑝𝑝𝑗𝑗 𝛾𝛾+1
2
𝜓𝜓 �pi � = � 2γ ( pi  )γ − (pi)(γ+1)/γ ) if pi  > ( 2 )γ/(γ−1)    Equation 38 
pj γ−1 pj pj pj γ+1
Where 𝜓𝜓 is the flow function, 𝑃𝑃𝐶𝐶 is the pressure of the 𝑖𝑖𝑎𝑎ℎ of the upstream pressure, 𝑃𝑃𝑟𝑟 is the 
pressure of the 𝑗𝑗𝑎𝑎ℎof downstream pressure, 𝛾𝛾 is the ratio of specific heats of the exhaust gas 
produced by the system or the engine. 
3.4.2 The Throttle Valve Equation 
Through the study of diesel engines, we learn that they are normally operated with fully opened 
throttle so as to allow high flow of EGR, through the use of a throttle actuator valve which 
allows certain operating conditions to generate a pressure drop in the intake manifold. For this 
reason, modelled equations take mostly the format of the mass flow equation through the 
orifice. However, because the throttle is never completely closed in diesel engines but leaves 
an aperture opening 𝜑𝜑0 as a bypass during operational running, the mass flow therefore through 
the throttle can be determined by using Equation 39 as: 
Ṁ =  𝐶𝐶 𝑅𝑅 𝑃𝑃𝑖𝑖𝑐𝑐𝑎𝑎ℎ 𝑎𝑎ℎ 𝑎𝑎ℎ 𝜓𝜓𝑎𝑎ℎ        Equation 39 �𝐸𝐸𝐼𝐼𝑖𝑖𝑐𝑐
Where Ṁ𝑎𝑎ℎ is the mass flow through the throttle, 𝑅𝑅𝑎𝑎ℎ is the throttle valve effective area, 𝐶𝐶𝑎𝑎ℎ is 
the discharge coefficient of the throttle valve where (𝐶𝐶𝑎𝑎ℎ = 1 ( assumed) and (0 ≤ 𝑢𝑢𝑎𝑎ℎ ≤ 1), 
 𝑃𝑃𝐶𝐶𝑎𝑎 is the intake manifold pressure, 𝑅𝑅𝐶𝐶𝐶𝐶𝑎𝑎 is the inlet intake air temperature, 𝜓𝜓 is the flow 
function through the throttle, 𝐸𝐸 is the specific gas constant. 
Thus, the throttle effective angle of the opening of the inlet intake throttle valve is given by the 
following equation as: 
83 
 
𝜑𝜑 = 𝑢𝑢𝑎𝑎ℎ �
𝜋𝜋 − 𝜑𝜑0� + 𝜑𝜑0        Equation 40 2
Where 𝜋𝜋 is pi 3.14, 𝜑𝜑 is the effective angle of opening of the throttle, 𝜑𝜑0 is the initial throttle 
angle of opening, 𝑢𝑢𝑎𝑎ℎ is the controlled scale variable of the throttle valve. 
But because the inlet intake throttle valve area varies as a function of the scaled control variable 
as per the second condition in equation 39 where 0 ≤ 𝑢𝑢𝑎𝑎ℎ ≤ 1, the effective area of the throttle 
valve is thus calculated as: 
2
𝑅𝑅 =  𝜋𝜋𝑑𝑑 𝑎𝑎ℎ �1 − 𝐶𝐶𝑂𝑂𝐸𝐸𝐶𝐶𝑎𝑎ℎ �        Equation 41 4 𝐶𝐶𝑂𝑂𝐸𝐸𝐶𝐶0
Where 𝑅𝑅𝑎𝑎ℎ is the throttle valve effective area, 𝜋𝜋 is pi 3.14, 𝑑𝑑 is the diameter of the throttle 
valve, 𝐶𝐶𝐶𝐶𝐷𝐷𝜑𝜑 is the final effective opening angle of the throttle valve, 𝐶𝐶𝐶𝐶𝐷𝐷𝜑𝜑0 is the initial 
effective angle of opening of the throttle valve. 
3.4.3 Fresh Air Fraction  
The fresh air fraction mass explains the amount of excess air associated with diesel engine 
operation. The diesel exhaust gases have and contain an amount of fresh air charge ṁ𝑎𝑎 which 
mathematically is defined as:  
 𝛾𝛾 = ṁ𝑎𝑎          Equation 42 
ṁ𝑓𝑓
Where ṁ𝑎𝑎 is the fresh air charge, 𝛾𝛾 is the fresh air fraction mass, ṁ𝑓𝑓 is the intake fuel flow 
rate. 
This equation thus forms part of the essential quantities for the calculation to define and 
determine the specific heat when 𝑐𝑐𝑝𝑝 and 𝑐𝑐𝑑𝑑 are constant in terms of pressure and volume 
respectively, and when they are to be expressed as a function of the air fraction especially in 
the inlet intake and exhaust manifolds as 𝛾𝛾𝐶𝐶𝑎𝑎 and 𝛾𝛾𝑟𝑟𝑎𝑎 respectively and calculated thus from the 
work of [11] as: 
𝐶𝐶. ,𝑟𝑟𝑎𝑎 = 𝛾𝛾𝑟𝑟𝑎𝑎�𝑐𝑐. ,𝑎𝑎− 𝐶𝐶. ,𝑟𝑟𝑒𝑒 � +  𝐶𝐶𝐶𝐶. ,𝑟𝑟𝑒𝑒      Equation 43 
Where 𝐶𝐶. ,𝑟𝑟𝑒𝑒 = (𝜆𝜆 = 1) is the specific heat of the exhaust gases, 𝐶𝐶. ,𝑎𝑎 is the fresh air mass 
fraction, 𝐶𝐶. ,𝐶𝐶𝑎𝑎 is the specific heat of the inlet intake manifold gases. 
While for the inlet intake manifold the equation then becomes: 
84 
 
𝐶𝐶. ,𝐶𝐶𝑎𝑎 = 𝛾𝛾𝐶𝐶𝑎𝑎(𝐶𝐶. ,𝑎𝑎− 𝐶𝐶. ,𝑟𝑟𝑎𝑎 ) + 𝐶𝐶. ,𝑟𝑟𝑎𝑎       Equation 44 
Where 𝐶𝐶. ,𝐶𝐶𝑎𝑎 is the specific heat of the inlet intake manifold gases, 𝐶𝐶. ,𝑎𝑎 is the fresh air mass 
fraction, 𝛾𝛾𝐶𝐶𝑎𝑎 is the fresh air fraction mass if the intake manifold.  
However, for the purposes of this experiment and the experimental set-up, the EGR % mass 
flow rate was calculated using the Equation 45 as: 
(𝐸𝐸𝐸𝐸𝐸𝐸 %) ṁ𝐸𝐸𝐸𝐸𝐸𝐸𝑟𝑟𝑓𝑓 = × 100        Equation 45 ṁ𝑖𝑖
But 
ṁ𝐶𝐶 = ṁ𝑎𝑎 + ṁ𝑓𝑓 + ṁ𝐼𝐼𝐸𝐸𝐸𝐸        Equation 46 
Where 𝐸𝐸𝐸𝐸𝐸𝐸 %𝑟𝑟𝑓𝑓 is the percentage mass flow rate of the exhaust gases, ṁ𝐶𝐶 is the mass of the 
inlet intake air, ṁ𝑎𝑎 is the mass of the ambient air, ṁ𝑓𝑓 is the mass of the injected fuel into the 
engine, ṁ𝐼𝐼𝐸𝐸𝐸𝐸 is the mass of the exhaust recirculated gas into the engine. 
3.5 The EGR Cooler and The Exchange of Heat in the Exhaust Manifold 
The EGR cooler can be modelled following the heat exchangers and thus assumed to be having 
a constant wall temperature 𝑅𝑅𝑚𝑚 as suggested by the works and equations of [11, 12]. 
 𝑥𝑥 = 0 𝑑𝑑𝑄𝑄
 
 
 𝑅𝑅(𝑥𝑥) 𝑅𝑅𝑟𝑟𝑎𝑎 
𝑅𝑅𝑟𝑟𝑒𝑒𝑟𝑟 
 𝑑𝑑𝑥𝑥 
 
𝑅𝑅𝑚𝑚 
 
Figure 3.2: The EGR cooler heat balance definitions and nomenclature 
 
Figure 3.2 shows the EGR cooler heat balance definitions and nomenclature, where 𝑅𝑅𝑚𝑚 is the 
wall temperature, 𝑅𝑅𝑟𝑟𝑎𝑎 is the exhaust manifold temperature,𝑅𝑅𝑟𝑟𝑒𝑒𝑟𝑟 is the temperatures of the EGR 
gases, 𝐿𝐿 is the length of the EGR cooler,𝑅𝑅𝑒𝑒 is the temperature as a function of 𝑥𝑥 in the cooler.  
85 
 
Looking at Figure 3.2 and utilizing a small control volume and assuming that the heat transfer 
ℎ and the specific heat 𝑐𝑐𝑝𝑝 at constant pressure is independent of the temperature 𝑅𝑅 and the 
function 𝑥𝑥, the heat flow into the cooler wall can be defined as: 
−𝑑𝑑𝑄𝑄 = 𝜋𝜋𝑑𝑑 ∗ 𝑑𝑑𝑥𝑥 ∗ ℎ[𝑅𝑅(𝑥𝑥) − 𝑅𝑅𝑚𝑚]       Equation 47 𝑑𝑑𝑎𝑎
Where −𝑑𝑑𝑄𝑄 is the rate of change of internal energy, 𝜋𝜋 is pi 3.14, 𝑑𝑑 is the diameter of the pipes 
𝑑𝑑𝑎𝑎
making up the EGR cooler, ℎ is the heat transfer, 𝑑𝑑𝑥𝑥 is the rate of change of x, 𝑅𝑅(𝑥𝑥) is the 
temperature as a function of x, 𝑅𝑅𝑚𝑚 is the wall temperature. 
But the change of internal energy of the EGR % mass flow rate is defined by Equation 48 as 
follows: 
𝑑𝑑𝑄𝑄 = 𝑑𝑑𝑅𝑅 × ṁ𝑟𝑟𝑒𝑒𝑟𝑟 × 𝑐𝑐𝑝𝑝        Equation 48 𝑑𝑑𝑎𝑎
Where 𝑑𝑑𝑄𝑄⁄𝑑𝑑𝑖𝑖 is the change in internal energy of the EGR % mass flow rate, 𝑑𝑑 is the diameter 
of the pipes making the EGR cooler, 𝑅𝑅 is the temperature, ṁ𝑟𝑟𝑒𝑒𝑟𝑟 is the mass flow rate of the 
EGR % gas, 𝑐𝑐𝑝𝑝 is the specific heat at constant pressure. 
In order to get the differential equation for the gas temperature Equations 47 and 48 are 
combined to become Equation 49 as: 
𝑑𝑑𝐼𝐼 = 𝜋𝜋𝑑𝑑ℎ (𝑅𝑅(𝑥𝑥) − 𝑅𝑅𝑚𝑚 )        Equation 49 𝑑𝑑𝑒𝑒 ṁ𝑒𝑒𝑒𝑒𝑓𝑓𝐶𝐶𝑝𝑝
Where 𝑑𝑑𝑅𝑅⁄𝑑𝑑𝑥𝑥 is the rate of change in temperature, 𝜋𝜋 is pi 3.14, ℎ is the heat transfer, 𝑑𝑑 is the 
diameter of the pipes making the EGR cooler, ṁ𝑟𝑟𝑒𝑒𝑟𝑟 is the mass flow rate of the EGR % gas, 
𝑐𝑐𝑝𝑝 is the specific heat at constant pressure. 𝑅𝑅(𝑥𝑥) is the temperature as a function of x, 𝑅𝑅𝑚𝑚 is the 
wall temperature. 
But at the point where 𝑥𝑥 = 𝐿𝐿 the temperature at the outlet of the EGR cooler corresponds to 
the solution of Equation 48. Therefore, the equation for 𝑅𝑅𝑟𝑟𝑒𝑒𝑟𝑟 is defined and expressed as a 
function of the length L as:  
𝑅𝑅(𝐿𝐿) = 𝑅𝑅 𝜋𝜋𝑑𝑑ℎ𝑟𝑟𝑒𝑒𝑟𝑟 = 𝑅𝑅𝑚𝑚 + (𝑅𝑅𝑟𝑟𝑎𝑎 − 𝑅𝑅𝑚𝑚)𝑟𝑟 −       Equation 50 ṁ𝑒𝑒𝑒𝑒𝑓𝑓𝐶𝐶𝑝𝑝
Where 𝑅𝑅(𝐿𝐿) is the temperature as a function of length, 𝑅𝑅𝑟𝑟𝑒𝑒𝑟𝑟 is the EGR gas temperature, 𝑅𝑅𝑚𝑚 is 
the wall temperature, 𝑅𝑅𝑟𝑟𝑎𝑎 is the exhaust manifold temperature, 𝑒𝑒 is the engine unit and its 
components, 𝜋𝜋 is pi 3.14, 𝑑𝑑 is the diameter of the pipes making the EGR cooler, ℎ is the heat 
86 
 
transfer, ṁ𝑟𝑟𝑒𝑒𝑟𝑟 is the mass flow rate of the EGR % gas, 𝑐𝑐𝑝𝑝 is the specific heat at constant 
pressure.  
When ℇ𝑟𝑟𝑒𝑒𝑟𝑟 is defined as the EGR cooler efficiency and as a function of the EGR % mass flow 
rate it is expressed as:  
ℇ = 1 − 𝑒𝑒 − 𝜋𝜋𝑑𝑑𝜋𝜋×ℎ𝑒𝑒𝑒𝑒𝑓𝑓𝑟𝑟𝑒𝑒𝑟𝑟         Equation 51 𝐶𝐶𝑝𝑝×ṁ𝑒𝑒𝑒𝑒𝑓𝑓
Where ℇ𝑟𝑟𝑒𝑒𝑟𝑟 is the EGR cooler efficiency, 𝑒𝑒 is the engine unit and components of internal 
recirculated EGR, ℎ𝑟𝑟𝑒𝑒𝑟𝑟 is the specific heat transfer of the EGR gases, 𝑑𝑑 is the diameter of the 
pipes making up the EGR cooler, 𝑐𝑐𝑝𝑝 is the specific heat at constant pressure, ṁ𝑟𝑟𝑒𝑒𝑟𝑟 is the mass 
flow rate of the EGR % gas, 𝜋𝜋 is pi 3.14, 𝐿𝐿 is the total length of the EGR cooler. 
But since the engine cooling medium used in this case is water, approximating 𝑅𝑅𝑟𝑟 as wall 
temperature using Equation 49 becomes as follows: 
𝑅𝑅𝑟𝑟𝑒𝑒𝑟𝑟 = 𝑅𝑅𝑟𝑟𝑎𝑎 − ℇ𝑟𝑟𝑒𝑒𝑟𝑟(𝑅𝑅𝑟𝑟𝑎𝑎 − 𝑅𝑅𝑟𝑟)       Equation 52 
Where 𝑅𝑅𝑟𝑟𝑒𝑒𝑟𝑟 is the EGR gas temperature, 𝑅𝑅𝑟𝑟𝑎𝑎 is the exhaust manifold temperature, ℇ𝑟𝑟𝑒𝑒𝑟𝑟 is the 
EGR cooler efficiency, 𝑅𝑅𝑟𝑟 is the engine unit temperature. 
However, since in this case the heat transfer is still unknown for the internal convection 
between the heat transferring and exchanging fluids and the wall of the EGR cooler, we have 
to identify it by using the 𝑁𝑁𝑢𝑢𝑔𝑔𝑔𝑔𝑒𝑒𝑁𝑁𝑖𝑖 𝑛𝑛𝑢𝑢𝑛𝑛𝑎𝑎𝑒𝑒𝑟𝑟 (𝑁𝑁𝑢𝑢), it should be mentioned here that the heat 
transfer can be calculated by two methods using the 𝑃𝑃𝑟𝑟𝑎𝑎𝑛𝑛𝑑𝑑𝑖𝑖𝑁𝑁 𝑛𝑛𝑢𝑢𝑛𝑛𝑎𝑎𝑒𝑒𝑟𝑟 and 𝑁𝑁𝑢𝑢𝑔𝑔𝑔𝑔𝑒𝑒𝑁𝑁𝑖𝑖 𝑛𝑛𝑢𝑢𝑛𝑛𝑎𝑎𝑒𝑒𝑟𝑟. 
The 𝑁𝑁𝑢𝑢𝑔𝑔𝑔𝑔𝑒𝑒𝑁𝑁𝑖𝑖 𝑛𝑛𝑢𝑢𝑛𝑛𝑎𝑎𝑒𝑒𝑟𝑟 is preferred due to its values remaining almost constant for all gases as 
expressed in Equations 53 and 54 as two definitions: 
𝑁𝑁𝑢𝑢 =  ℎ𝑒𝑒𝑒𝑒𝑓𝑓𝜋𝜋          Equation 53 
𝑘𝑘
𝑁𝑁𝑢𝑢 = 𝐶𝐶0 × 𝐸𝐸𝑒𝑒𝐶𝐶𝑖𝑖         Equation 54 
Where 𝑐𝑐𝐶𝐶 is the empirical constant, 𝑘𝑘 is the thermal conductivity of the exhaust gases, 𝐿𝐿 is the 
total length of the EGR cooler, ℎ𝑟𝑟𝑒𝑒𝑟𝑟 is the specific heat transfer of the EGR gases, 𝐸𝐸𝑒𝑒 is the 
Reynold’s number, 𝑁𝑁𝑢𝑢 is the Nusselt number.  
But for circular pipes which the exhaust manifold and the calorimeter can be classified as one 
the Reynolds number is defined and expressed in Equation 55 as: 
87 
 
𝐸𝐸𝑒𝑒 = 𝜌𝜌𝑒𝑒𝑒𝑒𝑓𝑓𝜐𝜐∞𝑑𝑑𝑒𝑒𝑒𝑒𝑓𝑓 = ṁ𝑒𝑒𝑒𝑒𝑓𝑓𝑑𝑑𝑒𝑒𝑒𝑒𝑓𝑓 = 4ṁ𝑒𝑒𝑒𝑒𝑓𝑓           Equation 55 
𝜇𝜇𝑒𝑒𝑒𝑒𝑓𝑓 𝜇𝜇𝑒𝑒𝑒𝑒𝑓𝑓𝐼𝐼𝑒𝑒𝑒𝑒𝑓𝑓 𝜋𝜋𝑑𝑑𝑒𝑒𝑒𝑒𝑓𝑓𝜇𝜇𝑒𝑒𝑒𝑒𝑓𝑓
Where 𝜇𝜇𝑟𝑟𝑒𝑒𝑟𝑟 is the dynamic viscosity, 𝜌𝜌𝑟𝑟𝑒𝑒𝑟𝑟 is the density of the EGR gases, 𝜐𝜐∞ is the velocity 
of the exiting exhaust gases and 𝑅𝑅𝑟𝑟𝑒𝑒𝑟𝑟 is the effective flow area of the EGR cooler, ṁ𝑟𝑟𝑒𝑒𝑟𝑟 is the 
mass flow rate of the EGR % gas, 𝜋𝜋 is pi 3.14, 𝑑𝑑𝑟𝑟𝑒𝑒𝑟𝑟 is the diameter of the EGR valve, 𝐸𝐸𝑒𝑒 is 
the Reynold’s number. If we combine Equations 54 and 55 we can now easily define the total 
heat coefficient of the EGR gases as: 
ℎ =  𝑘𝑘𝑒𝑒𝑒𝑒𝑓𝑓 𝑁𝑁𝑢𝑢 = 𝑘𝑘𝑒𝑒𝑒𝑒𝑓𝑓𝐶𝐶0𝑟𝑟𝑒𝑒𝑟𝑟 (
4ṁ𝑒𝑒𝑒𝑒𝑓𝑓 )𝐶𝐶𝑖𝑖 = 𝛼𝛼 × ṁ𝑟𝑟𝑒𝑒𝑟𝑟𝛽𝛽            Equation 56 𝜋𝜋𝑒𝑒𝑒𝑒𝑓𝑓 𝜋𝜋𝑒𝑒𝑒𝑒𝑓𝑓 𝜋𝜋𝑑𝑑𝑒𝑒𝑒𝑒𝑓𝑓𝜇𝜇𝑒𝑒𝑒𝑒𝑓𝑓
Where ℎ𝑟𝑟𝑒𝑒𝑟𝑟 is the coefficient of the EGR gases, 𝑘𝑘𝑟𝑟𝑒𝑒𝑟𝑟 is the thermal conductivity of the EGR 
gases, 𝐿𝐿𝑟𝑟𝑒𝑒𝑟𝑟 is the total distance of travel across the cooler of the EGR gases, 𝑁𝑁𝑢𝑢 is the 𝑁𝑁𝑢𝑢𝑔𝑔𝑔𝑔𝑒𝑒𝑁𝑁𝑖𝑖 
number, 𝐶𝐶0/𝐶𝐶𝐶𝐶 are empirical constants,𝑑𝑑𝑟𝑟𝑒𝑒𝑟𝑟 is the diameter of the EGR valve, 𝜇𝜇𝑟𝑟𝑒𝑒𝑟𝑟 is the 
dynamic viscosity of the EGR gases, ṁ𝑟𝑟𝑒𝑒𝑟𝑟 is the EGR mass flow rate, 𝛼𝛼 is the thermal 
diffusivity, 𝛽𝛽 is the reciprocal of the temperature of the EGR gases mean. 
3.5.1 The Axial Heat Transfers in the Exhaust Manifold 
 
Figure 3.3: Axial and radial components of the heat transfer model of the exhaust pipe and its nomenclature 
 
Figure 3.3 shows the axial and radial components of the heat transfer model of the exhaust 
pipe. The exchange of the burnt gases between the engine cylinders’ outlet and exhaust 
manifold is affected by mostly three types of heat factors i.e. heat along the exhaust pipe which 
occurs through conduction, the loss of heat from the pipe to the ambient environment through 
radiation and etc., and finally the loss of energy due to dissipation of radiation and convection. 
The nomenclature shown on Figure 3.3 indicates that the system is modelled in static mode 
where it is assumed that the inside wall of the pipe seems to operate with almost the same 
88 
 
values of temperature as the exhausting gases (this assumes there is no loss through internal 
structural convection). This is expressed in equation form as: 
𝑄𝑄𝑟𝑟𝑎𝑎 = 𝑄𝑄𝑎𝑎𝑒𝑒 + 𝑄𝑄𝑟𝑟𝑎𝑎                              Equation 57 
Where 𝑄𝑄𝑎𝑎𝑒𝑒 is the heat of axial direction relative to the engine, 𝑄𝑄𝑟𝑟𝑎𝑎 is the heat of radial direction 
flowing out relative to the ambient and 𝑄𝑄𝑟𝑟𝑎𝑎  is the heat from the exhaust manifold. 
But 𝑄𝑄, which here is the heat flux from the exhaust manifold with temperature component 𝑅𝑅1 
to the heat sink, which in this case is the cooling system coolant with temperature 𝑅𝑅2 across a 
surface area A, is expressed and defined as: 
𝑄𝑄 = 𝑅𝑅ℎ(𝑅𝑅1 − 𝑅𝑅2)                   Equation 58 
Where 𝑄𝑄 is the heat flux from the exhaust manifold, 𝑅𝑅 is the area, ℎ is the heat transfer, 𝑅𝑅1 is 
the temperature of the heat sink, 𝑅𝑅2 is the temperature of the cooling medium.  
However the heat normally in axial flow direction from the exhaust manifold wall relative to 
the engine block is in this case assumed to have the same temperature as that of the engine 
exhaust manifold, where the 𝑅𝑅𝑟𝑟𝑎𝑎 is the main source of the heat with the engine cooling system 
acting as the heat sink, thus defining the heat flux as:  
𝑄𝑄𝑎𝑎𝑒𝑒 = 𝑅𝑅𝑠𝑠𝑟𝑟𝐶𝐶ℎ𝑎𝑎𝑒𝑒(𝑅𝑅𝑟𝑟𝑎𝑎 − 𝑅𝑅𝑟𝑟)           Equation 59 
Where 𝑅𝑅𝑠𝑠𝑟𝑟𝐶𝐶 is the cross sectional area through which the gaseous exchange occurs in the axial 
direction and does correspond to the number of cylinder contained by the model engine, 𝑄𝑄𝑎𝑎𝑒𝑒 
is the axial heat transfer, ℎ𝑎𝑎𝑒𝑒 is the axial heat transfer coefficient, 𝑅𝑅𝑟𝑟 is the temperature of the 
engine and 𝑅𝑅𝑟𝑟𝑎𝑎 is the exhaust manifold temperature. 
From Equation 57 we can deduce a very important factor which is resistance to the flow of the 
gases along the exhaust manifold. The thermal resistance here consists of the conductive 
resistance 𝑟𝑟 , which is the reciprocal of or inversely proportional to ℎ (the heat transfer 
coefficient) and defined in two Equations 60 and 61 as: 
𝑟𝑟 = 1                           Equation 60 
ℎ
Where 𝑟𝑟 is the resistance to flow along the exhaust manifold, ℎ is the heat transfer along the 
exhaust manifold. 
But it can also be expressed as: 
89 
 
𝑄𝑄𝑎𝑎𝑒𝑒 =  
1                  Equation 61 
ℎ𝑎𝑎𝑎𝑎
Where 𝑄𝑄𝑎𝑎𝑒𝑒 is the axial heat transfer, ℎ𝑎𝑎𝑒𝑒 is the axial heat transfer coefficient. 
𝑟𝑟𝑎𝑎𝑒𝑒 
𝑅𝑅𝑟𝑟𝑎𝑎 𝑅𝑅𝑟𝑟 
 
Figure 3.4: Axial heat resistance model for the exhaust pipe with its nomenclature 
Figure 3.4 is showing the axial heat resistance model for the exhaust pipe with its nomenclature. 
𝑅𝑅𝑟𝑟𝑎𝑎 is the temperature of the exhaust manifold, 𝑟𝑟𝑎𝑎𝑒𝑒 is axial heat resistance within the system, 
𝑅𝑅𝑟𝑟𝐶𝐶𝑒𝑒 is the temperature from the engine itself. 
However, it should be noted at this point that the coefficient of heat transfer has so many 
methods of derivation from many geometric calculation and models, but this equation 
expression was chosen as it is a more direct expression of coefficient heat transfer and is 
expressed in Equation 62 as: 
ℎ 1𝑎𝑎𝑒𝑒 =  ⁄                  Equation 62 𝜋𝜋 𝐶𝐶𝐶𝐶𝑐𝑐𝑐𝑐
Where 𝐿𝐿⁄𝑛𝑛𝑐𝑐𝑦𝑦𝑁𝑁 is the overall length of the exhaust pipe equally divided to the number of 
cylinder the engine contains, ℎ𝑎𝑎𝑒𝑒 is the axial heat transfer coefficient. 
90 
 
3.5.2 The Radial Heat Exchange in the Exhaust Manifold  
 
Figure 3.5: Modelled radial heat resistance of the exhaust pipe 
 
Figure 3.5 shows the modelled radial heat resistance of the exhaust pipe, where all the thermal 
resistance is modelled and calculated like an electrical system in which the circuits are arranged 
in a series and parallel configurations. However, it is important to mention that the internal 
structural convection heat transfer has been neglected, thus the total overall thermal radial 
resistance as expressed in Equation 63 does not include internal structural convection heat 
transfer and is expressed as: 
𝑟𝑟 = 𝑟𝑟 + ( 1 + 1 )−1𝑟𝑟𝑎𝑎 𝐶𝐶𝑑𝑑         Equation 63 𝑟𝑟𝑐𝑐𝑐𝑐 𝑟𝑟𝑓𝑓𝑎𝑎𝑓𝑓
Where 𝑟𝑟𝑟𝑟𝑎𝑎 is the resistance due to the radial heat transfer, 𝑟𝑟𝐶𝐶𝑑𝑑 is the resistance due to conduction 
heat transfer, 𝑟𝑟𝐶𝐶𝑑𝑑 is the resistance due to the convectional heat transfer, 𝑟𝑟𝑟𝑟𝑎𝑎𝑑𝑑 is the resistance 
due to the radial heat transfer. 
3.5.3 The Calculation of Coefficients of Heat Transfer  
The coefficient of heat transfer can be calculated using different methods and formulas, 
namely, the radial conduction method, the free convection method which utilizes the Raleigh 
number (Ra), the Grashofs number (Gr) and the already mentioned Prandtl number (Pr), as per 
[11, 13]. 
The Raleigh number is expressed in the following equation:  
3
𝐸𝐸𝑎𝑎 = 𝐸𝐸𝑟𝑟 × 𝑃𝑃𝑟𝑟 = 𝑒𝑒𝛽𝛽(𝐼𝐼−𝐼𝐼∞)𝑑𝑑 × 𝑑𝑑       Equation 64 
𝑑𝑑2 𝛼𝛼
91 
 
Where the 𝐸𝐸𝑎𝑎 is the Raleigh number, 𝛼𝛼 is the thermal diffusivisity through the cooling medium, 
𝑣𝑣 is the kinematic viscosity of the air medium, 𝑔𝑔 is the force of gravitational acceleration acting 
on the cooling medium, 𝛽𝛽 is the reciprocal of the temperature mean between the source (in this 
case the exhaust manifold) and the heat sink ( cooling medium air). 𝐸𝐸𝑟𝑟 is the Grashofs number, 
𝑃𝑃𝑟𝑟 is the Prandtl number.  
Thus:  
β = 2           Equation 65 
𝐼𝐼+𝐼𝐼∞
Where 𝛽𝛽 is the reciprocal of the temperature mean between the source (in this case the exhaust 
manifold) and the heat sink (cooling medium air), 𝑅𝑅∞ is the temperature of the exiting gases. 
But for the exhaust manifold 𝑅𝑅 = 𝑅𝑅𝑟𝑟𝑎𝑎𝑅𝑅∞ = 𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏 and 𝑑𝑑 = 𝑑𝑑𝑜𝑜𝑢𝑢𝑎𝑎 and the coefficient of heat 
transfer therefore is calculated using the Nusselt number in its definition as: 
ℎ𝐶𝐶𝑑𝑑 = 𝑁𝑁𝑢𝑢 ×
𝑘𝑘𝑎𝑎          Equation 66 
𝑑𝑑𝑓𝑓𝑢𝑢𝑎𝑎
Where ℎ𝐶𝐶𝑑𝑑 is the heat transfer coefficient due to the convection heat transfer, 𝑁𝑁𝑢𝑢 is the Nusselt 
number, 𝑘𝑘𝑎𝑎𝑎𝑎𝑏𝑏 = the 𝑐𝑐𝑝𝑝 / 𝑐𝑐𝑣𝑣 (where 𝑐𝑐𝑝𝑝 and 𝑐𝑐𝑑𝑑 are the specific heat at constant pressure and 
volume in the system of the ambient air), 𝑑𝑑𝑜𝑜𝑢𝑢𝑎𝑎 is the diameter of the exhaust manifold.  
However, the Nusselt number as defined here can also be described and determined by just 
following the empirical formula which defines it as: 
1
𝑁𝑁𝑢𝑢 = (0.6 + 0.387𝐸𝐸𝑎𝑎
6 2
0.559 9 8
)         Equation 67 
[1+( )16]27𝑃𝑃𝑓𝑓
But looking at the energy of radiation exhibited in this process and heat transfer processes, the 
heat of radiation transfer then can be defined as:  
𝑄𝑄 = 𝑅𝑅𝑠𝑠ℎ𝜃𝜃𝜃𝜃(𝑅𝑅 − 𝑅𝑅∞)4        Equation 68 
Where 𝜃𝜃 = 5.6704 ∗ 10−8 is the Stefan-Boltzman constant, 𝑅𝑅𝑠𝑠ℎ is the area of the shell of the 
cylinder, 𝜃𝜃 is the measure of the emissivity angle relative to the black surface, 0.6 is the angle 
of emissivity for the exhaust manifold, 𝑅𝑅 =  𝑅𝑅𝑟𝑟𝑎𝑎 is exhaust manifold temperature and 𝑅𝑅∞ =
𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏 is the ambient temperature.  
92 
 
However, if we substitute into the power in the bracket in Equation 68 by using the products 
of exhaust manifold, where 𝑅𝑅 = 𝑅𝑅𝑟𝑟𝑎𝑎 and 𝑅𝑅∞ = 𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏, the radiation and heat transfer equation 
becomes: 
𝑄𝑄 2 2𝑟𝑟𝑎𝑎𝑑𝑑 = 𝑅𝑅𝑠𝑠ℎ𝜃𝜃𝜃𝜃(𝑅𝑅𝑟𝑟𝑎𝑎 + 𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏)(𝑅𝑅𝑟𝑟𝑎𝑎 + 𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏)(𝑅𝑅𝑟𝑟𝑎𝑎 − 𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏)    Equation 69 
Where 𝑅𝑅𝑠𝑠ℎ is the area of the shell of the cylinder, 𝑅𝑅𝑟𝑟𝑎𝑎 is the exhaust manifold temperature, 
𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏 is the ambient temperature, 𝑄𝑄𝑟𝑟𝑎𝑎𝑑𝑑 is the heat flux due to the radial heat transfer, 𝜃𝜃 is the 
Stefan-Boltzman constant, 𝜃𝜃 is the measure of the emissivity angle relative to the black surface. 
This equation thus makes it possible for the coefficient of heat of radiation transfer process to 
be calculated as in Equation 70 as: 
ℎ 2 2𝑟𝑟𝑎𝑎𝑑𝑑 = 𝜃𝜃𝜃𝜃(𝑅𝑅𝑟𝑟𝑎𝑎 + 𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏)(𝑅𝑅𝑟𝑟𝑎𝑎 + 𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏)      Equation 70 
Where ℎ𝑟𝑟𝑎𝑎𝑑𝑑 is the heat transfer coefficient due to radial heat transfer, 𝜃𝜃 is the measure of the 
emissivity angle relative to the black surface, 𝑅𝑅𝑟𝑟𝑎𝑎 is the exhaust manifold temperature, 𝑅𝑅𝑎𝑎𝑎𝑎𝑏𝑏 
is the ambient temperature, 𝜃𝜃 is the Stefan-Boltzman constant. 
References 
 
[1] G. Paul, A. Datta, and B. K. Mandal, "An experimental and numerical investigation of 
the performance, combustion and emission characteristics of a diesel engine fueled with 
jatropha biodiesel," Energy Procedia, vol. 54, pp. 455-467, 2014. 
[2] P. Divekar, Q. Tan, X. Chen, and M. Zheng, "Characterization of Exhaust Gas 
Recirculation for Diesel Low Temperature Combustion," IFAC-PapersOnLine, vol. 48, 
pp. 45-51, 2015. 
[3] M. Hamdan and R. H. Khalil, "Simulation of compression engine powered by 
Biofuels," Energy Conversion and Management, vol. 51, pp. 1714-1718, 2010. 
[4] M. Hawi, R. Kiplimo, and H. Ndiritu, "Effect of exhaust gas recirculation on 
performance and emission characteristics of a diesel-piloted biogas engine," Smart Grid 
and Renewable Energy, vol. 6, p. 49, 2015. 
[5] J. B. Heywood, "Internal combustion engine fundamentals," 2012. 
[6] S. R. Turns, An introduction to combustion vol. 287: McGraw-hill Education 
India,Private limited,New Delhi, 2012. 
[7] K. C. Rolle, Thermodynamics and heat power, 6 ed. Upper Saddle River, NJ: 
Pearson/Prentice Hall, 2005. 
[8] C. R. Ferguson and A. T. Kirkpatrick, Internal combustion engines: applied 
thermosciences: John Wiley & Sons, 2015. 
[9] L. Guzzella and C. Onder, Introduction to modeling and control of internal combustion 
engine systems: Springer Science & Business Media, 2009. 
[10] L. Kocher, E. Koeberlein, K. Stricker, D. Van Alstine, B. Biller, and G. M. Shaver, 
"Control-oriented modeling of diesel engine gas exchange," in American Control 
Conference (ACC), 2011, 2011, pp. 1555-1560. 
93 
 
[11] E. Alfieri, "Emissions-controlled diesel engine," 2009. 
[12] T. L. Bergman and F. P. Incropera, Fundamentals of heat and mass transfer: John 
Wiley & Sons, 2011. 
[13] M. F. Modest and D. C. Haworth, Radiative Heat Transfer in Turbulent Combustion 
Systems: Theory and Applications: Springer, 2016. 
 
 
 
 
 
 
 
 
94 
 
CHAPTER FOUR 
THE EFFECTS OF EXHAUST GAS RECIRCULATION ON 
THE PERFORMANCE AND EMISSION CHARACTERISTICS 
OF A DIESEL ENGINE – A CRITICAL REVIEW 
Journal article 
Maroa, S., & Inambao, F. (2017). The Effects of Exhaust Gas Recirculation on the 
Performance and Emission Characteristics of a Diesel Engine–A Critical Review. International 
Journal of Applied Engineering Research, 12(23), 13677-13689. 
http://www.ripublication.com 
 
  
 
 
95 
 
International Journal of Applied Engineering Research ISSN 0973-4562 Volume 12, Number 23 (2017) pp. 13677-13689 
© Research India Publications.  http://www.ripublication.com 
The Effects of Exhaust Gas Recirculation on the Performance and Emission 
Characteristics of a Diesel Engine – A Critical Review 
 
Maroa Semakula1and Prof Freddie Inambao2 
University of KwaZulu-Natal, Durban South-Africa. 
Orcid: 0000-0001-9922-5434 
 
Abstract engines. Other benefits of EGR, particularly cooled EGR, is 
that it enables more EGR flow and cooler intake charge 
Exhaust gas recirculation (EGR) is one of the most widely used 
temperatures, thus reducing peak in-cylinder temperatures 
methods of controlling emissions, particularly nitrogen oxide 
which improves ignition quality in diesel engines; reduction in 
(NOX), CO, CO2, HC (hydrocarbons) and particulate matter 
lubricating oil consumption; increase in fuel injection 
(PM). EGR is a system where portions of the exhaust gases are 
pressures; increase in use of diesel oxidation catalysts; and 
re-introduced back into the combustion chamber via the intake 
increased intake manifold boost pressure. Currently, more 
manifold. The systems that are affected by the application of 
efficiency is being demanded in the management systems of 
EGR are: brake specific fuel consumption (BSFC), lubricating 
diesel engines as a way of reducing energy losses, particularly 
system, diesel engine combustion characteristics, the brake 
those being wasted through the exhaust system which typically 
thermal efficiency (BTE), the combustion characteristics of DI 
wastes between 20% to 25%  of all energy released. Thus, there 
diesel engines, engine speed, CO2 and HC emissions, the 
has been tremendous development of ways or methods to 
performance characteristics of turbocharged diesel engine, 
continually reduce this energy wasted through re-tapping of the 
fuels, temperature, on EGR rates on the diesel engine 
exhaust gases by utilizing EGR systems, for example, in 
performance and emission, the durability of engine 
scavenging for pre-heating purposes. This recycling of waste 
components, and  cooled EGR and the challenges of cooled 
gate gases improves fuel efficiency economy, volumetric 
EGR. 
efficiency, brake thermal efficiency, brake specific fuel 
Keywords: EGR, BSFC, Carbon dioxide, PM, direct injection, consumption, and reduces emissions and PM [2]. 
brake thermal efficiency. 
The phenomenal growth of the middle class of most emerging 
 and developing countries, and the failure of most public 
transport systems and infrastructure in developing and 
INTRODUCTION 
emerging nations, has tended to increase the number of 
The exhaust gas recirculation (EGR) technique is gaining personalized transport vehicles on the road, with diesel 
widespread use as one of the most efficient methods or propelled vehicles gaining popularity including in day to day 
techniques for reducing emissions and particulate matter (PM) commercial activities. This growth has produced many studies 
effluents from diesel automobiles, particularly nitrogen dioxide looking into ways of reducing energy spending through 
(NOX), carbon monoxide (CO), and hydrocarbons (HC). This harnessing of systems gases and redirecting and harvesting the 
involves the circulation of part of the exhaust gas along with available energy within the gases which would have otherwise 
the intake fresh air charge, into the combustion chamber of the have been lost. This reduces demands on the primary sources 
diesel engine for combustion together with the fuel injected in of energy experienced in the world today. Although there have 
the normal power cycle of the diesel engine. EGR is an been tremendous achievements in technological advances in 
emission control system which allows significant reduction in automobile manufacture and technology, there is a need for us 
NOX gases in almost all types of diesel engines including Light to continue to reduce the amount of energy use and emissions, 
Duty, Medium Duty and Heavy Duty engine use or thereby reducing pollution in our primary and atmospheric 
applications, and primary engines as low as two stroke as used environment. These studies and technological changes can 
in marine operations and applications. initiate a reduction in global warming, which has seen a 
phenomenal growth in 21st century. It is imperative, therefore, 
The use of EGR systems in diesel engines is not limited to 
for automobile and engine manufacturing industries to 
increasing efficiency and reduction of emissions only [1], but 
incorporate these new studies regarding reduction in emissions 
also includes imparting knock resistance and reduction in the 
and energy wastage from engine exhaust systems and 
need for high load fuel enrichment in diesel engines. However, 
undertake deliberate measures to improve on and harness EGR 
in spark ignition (SI) engines EGR systems also aide in or help 
system gases [1].  
in the vaporization of liquid fuels, though it is an enabler for 
closed cycle diesel engines in particularly improving ignition Agarwal, Singh and Agarwal [3] state that despite the high fuel 
quality and reduction of high load fuel enrichment in diesel efficiency provided by diesel engines, they have the 
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© Research India Publications.  http://www.ripublication.com 
disadvantage of emitting nitrogen NOx and PM which is engine motor speed is increased. However, in increasing the 
considered as the main pollutants and main toxic gases emitted load and engine speed he noted that the exhaust gas 
to the atmosphere and as significant contributors to air temperature increased, whereas increasing the EGR% mass 
pollution in major urban centres and modern-day cities. Other flow rate reduced the temperatures (Figure 1). 
challenges or disadvantages include: more HCs and soluble 
organic fractions (SOF), PM, more condensation HC and PM 
deportation in cooler fouling (i.e. degraded heat transfer and 
higher flow resistance). These emissions are currently 
legislated by stringent environmental protection agencies in 
most countries, making diesel propelled engines and machines 
more expensive and less appealing. However, with the 
introduction of the EGR system, NOx emissions of diesel 
engines are greatly reduced making it increasingly compelling 
to incorporate such a system in the design and manufacture of 
diesel engines.  
 
THE REVIEW 
A number of studies have been published regarding the effects 
of (EGR on the performance and emission characteristics of 
diesel engines. Several of these have been considered and are  
reviewed here to highlight the developments in this area of 
Figure 1: Abaas (2016) 
study. 
 
Abaas [4] in his study “Effect of Exhaust Gas Recirculation 
(EGR) on the Performance Characteristics of a Direct Injection Figure 1 Showing the relationship between the EGT and BSFC 
Multi Cylinders Diesel Engine,” used a direct injection (DI), and application of EGR.  
four-cylinder FIAT engine, coupled to a dynamometer and 
The increase in temperature is explained by the fact that more 
thermocouple with Ni-Cr and Ni-Al elements, air and fuel flow 
load and speed need more fuel to be burned; by increasing the 
meters, together with engine speed tachometers. Fuel used 
load the work of dilution, as already explained, increases as 
during the test was Iraqi conventional diesel CN = 46.8 under more combustion reactants are exposed to the low temperature 
varying EGR rates of 5%, 10%, 20%, and 30%. Engine speed 
environment. Thus, increasing EGR% mass flow rate reduced 
was maintained at constant for all loads and variable speeds of 
exhaust gas temperature (EGT) by 1.41%, 15.64%, 16.08% and 
1250 rpm to 3000 rpm at constant engine load were used.  
26.38% for EGR% mass flow rate of 5%, 10%, 20% and30% 
Abaas [4] noted that increasing the EGR percentage or mass respectively (Abaas, 2016). On the brake power (BP), Abaas 
flow rate increased the brake specific fuel consumption [4] noted an increase as the engine speed increases. However, 
(BSFC). Other findings on the BSFC were that at low engine by increasing the percentage of the EGR mass flow rate, the 
load ranges the BSFC was found to be high as compared to brake power reduced 1.12%, 3.26%, 5.42%, and 16.79% when 
when medium engine load ranges were used. The researcher compared to EGR% mass flow rates of 5%, 10%, 20% and 
attributed this to the drop in combustion chamber temperature. 30%, which indicate true effects of dilution in the combustion 
However, with reducing temperature the researcher observed chamber and on the performance of a diesel engine. 
increased fuel delay period, thus reducing engine peak pressure 
Charola, Makwana and Makwana [5] in their study “Effects of 
resulting in lower engine output and a greater increase in the 
EGR (Exhaust Gas Recirculation) on the Performance and 
fuel consumed. Therefore, according to this study, the brake 
Emission of a CI Compression Ignition Engine Fueled with a 
thermal efficiency reduces with an increase in EGR % mass 
Blend” used a single cylinder, four stroke, water compression 
flow rate application. 
ignition engine, with the test fuel being methanol (industrial 
Other findings from Abaas [4] include the effect of reduction chemical) with purity or concentration level of 99%. The test 
in oxygen availability during combustion, which is the factor set up consisted of a test engine, a hydraulic dynamometer, fuel 
explaining the relative incomplete combustion and the increase and flow meters, and incorporated various measuring 
in particulate matter (PM), but credited with the reduction of equipment and relevant modifications for test study. 
NOX emissions. On volumetric efficiency, Abaas [4] argues In their findings, the BTE (brake thermal efficiency) increased 
that at low engine loads, since diesels operate on lean mixture, 
as the load increased. Using 10% methanol diesel blend they 
the volumetric efficiency was observed to be very high, but 
found that it increased the BTE by up to 12.85% and a 20% 
increasing the engine load. He observed that the engine 
blend of methanol and diesel improved the BTE by 26.23% on 
consumed more fuel in order to meet the load demands as the 
the level of 10% EGR, while in BSFC (brake specific fuel 
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© Research India Publications.  http://www.ripublication.com 
consumption) they observed a reduction of 3.57% in the BSFC the heat specific ratio of the intake air mixture by causing less 
in a 10%/90% methanol and diesel blend, and an increase of oxygen in the combustion chamber. The NOX emission 
4.28% when the 20%/80% was used [4]. On how the methanol findings indicate that for all the fuel blends used during the 
diesel blend affects emissions of CO, they found that with the study, the NOX emissions decreased with the increase in EGR% 
10%/90% methanol diesel blend there was a 30% reduction in mass flow rate. This is explained by two factors: reduction in 
the emissions of CO, while with the 20%/80%, methanol diesel the availability of oxygen during combustion and the lower 
blend there was a reduction of 40% during the experiment combustion temperatures. But since NOX is temperature 
study period. Regarding the HC in the exhaust system, it was dependent for its formation to take place, this explains its 
observed that there was a reduction of 13.04% in the 10%/90% decrease as flame temperatures drop. The HC and CO emission 
methanol diesel blend and a reduction of 17.39% with the showed an increase with the increase of the EGR% mass flow 
20%/80% methanol diesel blend. The NOX emission showed a rate above the 15% EGR rate level, as less oxygen becomes 
reduction of 5.55% with the 10%/90% methanol diesel blend, available for combustion, resulting in incomplete combustion 
and 3.33% in the 20%/80% methanol diesel blend [4]. The and more HC and PM.  
exhaust temperature gas reduction was observed to reduce or 
Kumar, Antony and Sahoo [7] In the “Effects of Varying EGR 
drop by 4°C, when the 10% EGR mass flow rate was used, 
Rates in The Performance and NOX Emission of an IDI Diesel 
compared to 6°C when the EGR% mass flow rate was applied. 
Engine Fueled with JB100, JB80, JB60, JB40, JB20 and 
Rao, Mutyalu and Ramakrishna [6] in “Effects of Exhaust Gas Diesel” used a single cylinder engine, water cooled IDI diesel 
Recirculation on The Performance and Emission engine, connected to an electrical loading system. The fuel line 
Characteristics of a Diesel Engine Fueled with Waste Cooking was connected to the fuel tank. NOX, CO, CO2 and HC were 
Oil Methyl Ester”, used a single cylinder, four stroke, water measured using an AVL gas analyzer, while the air intake mass 
cooled direct injection, compression ignition engine, with a flow rate was measured by a circle edge orifice plate. After 
dynamometer coupled to an output shaft. Modifications were conducting the study, they observed that at an EGR of 15% of 
carried to allow the experiment to be carried out. In order to the mass flow rate, using the blends of JB20% at 25% EGR 
collect data a digital panel was installed, while an exhaust gas mass flow rate, JB40% at 15% EGR mass flow rate, JB60% at 
analyzer was used to investigate emission characteristics of the 20% EGR mass flow rate, JB80% at 40% EGR mass flow rate 
various gases in this study: CO, CO2, HC, NOX and O2. Finally, and JB100 at 5%EGR mass flow rate [8], the NOX emissions 
the fuel they used was BO (Normal diesel), B10 were seen to be reduced by 10.10%, 11.94%, 13.4%, 15.2%, 
(10%WCOME+90%) diesel by volume, B20 19.8% and 24.8% respectively. Soot emission, PM, CO, and 
(20%WCOME+80%) diesel by volume, B30 CO2 increased. 
(30%WCOME+70%) diesel by volume. The EGR % mass 
The NOX emissions were observed in all the biodiesel blends 
flow rate ranged from 0% to 20% graduated in intervals of 5%. 
at high levels, explained by the fact that biodiesel blends 
During this experiment all the tests were conducted at constant 
contain higher amounts of oxygen content; this trend is most 
1500 rpm, in order to study the effects of EGR on the emission 
visible at all full loads. Among the diesel blends the following 
and engine performance characteristics when waste cooking oil 
values were found 882 ppm, 848 ppm, 806 ppm, 775 ppm, 737 
is used together with diesel blends. The parameters investigated 
ppm respectively, whereas for diesel only the value was 643 
and their variations were: BSFC, BTE, EGT and emission 
ppm [7]. The BTE was observed to be lower for all the JB 
characteristics of NOX, HC and CO. 
blends compared to diesel due to the lower calorific value of 
During their study they observed that the BTE remained the biodiesel but the viscosity index was seen to be very high. 
unaffected by the application of an EGR% mass flow rate that During the study, diesel and the JB20 blend of fuel showed a 
is above 15%, but tended to decrease as the EGR% mass flow better maximum BTE value among all the blends used in the 
rate was increased beyond 15% EGR. This drop is explained study with 5% EGR mass flow rate and the JB90 blend of fuel 
by the fact that the availability of oxygen is reduced during showing a maximum reduction of NOX reduction percentage 
combustion with the application of EGR of 15%. Regarding the but with a decreased BTE of 1.25%. The decrease in BTE 
BSFC it was observed to be independent of the rate of EGR% during this study was attributed to the reduced availability of 
mass flow rate, particularly at lower rates of EGR% mass flow oxygen in the fresh charge which is an effect from the EGR 
rate, but it increased as the rate of EGR% mass flow rate application, where the exhaust gases result in lower flame 
increased above 15%. This phenomenon is explained by the temperatures and velocity, leading to an increase in soot and 
fact that there is formation of a rich mixture due to the dilution PM. However, in the JB60 test fuel and 20% EGR mass flow 
effect in the combustion chamber. This was identical to the rate, the NOX emissions reduced to 19.81% with only a slight 
findings of Charola et al. [5]. increase in the BTE of 0.77, attributed to the reburning of the 
unburned hydro-carbons together with the EGR% mass flow 
The effects of EGT that were observed if the engine was run on 
rate so helping to complete the combustion of the fuel. It is 
cooled EGR were that the in-cylinder temperatures were 
important to note that during the study the exhaust gas 
lowered, as compared to when the engine exhaust gases were 
temperature for all the study fuels (biodiesel and the blends) 
run without being cooled. The cooling of the EGR gases affects 
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© Research India Publications.  http://www.ripublication.com 
was found to be lower than when conventional diesel alone was (by volume) for the biodiesel all under full engine load 
used.  conditions. Thus, they concluded that as the EGR % mass flow 
rate increases so do the CO emissions, caused primarily by the 
Manieniyan and Sivaprakasam [1] in their study “Experimental 
lack of oxygen and poor mixing during the combustion process. 
Analysis of Exhaust Gas Recirculation on Diesel Engines 
Operating with Biodiesel” used a four stroke, water cooled, The smoke density was generally found to be lower or 
single cylinder direct injection DI engine, rated at 5.2 kW decreased for biodiesel as compared to diesel fuel and this was 
power with a 1500 rpm as the rated speed. The experiment was observed for all engine loads. As the EGR % mass flow rate is 
conducted in two phases. Phase 1 investigated the performance, increased there is an observed increase in the smoke density. 
combustion and emission of diesel fuel and biodiesel blends of This is explained by the fact that EGR affects the oxygen 
B20 where the EGR application was 0 (zero) or not applied. availability leading to incomplete combustion and the 
Phase 2 of the experiment used the same equipment and test rig formation of (PM. 
and fuel i.e. B20, but with the EGR application employed. The 
The EGT was observed to increase with the increase in the load, 
range of distribution for EGR% mass flow was 5%, 10%, and 
especially noticeable for biodiesel compared to diesels at all 
20%, while the load distribution was 20%, 40%, 60%, 80% and 
engine applicable loads and conditions during the study. This 
100%. The emission gases were measured using a gas analyzer 
increase in temperature was attributed to higher oxygen content 
and the density of the smoke was measured using a smoke 
in all biodiesel fuels particularly those used in the study. This 
meter. Finally, the exhaust gas temperatures were measured 
study finding was identical to Hawi, Kiplimo and Ndiritu [9]. 
using a thermocouple. Other combustion parameters like 
cylinder pressure, heat release and cycle to cycle variations The cylinder pressure at full engine loads and EGR% at 0 (zero) 
were measured using a combustion analyzer. was observed to be comparable to biodiesel, arising from the 
fact that there was a good mixing from the biodiesel mixture 
The tests and study found that the BSFC for the test fuel 
under high load, because EGR gases act like heat absorbing 
without EGR under engine full load was seen to be 0.2779 
agents due to the presence of CO2 and H2O caused by 
kg/KW-hr and 0.2794 kg/kW-hr for biodiesel. The value of the 
disassociation in the chemical effects of EGR application. Thus 
BSFC and the engine at full load and EGR% mass flow rate at 
the combustion analysis showing or revealing significant heat 
5%, 10%, 15% and 20%, were 0.2853 kg/kW-hr, 0.2796 
release rates, with or without the application of EGR% mass 
kg/kW-hr, 0.2832 kg/kW-hr and 0.3050 kg/kW-hr 
flow rate. 
respectively, whereas for the higher EGR% mass flow rate of 
20% the BSFC increased for both diesel and biodiesel. This Abu-Jrai et al. [10] in their study “Performance, Combustion 
was attributed to the high calorific values, higher viscosity and Emissions a Diesel Engine Operated with Reformed EGR 
density and boiling point.  Comparison of Diesel and Gas to Liquid (GTL) Fueling,” used 
a Lister-Pitter TR1 engine, single cylinder, direct injection, air 
The BTE was found to be comparable for diesel and biodiesel 
cooled, and naturally aspirated. It was connected to an electric 
with or without EGR. At full load the BTE returned a figure of 
dynamometer with a motor and load cell. The crank shaft 
29.006% for a test with EGR and 29.4933% without EGR 
position timing was set at 22° CA (BTDC) in all the tests. The 
application in full load. At 20% EGR BTE was reduced by 
EGR% mass flow rate was set manually using a gate valve 
1.5% in diesel and biodiesel and more exhaust gases were 
while EGR levels were determined by volume. The rig also 
produced, predominantly due to the dilution effect of EGR 
included other instruments to do measurements for oil and air 
resulting in a drop in efficiency. 
inlet/exhaust manifold temperature respectively. 
The variations between the hydrocarbons and the brake power 
Data was acquired using software for measurement of: peak 
results and conclusions indicate that with an increase in EGR 
cylinder pressure, IMEP (indicated mean effective pressure), 
mass flow rate levels, the HC emissions also increased only for 
COV (percentage coefficient variation) of IMEP, average 
the biodiesel, attributed to the presence of high oxygen content 
values of COV of peak cylinder pressures, ROHR (rate of heat 
in most biodiesels, thus compensating for the lack of oxygen to 
release). The AVL gas analyzer was used in the measurement 
facilitate complete combustion thus the increase in BTE. The 
of NOX, CO and CO2 gases. The test fuels used for the test were 
values for the NOX during the study were found to be 736 ppm 
ULSD (ultra-low sulphur diesel) and the ultra-clean GTL (gas 
for diesel fuel running without EGR in full load environment 
to liquid) which was obtained through a LTFT (low 
and 796 ppm for the biodiesel at full load without EGR. 
temperature Fischer-Trosph) process, which is a synthetic fuel.  
However, with the application of EGR it was noticed that the 
levels of NOX reduced to 157 ppm for diesel and 158 ppm for The experiment was conducted at two engine speeds of 1200 
the biodiesel respectively, but a reduction in BTE was rpm and 1500 rpm and two loads only tested at 25% load and 
established with large increments in smoke density. 50% (medium load), while the EGR ratios chosen were 
classified into two; standard EGR and REGR (reformed 
Carbon monoxide (CO) in relation to diesel and biodiesel with 
exhaust recirculation) both using ratios of 0%, 10%, 20%, and 
various levels of EGR% mass flow rate and load condition was 
30%. The REGR was a simulation based on H2 gas and CO gas 
noted to vary from 0.14% (by volume) for the diesel and 0.18% 
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in the ratios of 25% = H2(gas) - 0% CO(gas)-75%EGR while engine increased as the engine speed increased in both the 
in the second scenario 15% = H2(gas)-10%CO(gas)-75% EGR.  phases of the experiment. i.e. when the inlet intake air 
preheating was applied without EGR application and vice 
The conclusions from this study were as follows: at low engine 
versa, a factor they attributed to the decrease in the duration of 
loads the GTL fuel was found to bring or return high BTE 
ignition delay and combustion chamber premixing time of the 
values as compared to the BTE efficiency values the ULSD 
inlet intake charge. In addition, they observed that the 
returned, but when REGR was introduced it was observed that 
application of EGR reduced the in-cylinder combustion peak 
GTL had a higher fuel replacement than ULSD, thus showing 
temperature due to the higher CO2 heat specific ratio. This 
an improvement in efficiency and reduced emissions. BTE was 
result is identical to results in other studies, particularly those 
also found to be load dependent, i.e. at low engine loads 
where the EGR% rate application was limited to between 20% 
efficiency decreased with increasing REGR [10]. However, 
to 25%, this being identical to the study done by [13]. 
when the loads were high it was observed that the high flame 
temperatures and velocity of hydrogen gas and the rapid However, on the relation between the effects of EGR on inlet 
expansion increased the BTE as the REGR was increased. intake air preheating and NOX emissions, they observed that 
with an increasing in the inlet intake air temperature, the 
On rate of heat release, the high octane number of GTL resulted 
emissions of NOX and CO indicated a decrease. This factor was 
in less pronounced premixed combustion, leading to lower heat 
linked to the ignition delay and premixed combustion mixture, 
release rate and in-cylinder pressure, leading to reduction in 
two factors whose time is shortened by the introduction of 
NOX emission where GTL fuels were employed, unlike when 
EGR. There are other identical studies whose findings are 
ULSD fuel was used in the experiment. Whereas the 
consistent with this study [14]. 
combustion pattern was affected by the REGR, the use of the 
30% REGR brought efficient gas fuel combustibility and an Lastly, the authors wanted to show the relationship between the 
increase in-cylinder pressure together with the ROHR [10]. At effects of adiabatic flame temperature on the emission of NOX 
low engine loads the authors noted that an increase of REGR particularly when conventional diesel fuel is applied or used. 
resulted in a decrease of both NOX and smoke emissions. When conventional diesel was used the authors observed that 
However, at higher engine loads higher REGR resulted in a as the equivalence ratio increased (hitting a maximum ratio of 
decrease in smoke emission but compromised NOX emission. 0.9), the adiabatic flame temperature also increased. However, 
the equivalence ratio began to decrease as it hit the maximum 
Lastly, when the combination was used of both REGR and 
value, though at this point the NOX emission continued to 
GTL fuel, the results showed a marked improvement in the 
increase and vice versa. 
reduction of emissions The 30% REGR when combined with 
the GTL fuel produced a 60% decrease in smoke emission, Regarding CO emissions, the application of inlet intake air 
while the NOX emission showed a 75% decrease as compared charge preheating at a controlled temperature of 55°C and 
to when ULSD fuel was used at 0% REGR and low engine without application of EGR% brought about a lowering of CO 
loads [10].  emissions to almost a minimum as explained by the fact that 
inlet intake air preheating provides CO oxidation. However, 
Shahadat, Nabi, Akhter and Tushar [11] in their study on 
when the EGR mass flow rate of 25% was applied it was 
“Combined Effect of EGR and Inlet Air Pre-Heating on Engine 
observed that the CO emissions increased greatly. This concurs 
Performance in Diesel Engines” used a four stroke diesel 
with the findings of a study conducted by Fathi, Saray and 
engine S195, single engine, with a maximum speed of 2000 
Checkel [15], where it was concluded that the application of 
rpm, direct injection, water cooled and naturally aspirated. The 
EGR does not have similar effects on the different species of 
fuel used for this study was conventional diesel, with a density 
emission characteristics. 
of 0.80g/cc@25°C. Other modifications to suit the experiment 
were made to provide measurement readings, particularly the The BTE during this study was found to be affected by the inlet 
preheating mechanism for inlet intake air. To maximize the intake air preheating in combination with the application of 
heat loss, the inlet intake air passage was insulated by plaster. 25% EGR mass flow rate but at varying engine speeds and 
constant load. The authors found that the BTE was higher for 
The overall heat transfer between the two exchange mediums 
the preheated inlet intake air than when compared to when 
(the inlet intake air and the exhaust gas) was calculated using 
preheating is not applied. This was seen to be due to higher 
the equation from [12]. The experiment was conducted in two 
amounts of heat energy that were being transferred from the 
phases or two parts; Phase 1 was conducted without the inlet 
exhaust gas to the inlet intake air, thus providing better burning 
air preheating and without EGR% mass flow rate application, 
of the fuel in the combustion chamber, hence minimizing the 
while Phase 2 involved preheating the inlet intake air charge 
overall heat loss by the system. In addition, they also observed 
and EGR% application. The gas emission measurements were 
that the BTE increased with the increase in engine speed both 
made using a gas analyzer to obtain values for comparative 
at low and intermediate speeds, but at higher speeds they noted 
analysis. 
a decrease in the BTE, a factor which they attributed to shorter 
After running the experiment, the following results and mixing time of the air fuel that is normally associated with 
conclusions were observed. The emission of NOX from the test higher engine speeds leading to poor or low fuel efficiency. 
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Abdelaal and Hegab [16] in “Combustion and Emission Applying EGR they noticed affected the peak in-cylinder 
Characteristics of a Natural Gas Fueled Diesel Engine with pressure by reducing it. This is explained by the fact that the 
EGR,” used a Ricardo E-6 engine modified to accommodate recycled exhaust gases in the form of EGR act as a thermal sink 
gaseous fuel and normal or conventional diesel use. The engine as earlier mentioned above, thus reducing the charge 
was single cylinder, indirect injection, four stroke, water temperature during combustion. The pressure rise during the 
cooled, with a maximum rated power of 9 kW, and a maximum combustion becomes reduced which is a means of avoiding 
rated speed of 3000 rpm. Modifications were made to testing engine knock. 
and gathering of data. Following testing and experimentation it 
Heat transfer was noticed to decrease with the application of 
was observed that the introduction of CO2, through the intake 
EGR, since the heat transfer coefficient is proportional to 
manifold resulted in a decrease in NOX emissions, but showed 
pressure and inversely proportional to temperature [15]. 
a relative increase in unburned hydrocarbons. 
Therefore, the heat transfer coefficient did not vary much with 
The disassociation of CO2 and its introduction into the intake the application of EGR, since by increasing EGR% mass flow 
charge is attributed to the reduction in NOX. This is a chemical rate both the pressure and temperature decreased. 
effect that is associated with the use of the EGR application. 
The gross heat release rate (HRR) was seen to be affected by 
Another observation the authors’ made was that by continuing 
the application of EGR flow rate because recycled exhaust 
to increase the inlet charge temperature there was an increase 
gases replace or substitute part of the fresh air intake charge, 
in the unburned hydrocarbons emissions mainly due to a 
thus affecting the in-cylinder equivalence ratio and the 
decrease in ignition delay in relation to the rise in the intake 
combustion process at large. EGR rate application can cause 
charge temperature. Increasing the intake charge temperature 
delay in SOC (start of combustion) hence prolonging the 
resulted in improved combustion characteristics due mainly to 
burning or exposing the elements of combustion to more time 
the drop in the ignition delay, thus improving and increasing 
to combust, which in return leads to increase in the IMEP limit 
the BTE and BHP (brake horse power) power parameters and 
and the operating range of the EGR rate. 
a decreasing the unburned HC emissions and carbon monoxide 
(CO). It is imperative to note at this point that the increase in The effects of EGR on combustion timing demonstrated that 
intake charge temperature during this experiment resulted in an since EGR increases specific heat ratio of the combustion 
increase in the NOX emission i.e. dual fuels reduce NOX elements, the in-cylinder inlet charge temperature decreases 
emissions during low engine loads only. The admission of thus delaying ignition as mentioned in the discussion above, 
diluents was shown to produce poor combustion characteristics thereby introducing retardation of SOC and increasing the 
if they were increased, thereby causing reduction in the NOX combustion duration. Defined as the indicated work output per 
due to the lowering of the cyclic cycle temperature and the unit of engine swept volume, the IMEP was seen to be less 
oxygen concentration due to the dilution of the intake charge affected by the application of the EGR rate which is identical 
by introduction of the EGR gases. with other studies conducted in this area [17]. 
Fathi et al. [15] in “The Influence of Exhaust Gas Recirculation Regarding the indicated specific fuel consumption (ISFC) the 
on Combustion and Emissions of N-Heptane and Natural Gas authors found that the retardation of the start of combustion 
Fuelled Homogeneous Charge Compression Ignition (HCCI),” (SOC) reduced ISFC, whereas when the EGR rate was applied 
conducted a study at the University of Alberta fuel and it produced the opposite effect i.e. increased the ISFC. This is 
combustion emission research facility. They used a Waukesha explained by the fact that introduction of EGR rate removed the 
CFR single cylinder research engine, water cooled, coupled to availability of oxygen thus causing incomplete combustion in 
a dynamometer. The speed was maintained at 800 rpm at wide the combustion process thereby decreasing efficiency. 
throttle. A temperature controlled air preheating system was 
The formation of NOX emissions as observed during this study 
installed using a 2.4 kW heater motor. Also included were two 
with EGR application, was due to the nature of the combustion 
port fuel injectors controlled electronically via an AFS sparrow 
in homogeneous charge compression ignition (HCCI) engines 
module. The EGR rate was manually controlled. The fuels used 
which show a fast combustion. The formation of NOX is limited 
for this experiment were n-heptane and natural gas either as 
due to the low charge temperature of the air mass flow rate. 
blends or used individually. 
This is explained by two factors: applying EGR rate flow 
During their study the authors observed that the recycled gases depresses peak in-cylinder temperature thus reducing NOX 
from the exhaust contained tri-atomic molecules of H2O and emissions formation, and secondly the application of EGR rate 
CO2 species which increase the specific heat capacity of the depresses the availability of O2 concentration in the 
combustion mixture and reduce the overall specific heat ratio. combustion chamber due to the EGR rate dilution effect. 
This makes the mixture act like a thermal heat absorbing sink, However, an increase of EGR rate was noticed to increase HC 
thereby decreasing the after-compression temperature. The and CO emissions as the EGR rate continued to increase. The 
other observation they made with EGR application is a explanation provided by the authors was the unavailability of 
reduction in peak in-cylinder charge temperature due to O2 and reduced inlet intake charge temperature which resulted 
delayed auto ignition of the combustion charge. in incomplete combustion due to poor mixing. 
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Hawi et al. [9] in the “Effects of EGR on Performance and Finally, on the CO emission and how they interacted with the 
Emission Characteristics of a Diesel Piloted Biogas Engine” EGR rate, the authors noted an increase of the CO emissions 
used a single cylinder, four stroke engine, water cooled, with an increase in the engine load. However, when a n EGR% 
naturally aspirated constant speed CI engine, coupled to a rate of 10% to 30% was applied, there was a minimal decrease 
hydraulic dynamometer for load application. Modifications associated with the substitution of fresh charge with the 
were made to the engine to enable it to run on dual fuel with elements of the exhaust gas which decreased the CO2 
conventional diesel as a pilot fuel and biogas as the main fuel. emissions. 
A gas analyzer was used, with a digital portable laser 
Hountalas, Mavropoulos and Binder [19] in the “Effect of 
tachometer for engine speed. Thermocouples were used to do 
Exhaust Gas Recirculation (EGR) Temperature for Various 
temperature readings. 
EGR Rates on Heavy Duty DI Diesel Engine Performance and 
After conducting tests and the experiment they came out with Emissions.” conducted their study using a heavy duty DI diesel, 
the following findings on BSFC. Through the variation of EGR single cylinder engine, with high compression peak pressure. 
rates from 0% to 100% of engine load, the BSFC was found to To improve the BSFC advanced timing was employed, while 
decrease as the rate of EGR% rate was increased, obtaining a for NOX emissions control the EGR was constantly cooled 
maximum value when the EGR rate was 20%. The decrease in throughout the experiment. The test engine condition was 
the BSFC with the application of the EGR rate was due to an maintained at full load with variations of engine speed 
increase in the inlet intake charge temperature, thus increasing employed, with the theoretical study on 3D multi zone model 
the rate of fuel consumption. modified to suit the experimental set up under study to include 
the effects of EGR% and temperature and for calculation 
On the brake power (BP) they observed that the application of 
analysis. The boost pressure and injection pressure were kept 
the EGR did not produce any significant effect on the total 
constant at 280 BAR and 140°C for the EGR gas. 
output power of the engine with the application rate of 30% 
EGR, while the BTE was observed to increase with the increase Figure 2 shows the comparison between the calculated and 
in the engine load for all operating conditions. However, it experimental cylinder pressure for EGR% rates at 1130 rpm, 
should be mentioned that the BTE showed some insignificant 100% load and -9 ATDC injection timing [19]. 
increase before the EGR rate passed the 20% mark, attributed 
to the returning of the hydrocarbons entering the combustion 
chamber for re-combustion together with EGR, and the 
increase of inlet intake charge temperature as a result of EGR 
rate application, which causes an increase in the rate of 
combustion. This is identical to the findings of Venkateswarlu, 
Murthy and Subbarao [18]. 
Regarding EGT, the authors concluded that it increases with 
engine load for all the operating conditions, which they 
attributed to the increase in input energy at high engine loads, 
as fuel consumption increases to accommodate high load and 
speed. Another observation they made during this study was 
that EGT temperature reduced with the increase of EGR% rate 
flow, which they attributed to the unavailability of O2 when 
EGR is applied, leading to lowered combustion temperatures. 
This is identical to other studies conducted [19].   
CO emissions variations in relation to the increase in EGR% Figure 2: Showing the calculated and experimental cylinder 
rate was noted as increasing with the increase in the percentage pressure in relation to different EGR% rates. 
of EGR, but was seen to significantly reduce as the EGR and 
engine load increased to 75% and 100%. This factor is Source: Hountalas, Mavropoulos and Binder [19] 
explained by the lack of O2 during combustion due to the  
dilution effect of EGR. On the other hand, HC emissions are 
As can be seen from Figure 2, simulated and theoretical 
noticed to increase in all the engine loads conditions as the 
measured values results show similarities and coincidences for 
EGR rate is increased, explained by the fact that the induced 
the values of the EGR rate examined [19]. The study indicates 
charge (intake air + EGR rate) has a higher CO2 content when 
a reduction in-cylinder pressure during combustion and 
compared to fresh air only. However, the increased load across 
expansion of the mixture in the combustion chamber. This 
the entire engine load shows a decrease in HC emissions 
increase in the SHC as occasioned by the presence of exhaust 
particularly when the dual fuel is employed for operation. This 
gas recirculation reduced the availability of O  thereby 
study finding was identical to those of Mahla, Das and Babu, 2
affecting the combustion process rate and the disassociation of 
[20]. 
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CO2 and H2O, further reducing the peak cylinder pressure The study was conducted using a four-cylinder diesel engine 
values as EGR% rate mass flow increased. During part engine with modified EGR application, naturally aspirated, air cooled, 
load different results were obtained, as the availability of O2 and with a maximum power of 64.9 kW @ 4500 rpm. The 
was higher compared to when the engine was operated at full engine was coupled to an eddy current ECB dynamometer that 
load. As the EGR inlet intake temperature increases with the was controlled by a Dynaelec load controller.  
EGR% rate mass flow this leads to delay in SOC leading to 
The result on the effect of EGR on the performance and 
increased peak cylinder pressure as EGR % rate mass flow is 
combustion characteristics of the palm biodiesel and mineral 
increasing. These results are identical to those of [9], [19], [21]. 
diesel under normal modes, EGR application and at full load 
The results indicated a negative effect of the EGR% rate mass condition, was observed. The test fuel, which in this case was 
flow on the BTE, especially at low engine speeds and full load, palm biodiesel, produced higher values of BSFC when the 
where an increase in EGR% rate mass flow showed a exhaust gas temperature was low compared to when the EGT 
significant lineal decrease in BTE % of 5.5% @ 15% EGR. In gas temperature was high. This increase in BSFC was 
relation to emissions of NOX and soot emissions, the study attributed to the power loss or drop in the efficiency of the 
found that the theoretical and measured values of NOX and soot engine and the lower heat release rate values that are associated 
at full engine load and varying EGR% rates showed that as the with palm biodiesel, in addition to the unavailability of O2 in 
rate of EGR mass flow increases it reduces the NOX emissions the inlet intake air mass flow with the application of EGR [23], 
linearly but increases the soot emissions significantly and which mixes the inlet intake air with the EGR gases. The study 
exponentially [19], with the effects becoming more critical at further showed palm biodiesel increased the exhaust gas 
low engine speeds. temperatures by 5.6% above the exhaust temperature values 
that were produced when mineral diesel was used [23]. This is 
The effects of EGR rate flow temperature on the engine 
attributed to the presence of greater O2 content in the palm 
performance and emission was investigated by Moiz [22]. The 
biodiesel leading to a high combustion efficiency and high in-
tests done for each case were 5% EGR, 10% EGR, and 15% 
cylinder temperature, being identical to the findings of 
EGR and the temperatures examined were from 90°C to 240°C 
Badruddin et al. [24]. The experimental engine set up test rig is 
calibrated in steps of 50°C. However, it is imperative to note 
shown in Figure 3. 
here that this temperature range is a representative range value 
because the cooling media is the engine coolant. As the test 
results show it is evident that the increase in EGR% rate of 
mass flow at constant boost pressure led to a decrease in the 
intake air mass flow rate sucked or induced per cycle, leading 
to a drop in the AFR (volumetric efficiency). This type of result 
should be expected especially when the EGR gas temperatures 
are increased with the increase in EGR%, thus forming a lineal 
relationship between the AFR and EGR gas temperature. In all 
the examined test cases the authors observed that any increase 
on the EGR% rate mass flow temperature resulted in a decrease 
in AFR, but was more pronounced during low engine speeds 
i.e. at full loads thermal throttling becomes significant and 
increases with increase in EGR% rate mass flow gas 
temperature [19]. 
The effects of high EGR% gas temperature on the peak 
combustion pressure was noted by the authors as a reduction 
on the peak combustion pressure, which was noted to be 
significant at very high EGR% rates where one should note that 
the AFR was minimum. However, it should be emphasized that  
the use of the high EGR% rate gas temperature reduces O2 Figure 3: Showing the experimental set-up rig. 
availability thereby reducing the combustion rate or SOC. It is 
Source: Yasin et al. [23] 
also important to remember or mention that the effect on 
ignition delay or retardation is negligible due to the high boost  
pressure and high charge temperature. 
The application of EGR% mass flow rate mode on the palm 
Yasin et al. [23] in their “Study of a Diesel Engine Performance biodiesel and mineral diesel fuels during the study lead to a 
with Exhaust Gas Recirculation (EGR) System Fuelled with noted decrease in the exhaust gas temperatures of 2.8% when 
Palm Biodiesel” intended to study the effects of EGR on the palm biodiesel was used and a 1.6% decrease when mineral 
parameters of SFC, EGT, and the emissions of NOX, CO, CO2 diesel was used [23]. This finding is similar to Saravanan [25] 
and UHC both in experimental and simulated study situations. and therefore confirms that most biodiesel fuels produce lower 
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EGT temperature values at varied engine speeds and constant Indus gas analyzer. All tests were conducted under full engine 
engine loads.This conclusion is attributed to the fact that the load conditions and speed. The result from this study indicate 
exhaust gases as mentioned earlier mix with the inlet intake that the BSFC is shown to be slightly independent at lower 
fresh charge which dilutes the methyl ester rich oxygen content EGR rates, but increases as the EGR% rates increase to 15%. 
in the combustion, thereby improving combustion completion. This can be explained from the fact that less oxygen means a 
rich mixture formation in the combustion chamber. 
The palm biodiesel and mineral diesel emission characteristics 
compared to normal modes under EGR application and full The results showed that BTE remains equally unaffected when 
engine load conditions at constant speed of 2500 rpm showed low EGR rates and full engine loads were applied. However, 
an increase in NOX emissions, a factor that is linked to the BTE decreased with an increase in EGR above 15% in the 
increase in the in-cylinder temperature and the presence of combustion chamber, due to the displacement and replacement 
more O2 content within the natural chemical composition of the of the fresh air charge by the EGT gases. The EGT partly 
palm biodiesel fuel used during the study. The palm biodiesel operated with cooled EGR% mass flow rate was observed to 
thus seemed to contribute to an increase in the emissions of lower the exhaust gas temperatures compared to EGT at normal 
NOX compared to the mineral diesel, because of the factor engine operating conditions without EGR. i.e. EGT decreased 
already mentioned regarding the oxygen content which causes with an increase in EGR rate. This is informed by two factors, 
high oxidation of the combustion mixture which leads to highly namely, unavailability of oxygen and the high specific heat 
flammable rapid combustion thus increasing the exhaust gas ratio of the intake air mixture [26]. It is imperative to note here 
temperatures. that a decrease in exhaust gas temperature is noticed 
irrespective of the continuous increase in EGR% rate even 
Although the results obtained showed that palm biodiesel 
above the value of 15% EGR. 
produced 4.7% higher NOX emissions compared to the mineral 
diesel with EGR mass flow rate under normal engine Regarding emission characteristics, the authors noticed a 
conditions and loads, it should be noted that low NOX emission significant decrease in NOX for all the test fuels used due to a 
of 5.4% was achieved when mineral diesel was used as reduction in O2 and a decrease in flame temperature within the 
compared to 22% when palm biodiesel was used. This is combustion chamber [26]. The hydrocarbon emissions showed 
explained by the fact that an increase in total heat capacity an increase with the application of EGR% rate for all the test 
(specific heat capacity) of the exhaust gases and reduction in fuels used, especially when the EGR% was above 15%. This is 
in-cylinder peak temperature tends to reduce the NOX explained by the incomplete combustion occasioned by the 
emissions [23], these results being identical to the findings of unavailability of O2 resulting in poor oxidation of the 
Abdelaal and Hegab [16].  combustion mixture causing incomplete combustion. The CO 
emissions were noted to increase with increase in EGR% rate 
The carbon emissions for the test fuels used with application of 
for all the fuels used during the experimental study, mainly due 
EGR showed a decrease in carbon emissions for the palm 
to the aforementioned O2 phenomenon. 
biodiesel compared to mineral diesel. This is explained by the 
complete burning of the mixture as the palm biodiesel contains Asad and Zheng [27] in their study of “Exhaust Gas 
more O2 content compared to mineral diesel, which provides Recirculation for Advanced Diesel Combustion Cycles” Used 
the oxidation of CO to CO2 which is an inert gas [23]. However, a four-cylinder Ford common rail diesel engine coupled to an 
from the result obtained it is evident that there is an increase in eddy current dynamometer. The engine was operated in two 
CO emission of 9.7% when palm biodiesel was employed modes. The first mode was a single cylinder mode (3 cyl to 1 
compared to a 2.45% increase using mineral diesel especially cyl) where the 3 cylinders are operated in a conventional mode, 
with the application of EGR, at full engine load i.e. both the with the #1 cylinder designated for research with independent 
fuels showed an affinity to produce CO emissions when EGR controls of the parameters of intake pressure, exhaust back 
is applied under all engine operating conditions and load. pressure and fuel injection. The second mode involved all the 
four cylinders with the engine running using a turbo charger, 
Rao [26] in the “Effect of EGR on Diesel Engine Performance 
air fuel and EGR with no major system alterations. The exhaust 
and Exhaust Emission Running with Cotton Seed Biodiesel.” 
gas analyzer was used to measure NOX, THC, CO, CO2 and O2 
used a single engine, four stroke, water cooled engine with 
concentration, while the intake CO2 and O2 were measured in 
cotton seed biodiesel (CSBD) and PBD (petroleum based 
the intake manifold using a different intake analyzer. The soot 
diesel) blends as test fuels, where 0% PBD = CSBD10, 
concentration was measured using an AVL smoke meter, while 
20%PBD = CSBD20 and 30%PBD = CSBD30. The EGR% 
all the cylinder pressure data was processed with a 0.1° angle 
rate of mass flow was limited to 0% to 20% in steps of 5%. The 
resolution using a pressure transducer mounted on cylinder #1 
engine speed was kept constant at 1500 rpm. The set up was to 
and connected to a data acquisition system.  
examine the following parameters in relation to the effects of 
EGR: BSFC, BTE, EGT and the emission characteristics of After testing the models and how they interacted with the 
NOX, HC and CO. The engine was modified to allow EGR% application, the following results and conclusions were 
experimental work to be done and for data to be obtained from reached. The impact of intake and in-cylinder charge 
the digital control panel. The gases were measured using an composition was characterized with in-cylinder excess air ratio 
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for the recycled O2 with the EGR as a function of the EGR this area and this review will present some of these studies and 
displacement [27]. The effects of EGR on intake pressure and their current position in so far as this area is concerned. 
engine load showed that the dilution of the intake charge with 
Mani, Nagarajan and Sampath [30] in “An Experimental 
application of EGR could correctly be predicted through the 
Investigation on a DI Diesel Engine Using Waste Plastic Oil 
intake oxygen concentration as compared to EGR fraction, as 
with Exhaust Gas Recirculation” showed that plastic oil 
it was shown to lack additional information to show the 
produced the most NOX emissions but showed reduced NOX 
effectiveness of the EGR ratio in relation to displacement [28]. 
emissions with increasing EGR of 20% onwards. There was no 
It is imperative to note at this point about the EGR% ratio: there 
noticeable change in the smoke, CO and HC levels of 
are two methods which Asad and Zheng, [27] propose in this 
emissions, thus reaching a conclusion that with the application 
study which are presently widely used to measure and quantify 
of EGR% rate levels plastic oil can be used as an alternative to 
the value of EGR rate, the gas concentration based method and 
biodiesel fuel.  
the mass based method. So far there has not been a standard 
method of defining EGR rate. The two equations for defining Xu, Gui and Deng [31] in “Fuel injection and EGR control 
the EGR% rate flow as proposed by [27] are: strategy on smooth switching of CI/HCCI mode in a diesel 
engine” conducted an experiment by variation of the cam phase 
[𝐶𝑂2]𝑖𝑛𝑡− [𝐶𝑂2]𝑎𝑚𝑏 [𝐶𝑂2]𝑖𝑛𝑡
 𝑟𝐸𝐺𝑅 ≈  ≈  equation 1 [𝐶𝑂 ]𝑒𝑥ℎ−[𝐶𝑂 ]𝑎𝑚𝑏 [𝐶𝑂 ]𝑒𝑥ℎ and tapped EGR gases as a heating or vaporizing mechanism 2 2 2
of the fuel and to harmonize the mixture charge, while using 
ṁ
 𝑟 = 𝐸𝐺𝑅𝐸𝐺𝑅    equation 2 ṁ +ṁ +ṁ the fuel injection to administer HCCI in an electronically 𝑎𝑖𝑟 𝑓 𝐸𝐺𝑅
controlled fuel injector. They used HCCI combustion 
But since the fuel flow measurement is less than the mass air modelling to study the effects of a premixed ratio in relation to 
flow [27], equation 2 is modified to define the EGR ratio rate performance and emission characteristics, when different EGR 
as shown in equation 3: % rates are applied. They observed a 78% drop in NOX and 
ṁ𝐸𝐺𝑅 ṁ𝑎𝑖𝑟 lowered smoke and soot emissions of 40%, when premixed 𝑟𝐸𝐺𝑅 = = 1 −       equation 3 ṁ𝑖𝑛𝑡 ṁ𝑖𝑛𝑡 ratio of 80% and 30% EGR respectively were applied. 
where 𝑟𝐸𝐺𝑅 is the EGR fraction mass rate of the recirculated Increased PR lead to increase in IMEP and BSFC, but fared 
gases, ṁ𝐸𝐺𝑅  is the flow rate of the mass of EGR gas, ṁ𝑎𝑖𝑟is the 
poorly on the HC and CO emissions.  
gas flow rate of intake air mass, ṁ𝑓is the flow rate of the Haber and Wang [32] in “A Robust Control Approach on 
injected fuel mass, [𝐶𝑂2]𝑎𝑚𝑏 is the carbon dioxide gas in the Diesel Engines with Dual-Loop Exhaust Gas Recirculation 
ambient air, [𝐶𝑂2]𝑖𝑛𝑡 is the carbon dioxide gas in the inlet Systems,” discuss a number of control strategies to be 
intake air, 𝐶𝑂2]𝑒𝑥ℎ is the carbon dioxide gas in the exhaust employed in mitigating the effects of EGR and emission. One 
system. of the methods suggested is EGR. The authors point out that 
the application of EGR alters or affects the normal combustion 
During low engine loads and low EGR it was noticed that the 
process in several ways. Firstly, it alters the physical properties 
effectiveness of EGR decreased as EGR gases displaced small 
of the charge mixture; secondly it alters the air fuel ratio (A/F); 
amounts of fresh air in the intake system due to the fact that 
and thirdly it alters the start of ignition and finally alters the 
EGR gases contain O2 and N2 elements and vice versa i.e. 
heat release pattern [32]. Other effects they noted as 
increasing the intake reduced the intake charge dilution for the 
responsible for NOX reduction were the dilution effect, the 
same EGR ratio [27], while increase in EGR (intake dilution) 
chemical effect, and the thermal effect. They also noted that the 
produced a leaner air fuel ratio with increased premixed 
application of EGR resulted in modification of both the 
combustion, which caused reduced soot emission, which is 
composition and physical conditions of the inlet charge (by 
identical to the findings of Aoyagi, Osada, Misawa, Goto, & 
modification of composition they mean addition of CO2 and 
Ishii [29] and improved on combustion efficiency [28]. It 
H2O, and by modification of physical conditions they mean 
should be emphasized here that O2 in the EGR is both a 
temperature density) [32]. By replacing the intake air charge 
function of the engine load and EGR ratio.  
using EGR they observed that it decreases the air fuel ratio 
In mentioning and discussing the combustion models, it should leading to a shortened delay in ignition. Other effects they 
suffice to say that the development of alternative fuels becomes noted in their study is that reduced oxygen availability due to 
of paramount importance. There are several methods that have the admission of EGR also explains the phenomenon of 
been proposed for processing the alternative fuels otherwise delayed ignition, which they observed to be due to increased 
called biodiesels, the commonly used ones being mixing time between the amount of injected fuel and fresh 
transesterification, hydrogenation, pyrolysis (used in plastics) oxygen [32]. In conclusion they argue that as a control strategy 
and dilution. The use of alternative fuel for emission control EGR cannot be used alone because of the increased soot 
has been increasing, particularly with the use of the EGR emissions, fuel penalties, degradation of the combustion 
control technique to control and reduce NOX, HC, CO and soot quality, increased piston cylinder wearing, deteriorating 
emissions. A number of authors have carried out research in efficiency, operational instability, and higher EGR temperature 
accounting for low fuel efficiency [32].  
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Lehto, Elonheimo, Hakkinen, Sarjovaara and Larmi [33] in levels it was observed that the values of λ was recorded was λ 
their study “Emission Reduction using Hydro Treated ≤ 3.0 thereby producing knock and a very rough combustion 
Vegetable Oil (HVO) with Miller Timing and EGR in Diesel experience. 
Combustion.” noted a marked reduction in the emissions of 
 
NOX and smoke compared to conventional diesel. EN590, a 
HVO fuel blend, can allow the use of higher EGR% rates CONCLUSION 
without compromising the PM, HC and smoke emissions but 
It can be safely concluded that the use of the EGR as a 
with exponential reduction in the NOX emissions.  
technique for controlling emissions of NOX, CO, CO2, HC and 
Chen, Liu, Wu and Lee [34] in “Effects of Port Fuel Injection SOF in conjunction with other techniques discussed in this 
(PFI) Of N-Butanol and EGR On Combustion and Emissions review when implemented will adequately address the 
of a Direct Injection Diesel Engine,” studied the effects of PFI underlying issues of the effects of EGR on diesel engine 
on n-butanol and application of EGR and how they influence performance parameters and reduction in emissions as was 
emission and combustion characteristics in a diesel engine. In seen during the studies that were reviewed here.  
their study the engine was equipped with independent inlet and 
However, more needs to be done to bridge the gap and 
outlet systems. The researchers varied the EGR% rate which 
disconnect between research and industry. For example, EGR 
they divided into three sections 0%, 15% and 45%, and the 
cooling is not widely used in stationery engine systems and in 
premixed fuel of the n-butanol was divided into four premixed 
marine applications and in industrial production where carbon 
ratios of 0%, 20%, 38% and 47% by volume. 
emissions are more pronounced, but is relegated and left to 
They observed and noted that as the rate of EGR% increased to industry handlers and policy makers to decide, something they 
45% there was a significant decrease of 97% of the NOX might not be able to do.  
emissions though there was an exponential increase in the soot 
EGR cooling will allow us to retain the benefits of low NOX 
emissions and the SOFs. However, when they combined the 
emissions but without compromising engine efficiency. EGR 
use of the PFI and the n-butanol of 47% while maintaining the 
cooling is necessary as it reduces soot and soluble organic 
EGR% rate at 45%, the emissions of NOx and those of soot 
fraction (SOF) emissions, for example when used with other 
were drastically reduced by 88% and 17% respectively, thereby 
techniques like VGT technology, which proposes to recirculate 
addressing the traditional basic fundamental problem of the 
all exhaust gases together with the intake mixture thus 
NOX – soot trade-off. 
increasing the mixture heat capacity but leading to lowered 
Zhao, Wang, Li, Lei and Liu [35] in their study on combustion temperature and consequently lower NOX 
“Combustion and Emission Characteristics of a DME emissions; due to O2 concentration levels being maintained, 
(Dimethyl Ether)-Diesel Dual Fuel Premixed Charge oxidation of the soot emissions becomes insignificant.  
Compression Ignition Engine with EGR (Exhaust Gas 
During the review it was observed and noted, based on the 
Recirculation)” studied the effects of external EGR and di- 
studies that have been done, that the preheating of the inlet air 
methyl ether (DME) on the premixed ratio on the performance 
reduced CO emissions by the introduction of EGR into the 
and emission characteristics of a two cylinder naturally 
system, though it is observed that CO emissions increased 
aspirated diesel engine. They varied EGR% from 0% to 27% 
linearly as the rate of EGR% was being increased; which raises 
while they kept the premixed ratio at between 0% and 30%.  
the question of how the two systems can be harmonized to 
Their results indicated that an increase in EGR% mass flow achieve low carbon emissions without compromising engine 
reduced NOX emissions, but increased CO2 and HC emissions performance and emission control systems. From the 
considerably. However, a very good result was obtained with combustion control strategies discussed, the question arises 
higher PR% and higher EGR% rate especially in NOX and regarding what implementation strategies can be employed to 
smoke emissions.  improve combustion characteristics of the various fuels 
available, particularly the biodiesel fuels that have been 
Singh and Agarwal [8] in “Combustion Characteristics of 
developed, especially their chemical and physical properties, 
Diesel HCCI Engine: An Experimental Investigation Using 
which sometimes lack proper testing and analysis facilities. 
External Mixture Formation Technique,” used a twin piston DI 
Biodiesel is still a grey area in terms of industrial commercial 
diesel engine to conduct a study by the modification of one 
development and industry regulators who impose stringent 
cylinder to use HCCI premixed fuel and the other one using 
control on its use 
conventional diesel. During the study they varied their mixture 
concentration from λ = 4.96 (as the lean mixture level) or the  
misfiring limit and λ = 2.56 (considered as the rich mixture 
ACKNOWLEDGEMENTS 
level) or the knock limit. The EGR applied was at the rate 20%. 
The results observed indicated an increased combustion First and foremost, to the good Lord for His enabling mercies, 
efficiency, when the value of λ ≤ 3.70 which is medium range to my supervisor Prof. F. N. Inambao for his tireless effort to 
engine loads, but when the engine load was increased to high mold me, to my brother D.K. Marwa for his support and 
13687 
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© Research India Publications.  http://www.ripublication.com 
visionary leadership and lastly to the UKZN fraternity for the Engine,” Smart Grid and Renewable Energy, 6(04), p. 
suitable environment, especially my fellow post-graduate 49.  
students, which enabled me to complete this work. 
[10] Abu-Jrai, A., Rodríguez-Fernández, J., Tsolakis, A., 
 Megaritis, A., Theinnoi, K., Cracknell, R., and Clark, 
R., 2009, “Performance, Combustion and Emissions of 
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13689 
CHAPTER FIVE 
TRANSPORTATION, POLLUTION AND THE 
ENVIRONMENT 
Journal article 
Maroa, S., & Inambao, F. (2017). The Effects of Exhaust Gas Recirculation on the 
Performance and Emission Characteristics of a Diesel Engine–A Critical Review. International 
Journal of Applied Engineering Research, 12(23), 13677-13689. 
http://www.ripublication.com 
 
 
108 
 
International Journal of Applied Engineering Research ISSN 0973-4562 Volume 13, Number 6 (2018) pp. 3187-3199 
© Research India Publications.  http://www.ripublication.com 
Transportation, Pollution and the Environment 
 
Maroa Semakula1and Prof Freddie Inambao2 
1,2University of Kwazulu-Natal Durban South-Africa. 
 
 
Abstract IEA International Energy Agency 
The importance of modern day transport systems and the IF Inorganic Fraction 
transportation industry in the world economy cannot be over 
emphasized. However, it is now widely recognized that air IPCC Intergovernmental Panel on Climate Change 
pollution and environmental degradation are directly linked to NAC NOX Absorber Catalyst 
the rapid growth of automobile, air, rail, and sea modes of 
transport. These transportation services and their allied LNT Low NOX Trap 
industries are key contributors to every nation’s gross domestic NASA National Aeronautical and Space 
product. This work will endeavour to highlight the general Administration 
areas of concern with regard to pollution and environmental 
degradation as a consequence of this rapid growth and NH3 Ammonia 
expansion, with a focus on diesel emissions. These areas are (1) NNH NNH Route of NOX Formation 
Pollution in highly populated urban areas (cities) with their 
high rate of personal or private transport rather than public NOAA National Oceanic and Atmospheric 
transport systems; (2) The effects of modern day transport Administration 
systems on the depletion of the ozone layer, which is NOX Agglomerates of Oxides of Nitrogen 
responsible for climatic changes; and (3) The legislative and 
effective control systems required to meet the current NSR NOX Storage Catalyst 
environmental demands, which this work will endeavour to OECD Organisation for Economic Cooperation and 
critically present and analyse. Development 
Keywords; Environmental Degradation, Transport Industry, PAGASA National State Weather Bureau of Philippine 
Urban Pollution, Modes of Transport. 
PCV Positive Crankcase Ventilation 
 
Pd Palladium 
Nomenclature and Abbreviations 
PM  Particulate Matter 
ṁ𝐸𝐺𝑅 Mass Flow Rate of Egr Gases 
PM10 Particulate Matter of Size 10nm 
ṁ𝑎𝑖𝑟 Mass Flow Rate of Air  PM2.5 Particulate Matter of Size 2.5nm 
ṁ𝑒𝑥ℎ Mass Flow Rate from The Exhaust  Pt Platinum 
ṁ𝑓 Mass of Fuel Injected Rh Rhodium 
ṁ𝑖𝑛𝑡  Mass Flow Rate from The Intake Side SAWS South-African Weather Services 
𝑟𝐸𝐺𝑅 Percentage of Recirculated Egr Gases SCR Selective Catalytic Reduction 
CeO2 Cerium Oxide SI Spark Ignition 
CO Carbon Monoxide SO3 Sulphur Trioxide 
CO2 Carbon Dioxide SOF Soluble Organic Fraction 
DOC Diesel Oxidation Catalyst TDC Top Dead Centre 
DPF Diesel Particulate Filter VOCs Volatile Organic Compounds 
EC Elemental Carbon WHO World Health Organization  
EPA  Environmental Protection Agency ZrO2 Zirconium Oxide 
GBD Global Burden of Disease λ The Air Excess Factor Symbol Lambda 
HC Hydrocarbon Emissions  
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© Research India Publications.  http://www.ripublication.com 
INTRODUCTION  Cooperation and Development (OECD), the Intergovernmental 
Panel on Climate Change (IPCC), the European Environment 
The word ‘pollution’ has a wide variety of contextual meanings 
Agency (EEA) and the International Energy Agency (IEA) 
and applications, but in the context of this work it means the 
being constituted and empowered to act in this regard. To this 
damage caused by human activity to the environment and 
end these agencies have recommended legal measures and 
ecological systems. Considering all sources of the 
control systems, as well as supporting relevant technological 
environmental pollution and categories of pollution emitted and 
research and development in research to improve the pollutions 
excluding land use for the purpose of agriculture, the transport 
standards of the transport service industry. 
industry in the world is responsible for about one third of all 
environmental emissions of volatile organic compounds The reason this study focuses on diesel propelled engines is 
(VOCs) released into the atmosphere and the environment, because they have the highest proportion of usage globally as 
including two thirds of the carbon-monoxide (CO). The VOCs, compared to gasoline propelled engines, given their numerous 
which are hydrocarbon based, are a direct result of combustion, merits such as low-operational cost, energy efficiency, 
particularly inefficient combustion emitted from vehicle tail reliability, and durability. These merits have made diesel 
pipes as exhaust gases when they burn fuel to produce propelled engines the top choice of business leaders in the 
propulsion power. Some hydrocarbons are known carcinogens. transport industry as power sources for commercial 
Among the greenhouse gases, CO2 is the main culprit of global transportation in trucks, buses, trains, ships and for power 
warming and has reached 34 billion tonnes per year with an generation. Their environmental impact is caused by the 
increase of 3 % in 2011 alone [1] and is projected to rise to exhaust gases released via the tail pipe which contain very high 
41 billion tonnes by the year 2020 [2]. amounts of particulate matter (PM) and NOX emissions and 
other pollutants. These pollutants are responsible for severe 
Human activities, especially the use of and burning of fossil 
environmental degradation of air quality causing smog, acid 
fuels, enhance the greenhouse gas effect by increasing their 
rain and disturbance to the ozone balance. They also cause 
concentration levels in the atmosphere. The major gases 
major health problems, especially diseases related to the 
emitted through human activity are CO2, methane, oxides of respiratory system, and cancer [11]. 
nitrogen (NOX), fluorinated gases (hydrofluorocarbons, 
perfluorocarbons and sulphur hexafluoride) [3].   
There are several methods of VOC emissions control in the DIESEL ENGINE STATUS ON EMISSIONS IN THE 
modern day automobile, through the use of high pressure fuel TRANSPORT SECTOR  
injection systems, catalytic converters and positive crankcase 
From Tables 1 and 2 and presented below it can be seen that 
ventilation systems (PCV). However, 20 % of all uncontrolled 
transport contributes the largest amounts of NO
vehicular emissions are as a result of crankcase blow-by and X
, PM, CO, SO3 
and HC emissions annually, globally. The PM percentage 
evaporating oil. The particulate VOC pollutants, which result 
referred to in Tables 1 and 2 does not include dust from the 
from incomplete combustion and evaporating lubricating oil, 
road, rubber particulates from tyres, the photochemical smog 
have an emission rate of 0.6 g/bkWh to 2.2 g/bkWh for light 
particles arising from brake pads and asbestos brake linings 
duty diesel automotive propelled engines, whereas the large 
used in braking systems of automobiles passenger and 
ones have 0.5 g/bkWh to 1.5 g/bkWh [4, 5]. 
commercial vehicles. In addition, there is noise pollution which 
The global threat caused by human activities to the is predominantly attributed to heavy duty diesel vehicles 
environment is the second most serious issue after war facing especially around major urban centres and cities (from 
the world today. This is due to the rise in global warming levels highways, ports and stations) and air transport. 
above the pre-industrial levels by over 0.94 ºC, acidifying of 
 
the oceans, and 3.2 cm of sea level rise per decade. Extreme 
heat waves, global temperature rise and drought have affected Table 1. Atmospheric pollutants and their sources in 
food chains across the globe [6, 7] Unless these activities are percentage form 
legislated for, curbed and controlled, it is expected that globally 
the negative effects of climatic change will cause a warming of SOURCE NOX % PM % CO % SO3 % HC % 
4 ⁰ C and a sea level rise of 0.5 m to 1.0 m as early as 2060 [8]. INDUSTRY 22 44 4 32 26 
The transport industry which is a major contributor to the POWER GENERATION 32 21 0 48 0 
global economy is also a major cause of environmental SPACE HEATING 5 14 3 12 3 
pollution and climate change, especially due to the emission REFUSE BURNING 0 7 1 4 6 
from motor vehicles which accounts for 22 % of all the global 
TRANSPORT 42 14 92 4 65 
CO2 emissions [9]. Air pollution in developing countries and 
emerging economies accounts for tens of thousands of deaths,  
with the World Health Organization estimating an annual death Table 2 showing emissions of air pollutants by source in the 
toll from air pollution alone to be around 2.4 million people United States in 2011[12] 
using 2009 estimates and projections [10]. Thus, environmental 
protection, air pollution control and climatic change  
management have become issues of global concern, with many  
agencies such as the Environmental Protection Agency (EPA) 
of the United States of America, the Organisation for Economic  
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© Research India Publications.  http://www.ripublication.com 
Table  2.
 
 
Figure 1 shows the compositions of hydrocarbon pollutant of being released to the atmosphere via the exhaust manifold. 
diesel exhaust gas as analysed at the tail pipe exit [2]. Molecular UHC emissions are inherently independent of the air-fuel ratio 
nitrogen, which is an agent of NOX emissions, is the largest of a working engine, as they do not depend on the engine alone 
percentage at approximately 67 %. but also on other factors outside the engine itself. Such 
emissions can also be created, for example, if there is a 
 
malfunction of other systems or an operational malfunction of 
other systems, particularly the input data systems in modern 
injection and engine management systems.  
There are two features of a fuel injection systems in spark 
ignition engines (SI) which cause or can bring about HC 
emissions, namely, secondary injection and high nozzle sac 
volumes. However, when it comes to compression ignition 
engines, which this work is dealing, the hydrocarbon emissions 
are caused due to insufficient temperature near most cylinder 
walls of the compression ignition engines. Diesel propelled 
engines emit low hydrocarbons principally at light loads; the 
lean air fuel mixture which is the main source of light load 
emissions is caused by low speed flames which disallow 
completion of combustion during the normal cycle of the power 
 stroke and secondly due to lack of the actual combustion thus 
Figure 1. The compositions of hydrocarbon pollutants of leading to high emissions of hydrocarbons. 
diesel exhaust gas as analysed at the tail pipe exit [2]. 
There are three main factors that may affect the type of 
 hydrocarbons emitted, namely, engine adjustments, design and 
fuel type. Engine operating environment can also affect the 
THE FORMATION OF DIESEL ENGINE 
content of hydrocarbons, where temperatures are 400 ⁰ C to 
POLLUTANTS, AND THEIR EFFECTS  
600 ⁰ C in the combustion cylinder and when oxygen is still 
Unburned Hydrocarbons (UHC) available, hydrocarbons will continue to react in the exhaust. 
Therefore, this will lower emissions exiting the tail-pipe as 
Unburned Hydrocarbon (UHC) emissions are an indication of compared to the engine cylinder prior to it, stated in terms of 
the quality of combustion within an internal combustion their CH4 equivalent content [13, 14]. Emission consists of 
engine, arising from vaporized unburnt fuel and partially thousands of chemical species of alkanes, alkenes and 
burned fuel by-products exiting the combustion chamber and aromatics.  
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Hydrocarbon emissions are not limited to the vehicle exhaust chamber, namely: The Zeldovich mechanism, the prompt 
system only but also occur in the vehicle fuel system, from mechanism, through engine fuel mechanism, and the NNH 
vapors during fuel distribution and dispensing (15 % to 25 %) mechanism [19]. The Zeldovich mechanism is represented in 
and the engine crank-case itself (20 % to 30 %). However, the Equation 1, 2 and 3 with Equation 2 as the reaction rate 
largest culprit of all vehicular hydrocarbon emissions is the determining step equation [20]. 
exhaust tail pipe which accounts for 50 % to 60 % of all 
𝑂2 ⟷ 2𝑂                                                                  Equation 1 hydrocarbon emissions [15, 16]. 
𝑁2 + 𝑂 ⟷ 𝑁𝑂 + 𝑁                                                Equation 2  
𝑁 + 𝑂 ⟷ 𝑁𝑂 + 𝑂                                                Equation 3 
Oxides of Nitrogen (NOx) Emissions 2
The prompt mechanism, otherwise known as the Fenimore 
As shown and indicated in the literature discussed above, the 
mechanism [21] (named after the man who discovered it), 
transport industry (particularly the road component) is the 
involves reactions of the fragmented hydrocarbons with 
principal and primary source of nitrogen oxide (NO) emission 
molecular nitrogen under weak temperature dependency, thus 
and its oxidized products nitrogen dioxide (NO2). NO and NO2 
accounting for a relatively smaller portion of the hydrocarbon 
constitute 85 % to 95 %, while NO and N2 are lumped together 
emissions. Equations 4, 5 and 6 show the occurrence of prompt 
as NOX. There are fundamental differences between the two 
NO  [22]. 
gases; NO is colourless and odourless, while its counterpart X
NO2 is a reddish brown gas with a pungent smell and odour 𝑂 + 𝑁2  + 𝑀 →  𝑁2𝑂 + 𝑀                                   Equation 4 
[17]. These two gases are considered to be extremely toxic, 
especially NO2 which is five times more toxic than NO, which 
𝐻 + 𝑁2𝑂 ⟷ 𝑁𝑂 + 𝑁𝐻                                        Equation 5 
is a threat to human health as it affects the respiratory system 𝑂 + 𝑁2𝑂 ⟷ 𝑁𝑂 + 𝑁𝑂                                        Equation 6 
causing irritation and poor resistance to even simple respiratory 
infections like the common cold and influenza [2]. As mentioned earlier, the oxides of nitrogen form an important 
constituent of photochemical smog besides the greenhouse gas 
The temperature dependence of oxides of nitrogen is due to the effects, thus contributing in many ways to the depletion of the 
equilibrium concentration of NOX compounds of oxygen and ozone, causing acid rain, air pollution and a general decrease in 
nitrogen when they are mixed in a very high flame temperature the air quality in modern urban metropolises and cities in the 
adiabatically with temperatures of 2000 K to 3000 K [18]. world today. 
There are basically four types of mechanisms through which 
the nitrogen oxide emissions are formed and constituted in the 
combustion chamber and within the automobile combustion 
 
Table 3. Transport patterns and their contribution to overall emissions in selected cities 
 
 
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Table 3 shows transport patterns in 23 sampled cities in various Table 4. The components of SOF emissions and their total 
regions and the contribution of individual modes of transport to fraction percentages [4]. 
the overall emission percentages in these cities. The leading 
percentages of pollution by source as seen in Table3 is caused Fraction  Components of fraction Percentages of 
by public transport and private public transport vehicles, with total fraction 
Dallas in USA as the leading city in terms of private public Acidic Aromatic or aliphatic 3-15 
transport pollution, and Tokyo in Japan is the leading city in in Acidic functional groups 
terms of public transport pollution. Phenolic and carboxylic 
The amount of NOX produced during emission from the engine acids 
is a function of three factors, namely: the amount of time taken Basic Aromatic or aliphatic < 1-2 
or the combustion duration, the level of concentration of Acidic functional group 
oxygen, and, because NOX is temperature dependent, the Amines 
maximum existing temperature in the combustion cylinder. 
Most of the NOX emitted is formed in the early combustion Paraffin Aliphatic, normal and 34-65 
phase process with the engine piston nearing the end of branched 
compression or top dead centre (TDC) at which period the Numerous isomers 
combustion process propagated flame temperature is at From unburned fuel and /or 
maximum or at peak levels. The amount of NOX increases lubricant 
three-fold for every 100 ⁰ C increase in the cylinder Aromatic From unburned fuel, partial 3-4 
combustion temperature [16]. combustion, recombination 
 of combustion, from 
lubricants,  
Particulate Matter (PM) Emissions Single ring compounds 
PM emissions can be defined as the agglomerations of small Polynuclear aromatics 
particles resulting from the combustion process due to partly Oxygenated Polar functional groups but 7-15 
burned lubrication oil, ash content of the fuel used by the not acidic or basic 
engine, sulphates from the engine cylinder wall lubrication oil, Aldehydes, ketones, or 
and water from condensation and the combustion process. For alccohols 
example, in an experimental study of a heavy duty diesel Aromatic phenols and 
engine, the researchers classified and characterized the quinones 
emissions as 41 % as carbon, 7 % as unburned fuel, 25 % as 
unburned oil, 14 % as sulphates and water, and 13 % as ash and Transitional Aromatic or aliphatic 1-6 
other component emissions [23]. In an earlier study conducted Carbonyl functional groups 
by Agarwal (2007)[24] it was observed that PM emissions Ketones, aldehydes, esters, 
consisted of ≅ 31 % elemental carbon, ≅ 14 % sulphates and ethers 
moisture, ≅ 7 % unburnt fuel, and ≅ 40 % unburnt lubricating Insoluble High molecular weight 6-25 
oil, with the remaining percentages being metals and a variety organic species 
of other substances. Hydroxyl and carbonyl 
Diesel PM are particles usually in the region of groups 
15 𝑛𝑚 𝑡𝑜 40 𝑛𝑚 measured by diameter, though almost 85 % Inorganic compounds 
to 90 % of all PM emissions are less than 1 𝜇𝑚 in measured Glass fibre from filters 
diameter. PM emissions can be divided into three forms or main  
components: soluble organic fraction (SOF), soot, and the 
The size distribution of particles exhausted from diesel engines 
inorganic fraction (IF) of which 50 % is emitted as soot in the 
generally has two peaks: the nucleation mode which includes 
form of black smoke in exhaust tail-pipes. SOF is made up of 
all volatile hydrocarbons that are below 50 𝜇𝑚, and the 
absorbed or condensed hydrocarbons that are embedded within 
accumulation mode which is made up of soot and particles that 
the soot emissions in very fine particles derived partly, as 
are above 50 𝜇𝑚 [26, 14]. PM emissions formation is 
already mentioned, from lubrication oil, unburned fuel and 
controlled by a number of factors such as the combustion 
compounds of the combustion process from the combustion 
process, fuel quality (sulphur and ash content in the fuel), the 
chamber (Table 4). These emissions are more pronounced 
power stroke or the expansion process of the fuel, the engine 
when starting and during idling when the exhaust temperatures 
lubrication quality, the fuel consumption per engine cycle and 
are low [25]. 
the exhaust gas cooling [27].  
 
PM emissions impact negatively on human health as well as the 
 environment. For example, in 2010 alone the global burden of 
diseases (GBD) index attributed 3.2 million deaths to ambient 
 
pollution [28] where PM2.5 emissions were the main 
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contributor, thus highlighting the seriousness of air pollution Figure 2 shows the cause specific global mortality attributable 
from the transport industry.  to ambient PM2.5 emissions. 
 
 
Figure 2 shows global and regional distributions of deaths as a function of ambient PM2.5 concentration. 
 
 
Figure 3. Global and regional distribution of deaths as a function of ambient PM2.5 in different countries in quantile percentages. 
 
 
Carbon Monoxide Emissions  acceleration where rich mixtures are a conditional requirement. 
CO is defined as a colourless, tasteless and odourless toxic gas 
Carbon monoxide emissions result from the incomplete 
produced primarily due to the incomplete combustion of the 
combustion of hydrocarbon fuels as a result of failure of 
carbon containing fuels. The United States of America is the 
oxidation during the combustion process in a diesel engine, 
single largest producer of this pollutant gas and a leading 
particularly where excess air factor λ satisfies the condition 𝜆 <
producer of pollution from anthropogenic sources – 
1 for SI engines and thus the mixture is classified as a rich 
anthropogenic sources being the major contributor of this gas, 
mixture, which occurs especially during start-ups and sudden 
as seen from Figure 4.  
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Effects of Traffic Pollutants on the Natural Environment 
The demand for energy supply associated with economic 
growth and development causes emissions which deplete the 
ozone layer. The ozone layer raises particular concern for 
ecology since it leads to sun radiation exposure to ultra-violet 
rays in the 200 𝜇𝑚 to 300 𝜇𝑚 range which is toxic to 
unicellular organisms and most surface cells of animals and 
[31].  
Pollution from automobiles is one of the leading causes of 
global warming, as cars (light duty and heavy duty) release 
greenhouse gases and carbon dioxide into the atmosphere 
through their exhaust systems. These types of gases tend to trap 
heat in the atmospheric environment thus causing a global rise 
in temperatures. Because of the combustion of increasing 
 volumes of fossil fuels, global temperatures have risen by 0.5 
⁰ C to 1 ⁰ C since the pre-industrial age. The global rise in 
Figure 4. carbon monoxide (CO) emissions by anthropogenic 
temperature or global warming affects the natural environment 
and biogenic sources in the USA [29]. 
by causing higher sea levels, damaging farming activities 
 which have a bearing on food security, damaging wild life, and 
destroying natural habitats and landscapes 
The measured average CO concentration for fuel rich mixtures 
is very close to equilibrium in burned gas especially during the  
expansion process. However, for lean fuel mixtures of which 
Effect of Greenhouse Gases and Global Warming 
diesel engines form a large segment, measured CO emissions 
are higher than most prediction models of any kind. One Since the first recorded temperatures in 1860, earth 
possible explanation for this is that from a practical point of temperatures have risen by over 1 ⁰ C, and 2017 was one of the 
view, CO oxidation mechanisms are determined by the fuel air hottest years after 2016 in human history. The first six months 
equivalence ratio. This phenomenon in the rich mixture areas of 2017 recorded an average surface temperature of 1.1⁰  C 
and lack of availability of enough oxygen makes the reactant above the 1950 to 1980 averages [32]. 
concentration fail to convert fully during combustion, thus 
exhausting a smaller portion through the tail-pipe due to the 
kinetic effects. CO emissions tend to form and accumulate in 
areas of heavy concentrated traffic, parking garages, and under 
building overheads and overhangs. Among the human health 
effects of CO emissions due to exposure are headaches, 
dizziness and in extreme cases of exposure, death.  
 
Figure 6. Annual global temperature anomalies from 1901 to 
2016 relative to the global annual average temperature for the 
20th century (1901-2000) [32]. 
 
Figure 6 shows the way in which the annual global temperature 
has consistently been above the global average for the 20th 
century, climbing steeply since 2010. Data is from the US 
National Oceanic and Atmospheric Administration (NOAA) as 
cited in [32]. 
Coming closer home to South Africa, the Western Cape 
 province has experienced unprecedented drought and severe 
unpredictable weather patterns for the last three years. South 
Figure 5 is a pie chart showing how anthropogenic sources 
Africa had intense 2016 January heatwaves in about 62 
contribute to carbon monoxide emissions with the largest share 
locations nationally with recorded temperatures reaching up to 
of 55 % being contributed by the transportation industry whose 
46 ⁰ C in Tosca in North West province according to the South 
core players are heavy and light duty vehicles on the road [30]. 
African Weather Services (SAWS) [33] 2016. Botswana 
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recorded and broke a national record of 72 years for its Table 5. EURO standards for heavy duty vehicles according 
maximum temperature in Maun at 44⁰ C. to Delphi 2016 to 2017 as per Walker (2016) [37] and [2]. 
Alaska in the United States of America had its warmest month  CO HC NOx PM 
in February 2016 breaking a 96-year national record predating 
th (g/kWh) (g/kWh) (g/kWh) (g/kWh) the 20  century average with increment levels of 6.9 ⁰ C 
according to the NOAA, with California and Australia Euro I 4.5 1.1 8.0 0.61 
recording unending wild fires due to excessive drought and Euro II 4 1.1 7.0 0.15 
heat, leading to massive loss of life and property besides the Euro III 2.1 0.66 5.0 0.13 
economic loss suffered by individuals and the state [32]. 
Euro IV 1.5 0.46 3.5 0.02 
 Euro V 1.5 0.46 2.0 0.02 
The Effects of Traffic Pollution On Agriculture and Food Euro VI 1.5 0.13 0.4 0.01 
Security  
As a result of the climatic changes that have been enumerated From Table 5 it can clearly can be seen that the standards have 
and discussed in previous sections, there has been a shift in the become more stringent, thus obliging the vehicle manufacturers 
meteorological equator leading to changed forecast trends and and industry service providers to work harder towards reducing 
scenarios of desert expansion, drought and change in weather emissions from public, private, and commercial vehicles. For 
patterns across the globe [34]. The following are examples of almost two and a half decades now (from 1993 to 2017) studies 
scenarios that have already started happening: have been carried out on engine modifications and electronic 
 Reduction of rainfall, turning some areas like California control systems for fuel injection have been introduced, along 
into deserts, with similar changes being experienced in with tremendous improvement on fuel properties and 
the Cape provinces in South Africa, where the water development. Among the techniques investigated to cut down 
conditions are worsening. Punjab in India and other emissions in combination with other techniques are: the use of 
areas are facing intense heat waves with the countryside exhaust gas recirculation (EGR), lean NOX trap (LNT), diesel 
experiencing droughts of untold proportions. oxidation catalyst (DOC), diesel particulate filter (DPF) and the 
selective catalytic reduction (SCR) technique [38]. However, 
 There has been an increase in the intensity of severe these developments have failed to attain sufficient reduction of 
weather conditions, where the severity of typhoons has pollutant emissions to the required regulatory standards as 
been increasing in places like Philippines with 2016 prescribed by the controlling environmental protection 
alone accounting for more than 20 tropical cyclones agencies.  
according to PAGASA, the national state weather 
bureau of Philippines. In 2013 35 000 to 39 000 people  
lost their lives and damage of more than 40 billion USD Exhaust Gas Recirculation 
was caused by extreme weather conditions, according to 
the national disaster risk reduction management council Exhaust gas recirculation is a system which allows the 
of Philippines [35, 36].  recirculation of the exhaust gases back into the combustion 
chamber for mixing and reburning with the fresh charge [39] 
 Expansion of the African desert beyond Africa, to Spain (Figure 7). This technology, though able to reduce NOX, leads 
Greece and Italy, besides the sub-Saharan countries to other problems such as an increase in HC and CO emissions 
known as the Sahel countries, which are currently due to the lowered combustion temperature besides affecting 
experiencing severe drought and food insecurity. overall engine efficiency. While it is accepted that universally 
 there is no standard definition of EGR to be able to quantify the 
amount of EGR recirculated, there are two methods that are 
AUTOMOBILE POLLUTION CONTROL MEASURES widely accepted and are commonly used to define EGR: the 
The world is now pre-occupied by environmental protection mass based method and the gas concentration method [40]. 
and the dangers of human activities that degrade the 
environment. From the United Nations to organizations in 
regional blocks and individual countries, everyone is 
embarking on saving and preventing extensive damage to the 
environment and human health as a result of emissions of 
pollutants. 
Due to the effects of diesel engine emissions on human health 
and the environment, government agencies in charge of human  
health and the environmental development have put forward Figure 7. Schematic representation of an EGR system and 
requirements and permissible standards to curb vehicular some of its nomenclature and control design for the EGR 
emissions. In Europe the European Environmental Agency valve. 
(EEA) has been in operation since 1993 with recommendations 
and standards of Euro I to Euro IV, as shown in Table 5. 
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Both of these methods can be demonstrated from the diagram Low NOX Trap (LNT) 
shown in Figure 7 and can also be expressed mathematically as 
The low NOX trap (LNT) system is has two other names by follows in Equations 7 and 8 [41].  
which it is also known: NOX storage reduction (NSR) and NOX 
ṁ
𝑟 = 𝐸𝐺𝑅                                              Equation 7 absorber catalyst (NAC). The LNT has three main components, 𝐸𝐺𝑅 ṁ𝑎𝑖𝑟+ṁ𝑓+ṁ𝐸𝐺𝑅 namely, the oxidation catalyst made of platinum (Pt), the NOX 
[𝐶𝑂2]𝑖𝑛𝑡−[𝐶𝑂2]𝑎𝑚𝑏 [𝐶𝑂2]𝑖𝑛𝑡 storage ambience made up of barium (Ba) together with other 𝑟𝐸𝐺𝑅 ≈ ≈                         Equation 8 [𝐶𝑂2]𝑒𝑥ℎ−[𝐶𝑂2]𝑎𝑚𝑏 [𝐶𝑂2]𝑒𝑥ℎ oxides, and a reduction catalyst made up of Rhodium (Rh). A 
platinum catalyst is preferred due to its ability to reduce NO  
Where the ṁ𝐸𝐺𝑅 is the mass flow rate of the gas recirculated, 
X
emissions even at very low temperatures while still offering 
ṁ𝑎𝑖𝑟 is the mass flow rate of fresh air, ṁ𝑓 is the mass flow rate stability in the presence of sulphur and water moisture. 
of the injected fuel and 𝑟𝐸𝐺𝑅 is the mass fraction of the However, this technology cannot offer all the solutions to 
recirculated exhaust gases. 
vehicular emissions as a stand-alone technology. 
 
 
 
Figure 7. The low NOX trap (LNT) has three operating modes of operation: NOx storage during lean engine operation, NOx 
reduction during rich operation phases, low NOX trap (LNT) desulfurization under rich conditions and high temperatures [42]. 
 
Selective Catalyst Reduction  After the hydrolysis and thermolysis process, the following are 
the chemical reactions which take place in the SCR catalyst as 
The most recent developments have seen the introduction of the 
indicated by Equations 11, 12 and 13 [22] and [4]. 
SCR technology especially for heavy duty vehicles [43]. It has 
already been in use with light duty vehicles. For example, Audi 4𝑁𝑂 + 4𝑁𝐻3 + 𝑂2 → 4𝑁2 + 6𝐻2𝑂                        Equation 11 
motors and Volkswagen motors have widely adopted these 
 2𝑁𝑂 + 2𝑁𝑂 + 4𝑁𝐻 → 4𝑁 + 6𝐻 𝑂                 Equation 12 
technologies for most of their passenger vehicles. SCR works 2 3 2 2
by utilizing ammonia as a reductant to minimize NOX 6𝑁𝑂2 + 8𝑁𝐻3 → 7𝑁2 + 12𝐻2𝑂                            Equation 13 
emissions in the exhaust gases emitted releasing N2 and H2O. 
There are two processes that an SCR catalyst system undergoes Figure 8 shows a schematic diagram of a car exhaust gas 
namely hydrolysis and thermolysis. Equation 9 for the emissions control system comprising an oxidation catalyst, 
hydrolysis process and Equation 10 for the thermolysis process wall-flow particulate filter, and flow-through SCR catalyst. 
summarizes the two processes [2] and [22]. Key components include a urea solution tank (heated in cold 
weather), dosing spray module and static mixer, temperature 
𝐻𝑁𝐶𝑂 + 𝐻2𝑂 → 𝑁𝐻3 + 𝐶𝑂2                                  Equation 9 and NOX sensors (Robert Bosch GmbH) [44]. 
(𝑁𝐻2)2𝐶𝑂 → 𝑁𝐻3 + 𝐻𝑁𝐶𝑂                                 Equation 10 
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Figure 8. Schematic diagram of the SCR NOX control system as used in a standard production vehicle [44]. 
 
Diesel Particulate Filters (DPF) carbon particles. The DOC is made up of a metal or a ceramic 
structure with an oxide mixture (wash coat) composed of 
DPF have been applied in automotive manufacturing 
aluminum oxide (Al2O3), cerium oxide (CeO2), zirconium 
production units since the year 2000 and are primarily used to 
oxide (ZrO2) and an active catalytic metal such as platinum 
remove PM emissions from the exhaust gases through physical 
(Pt), palladium (Pd) and rhodium (Rh) [46]. 
filtration. Most of the DPFs are made in a honey comb structure 
of silicone carbide or cordierite (2𝑀𝑔𝑂 − 2𝐴𝑙2𝑂2 − 5𝑆𝑖𝑂2)  
with both ends of the monolith structure blocked so that 
particulate matter is forced through the porous substrate walls 
thus acting as a mechanical filtering system. The walls of the 
DPF filter are made in such a way that they enable exhaust 
gases to pass through the walls with little resistance while 
maintaining the capability to collect PM emissions particle 
species as shown on Figure 9. 
               
Figure 10 shows a schematic diagram of a DOC and its 
operation in reducing emissions through the process of 
oxidation. 
 
There are six factors which influence the choice of DOCs, 
namely conversion efficiency, temperature stability, light-off 
 temperature, tolerance to poisoning, the cost of manufacturing, 
and parametrical factors. Parametrical factors include channel 
Figure 9. Schematic of the working mechanism of a diesel 
density measured in channels per square inch, the individual 
particulate filter (DPF) filter. With DPF, care must be taken to 
channel wall thickness, the cross-sectional area, and the length 
avoid excessive saturation of the filter as this builds up back 
of the channels (external dimensions) [47]. 
pressure which is harmful for engine operation and durability, 
increases fuel consumption, increases engine stress levels and The DOCs are the preffered choice for most emission control 
premature failure of the filter and engine [45]. systems for heavy duty and light duty vehicles in Europe, the 
USA and Japan, with the DOCs that contain platinum and 
 
palladium being the most popular among manufacturers and 
The Diesel Oxidation Catalyst (DOC) consumers alike in the world market. However, DOCs cause a 
reaction with sulphur oxide and sulphur tri-oxide thus 
The main purpose of a diesel oxidation catalyst (DOC) is the 
generating sulphates and sulphuric acid which shortens the life 
oxidation of the CO and HC emissions by the reduction of the 
of such control systems and causes several environmental and 
PM through oxidation of the hydrocarbons absorbed into the 
human health issues. 
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CONCLUSION The Hague: PBL Netherlands Environmental 
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From the current work and as discussed in the literature, the 
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[4] Heywood, J. B., 2012, Internal Combustion Engine 
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ACKNOWLEDGEMENTS Clinical Immunology, 136(2), pp. 441-453.  
First and foremost, to the good Lord for His enabling grace and [12]   IEA, C. (2012). emissions from Fuel Combustion. 
mercies, to my supervisor Prof FN Inambao for his tireless International Energy Agency. [13] Yamada, H., Misawa, K., 
effort in molding me, to my brother David K. Marwa for his Suzuki, D., Tanaka, K., Matsumoto, J., Fujii, M. and Tanaka, 
vision and effort and tireless support, to my family members K., 2011, “Detailed Analysis of Diesel Vehicle Exhaust 
especially my sisters Prof M.B Mwita and Dr I.N Marwa. Emissions: Nitrogen Oxides, Hydrocarbons and Particulate 
Lastly, to the UKZN University fraternity for the enabling Size Distributions,” Proceedings of the Combustion Institute, 
environment, and to my fellow post-graduate students who 33(2) pp. 2895-2902.  
helped me to complete this work.  
[14] Shirneshan, A., 2013, “HC, CO, CO2 and NOx 
 Emission Evaluation of a Diesel Engine Fueled with 
Waste Frying Oil Methyl Ester,” Procedia-Social and 
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Cycles,” Applied Energy, 123, pp. 242-252.  
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Combustion Characteristics and Particle Number and Exhaust Aftertreatment Systems, FEV Motorentechnik 
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[43] Yun, B. K. and Kim, M. Y., 2013, “Modeling the 
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Exhaust Gases,” Applied Thermal Engineering, 50(1), 
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[44] Nova, I. and Tronconi, E., 2014, Urea-SCR technology 
for deNOx Aftertreatment of Diesel Exhausts, Springer, 
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Lefebvre, D., 2017, “Model-Based Passive and Active 
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Applied Thermal Engineering, 110, pp. 962-971.  
 
3199 
CHAPTER SIX 
THE FORMATION, EFFECTS AND CONTROL OF OXIDES 
OF NITROGEN IN DIESEL ENGINES 
Journal article 
Maroa, S., & Inambao, F. (2018). The Formation, Effects and Control of Oxides of Nitrogen 
in Diesel Engines. International Journal of Applied Engineering Research, 13(6), 3200-3209. 
http://www.ripublication.com 
 
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© Research India Publications.  http://www.ripublication.com 
The Formation, Effects and Control of Oxides of Nitrogen in Diesel Engines 
 
Maroa Semakula1and Prof Freddie Inambao2 
1,2University of Kwazulu-Natal Durban South-Africa. 
 
 
Abstract HPL High Pressure Loop EGR 
The transport service industry is a heavy user of diesel IDICI Indirect Injection Compression Ignition 
propelled engines as prime movers of goods and services. The 
diesel propelled engine is praised due to its high fuel efficiency, L Length or Piston Stroke 
reliability and durability. However, the nitrogen emissions as a 
result of diesel fuel combustion characteristics raise major LPL Low Pressure Loop EGR 
concerns for the manufacturing industry, environmentalists and 
health care researchers. The manner in which diesel engines M Organic Residue 
combust their fuel is the main cause of the nitrogen oxide 
NG Natural Gas 
emission proportion. Although there are other sources of 
nitrogen oxide emission, this work will cover the sources of NO  Nitrogen Oxide 
nitrogen oxides and their formation within the diesel engine, 
their routes of formation, identify the mechanisms under which NOX Oxides of Nitrogen Excluding Nitrogen Trioxide 
the formations occur, identify their types and interactions, look Up 
at the various effects of the oxides of nitrogen on human health 
and the overall damage to the natural environment, and look OH Water or Hydroxide Radical 
critically at control systems.  
P Pressure 
Keywords: Nitrogen Oxides, Human Health, Formation 
Mechanisms, Durability. PCCI Pre-Mixed Charge Compression Ignition 
 PM Particulate Matter 
NOMENCLATURE AND ABBREVIATIONS rc Compression Ratio 
A/F Air-Fuel Ratio 
RCCI Reactivity Charge Compression Ignition 
ATDC After Top Dead Centre 
SCR Selective Catalytic Reduction 
B Cylinder Bore 
SI Spark Ignition 
BTDC Before Bottom Dead Centre 
SO2 Sulphur Dioxide 
CA Crank Angle 
SOI Start of Injection 
CI Compression Ignition 
UV Ultra-Violet Rays 
CN  Carbon of Nitrogen  
Vc Clearance Volume 
CO Carbon Oxide 
Vd Swept or Displaced Volume 
DICI Direct Injection Compression Ignition 
λ The Air Excess Factor Symbol Lambda 
DOC Diesel Oxidation Catalyst  
DPF Diesel Particulate Filter INTRODUCTION 
EGR Exhaust Gas Recirculation The modern-day diesel engine, also known as the compression 
ignition engine, offers high fuel efficiency, low engine noise, 
F/A Fuel-Air Ratio reliability and durability during its life and service. However, 
diesel engines produce more oxides of nitrogen emissions than 
HCCI Homogeneous Charge Compression Ignition  their counterparts, spark ignition (SI) engines. This collection 
of oxides of nitrogen emissions are collectively referred to as 
HCN Hydrogen Cyanide 
NOX. Along with particulate matter emissions (PM), these 
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emissions have become major environmental and health Prompt NOx or The Fenimore Mechanism 
concerns and have also become important themes in global 
Prompt NOX, also known as Fenimore mechanism (named after discussions and forums.  
the person who discovered it [8], accounts for the smallest 
Due to the ever increasingly stringent regulations on emissions contribution to the quantity of NOX. It is usually formed at the 
by environmental protection agencies and governments, there flame front especially in rich fuel-air ratio areas due to 
has been development and improvement in design to unavailability of oxygen, from the radicals of CH through their 
accommodate and conform with the growing list of emission oxidation when they react with molecular nitrogen (N2), 
standards, especially developments in diesel fuel improvement leading to the formation of cyanhydric acid (HCN) and nitrogen 
[1] and [2]. Among the fuels for the 21st century being oxide (NO) at the termination of the reactions. The general 
proposed as an alternative to diesel fuel is natural gas (NG), scheme of prompt NOX causes hydrocarbon radicals to react 
which is highly promising and attractive according to Abdelaal with molecular nitrogen to form amines or cyano compounds 
and Hegab (2012) [3]. The advantages of NG include which are then transferred and converted to intermediate 
availability, lower price, and reduction in CO emissions due the compounds thus forming NO as can be shown in equation 1 and 
low carbon to hydrogen ratio. Further, because of its high equation 2 below and expressed as: 
octane number, NG has a high auto-ignition temperature 
𝐶𝐻 + 𝑁 ⟷ 𝐻𝐶𝑁 + 𝑁    Equation 1 
characteristic which eliminates compression ignition (CI) 2
engine knock such as it occurs during high compression ratios 𝐶 + 𝑁2 ⟷ 𝐶𝑁 + 𝑁    Equation 2 
due to the low octane number of diesel fuel. Above all, it is eco-
 
friendly with clean combustion compared to conventional 
diesel fuel engines. NOX emissions are approximately 50% to Where equation 2 forms the primary path and becomes the rate 
80% less, with zero smoke and PM emissions – something that limiting step in the reaction chain of the entire mechanism. 
is very hard to achieve in basic diesel propelled engines [3].  However, it is vital to mention here that in the diesel engine 
combustion process, NOX is generated through the Fenimore Besides the use of alternative fuels as a method of reducing 
mechanism or prompt NO  and the thermal mechanism or 
NOX emissions, the use of other control strategies and measures 
X
Zeldovich mechanism only. The conversion of hydrogen 
besides fuels has become equally important in mitigating the 
cyanide (HCN) to form NO takes the following path as 
effects and impact of NOX emissions. One of the techniques 
expressed in the following equations as: 
that has gained widespread use, although not as a standalone 
technology, is exhaust gas recirculation (EGR). This technique 𝐻𝐶𝑁 + 𝑂 ⟷ 𝑁𝐶𝑂 + 𝐻    Equation 3 
has been found to be an effective tool in reducing the emissions 
of NOX [4, 5]. 
𝑁𝐶𝑂 + 𝐻 ⟷ 𝑁𝐻 + 𝐶𝑂    Equation 4 
Although the combustion of biodiesels and its blends is 𝑁𝐻 + 𝐻 ⟷ 𝑁 + 𝐻2    Equation 5 
different from conventional diesel and fossil fuel in 𝑁 + 𝑂𝐻 ⟷ 𝑁𝑂 + 𝐻    Equation 6 
compression ignition engines due to their physio-chemical fuel 
properties, they do also cause emission of NOX, PM, CO and  
HC [6]. There has been a tremendous development in 2.2 The Thermal NOx (Zeldovich Mechanism) 
combustion technologies for biodiesel fuel combustion in 
diesel compression ignition engines, to accommodate the The Zeldovich mechanism which takes the intermediate route 
developments witnessed in the growth and expansion in is an important mechanism in lean fuels Փ < 0.8 operating 
alternative fuels, especially the biodiesels. Among the under low temperature conditions as can be seen in the three 
strategies that have been developed are: homogeneous charge equations below with M as the organic residue [9] as: 
controlled ignition (HCCI), pre-mixed charge controlled 
𝑂 + 𝑁 + 𝑀 → 𝑁 𝑂 + 𝑀    Equation 7  
ignition (PCCI) and reactive charge controlled ignition (RCCI). 2 2
These strategies have been extensively studied and found to 𝐻 + 𝑁2𝑂 ⟷ 𝑁𝑂 + 𝑁𝐻    Equation 8 
significantly reduce NOX and other emissions [7].  
𝑂 + 𝑁2 ⟷ 𝑁𝑂 + 𝑁𝑂    Equation 9 
 
The Zeldovich scheme mechanism consists of three main 
SOURCES OF OXIDES OF NITROGEN AND reaction chains as shown in the following equations and are 
FORMATION coupled to fuel combustion chemistry through O2, O and OH 
species [10] as follows: 
There are five mechanisms of NOX formation in a diesel engine 
combustion process:  𝑂 + 𝑁2 ⟷ 𝑁𝑂 + 𝑁    Equation 10 
 Fenimore CN and HCN pathways. 𝑁 + 𝑂2 ⟷ 𝑁𝑂 + 𝑂    Equation 11 
 N2O intermediate route or the Zeldovich mechanism. 𝑁 + 𝑂𝐻 → 𝑁𝑂 + 𝐻    Equation 12 
 Due to super equilibrium concentrations of O and OH. 
 Radicals in the extended Zeldovich mechanism The thermal NOX formation of oxides of nitrogen depend on 
reactions. the following factors for their formation and propagation:  
 The NNH route. 
 Temperature especially in the reaction zone. Other 
 than the disassociation process in the equation 2 of 
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oxygen, the reaction chain is inherently a temperature Due to the fact that hydrocarbon fuel contains organic bonded 
function as concluded by the studies conducted by nitrogen (organic nitrogenous compounds) some of the in-
Dangar and Rathod (2013) [11].  oxidized nitrogen will eventually be oxidized to be NOX. The 
 The equivalence ratio or the air-fuel ratio in the percentage amount of nitrogen going through this process 
reaction zone as it influences the atomic oxygen depends on the nature of the combustion process. This is 
concentration within a combustion mixture. It has compounded by the fact that light distillate fuels contain 0.06 
been observed that the NOX emissions decrease with a % by volume of organic hydrogen, while the heavy distillates 
decrease in the air-fuel ratio [12]. contain 1.5 % or above. Thus, it can clearly be seen that 
 The amount of time or duration of the reacting gases, depending on the amount of or percentage of nitrogen 
also called retention time, spent in the reaction zone at converted into fuel, NOX can contain large portions of and 
maximum temperature determine the amount of NOX percentages of total NOX emissions [20].  
formed. The shorter the time the smaller the quantity 
The oxidation of the nitrogenous compounds is known to be 
of NOX and vice versa. less temperature dependent; thus, oxidation can even occur at 
 low temperatures as the air excess ratio increases. However, 
during combustion a portion of the nitrogen fuel is changed to 
The NNH Mechanism NOX especially in fuel rich areas of combustion where 
This is a recently discovered reaction pathway [13], with two thermally formed NOX can be reduced to molecular nitrogen 
major reaction mechanisms which has been shown to be [21] as demonstrated by Figure 1 with the reactions taking the 
important especially in the combustion of hydrogen [14, 15] following forms: 
and hydrocarbon fuels which have large carbon to hydrogen 1
𝑁𝐻 + 𝑂 → 𝑁𝑂 + 𝐻     Equation 15 
ratios [16]. The following two equations illustrate this route as 2 2 2
follows: 1 1
𝑁𝐻3 → 𝑁2 + 1 𝐻2    Equation 16 2 2
𝑁2 + 𝐻 → 𝑁𝑁𝐻     Equation 13 1
𝑁𝑂 + 𝐶𝑂 → 𝑁2 + 𝐶𝑂2    Equation 17 𝑁𝑁𝐻 + 𝑂 → 𝑁𝑂 + 𝑁𝐻    Equation 14 2
1 1
 𝐻𝐶𝑁 + 2𝑁𝑂 → 1 𝑁2 + 𝐶𝑂 + 𝐻2 2 2 2
  Equation 18 
The Fuel NOx  
Fuel NOX can be defined as the reaction of fuel bound nitrogen 
from the compounds of N-H or N-C leading to the formation of 
ammonia (NH3) and cyanhydric acid (HCN), which, when 
disassociated during the chain reaction, form NO. For lean fuel-
air ratio mixtures almost two thirds of fuel bound nitrogen 
becomes converted to NO with the rest being converted to 
molecular nitrogen N2 [17]. However, for rich mixtures (that is 
fuel-air ratios that have high fuel-to-air ratios), less NO is 
formed but more ammonia and cyanhydric acid (HCN) is 
produced, which, when released to the atmosphere, form NO 
through decomposition [18]. It is imperative to mention here 
that the NO formed from this mechanism and route cannot be 
optimized and controlled through combustion. The various 
forms of nitrogen bound fuel from different fuel sources are 
shown in Table 1. 
Table 1: Forms of nitrogen bound fuel from different fuel 
 
sources [19] 
Figure 1: The relationship between temperature and NOX 
FUEL TYPE PERCENTAGE OF FUEL BOUND formation, the air excess factor also known as the oxygen 
NITROGEN  (WEIGHT%, DRY, ASH excess factor, and the parameters of thermal NOX formation, 
FREE BASIS) fuel NOX and prompt NOX [22] 
COAL 0.5-2.0 
 
BIOMASS (WOOD) < 0.5 
From Figure 1 we can see the relationship between combustion 
PEAT 1.5-2.5 
and the NOX formation path. The figure shows a positive 
FUEL OIL 0.0 
correlation between fuel and prompt NOX with temperature that 
NATURAL GAS 0.0 is almost linear with the amount or quantity of thermal NOX 
DERIVED GASES 0.1-1.0 increasing disproportionately with increasing combustion 
(>>1 chemical sources) temperature. In addition, the figure shows that those areas 
 where there is a rich fuel-air ratio naturally have reduced 
oxygen availability, thus making insufficient oxygen available 
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to react with the molecular nitrogen (N2) in the combustion THE EFFECTS OF OXIDES OF NITROGEN 
mixture, leading to reduced fuel conversion and thermal NOX. 
The oxides of nitrogen are recognised now as causing negative 
effects for human health and the natural environment. The 
harmful effects of oxides of nitrogen pollution do not 
necessarily take effect immediately after exposure, but after 
prolonged exposure. In the environment, oxides of nitrogen 
pollutants are known to cause destruction of the ozone layer, 
acid rain, poor visibility due to smog, poor air quality, and they 
contribute to the rise in earth surface temperatures leading to 
global warming, All these effects cause serious environmental 
degradation and poor human health, thereby increasing the rate 
of natural disasters and the global health burden, requiring 
billions of dollars for repair and reconstruction of the destroyed 
environment and of health. 
 
Global Warming 
Among the fundamental effects of oxides of nitrogen emission 
is global warming, although these are not the only emissions 
that contribute to global warming of the earth’s surface. The 
 
stratosphere ozone layer is destroyed by oxides of nitrogen 
Figure 2: The relationship between the various types of NOX emissions therefore increased high energy ultra-violet rays 
formation under the influence of varying combustion (UV) warm up the earth’s surface due to the decreased 
temperature ranges reflective capacity of the lower atmospheric layers. 
 The increase in earth’s surface temperature has serious 
Figure 2 shows the relationship between the various types of repercussions, shifting the climatic zones and thus shifting 
habitable regions of the earth leading to increase of the deserts 
NOX formation under the influence of varying combustion 
temperature ranges where thermal NO , prompt NO  and fuel and increasing the unpredictability of the weather patterns. For X X
NOX set in, and the quantity of NOX emissions produced under 
example, flooding and an increase in the intensity of typhoons 
have been witnessed in recent times or example in the USA 
varying temperatures. A conversion mechanism of NOX from 
fuel bound nitrogen shows the following trends and (hurricane Katrina) and in the Philippines (typhoon Haiyan), 
characteristics as can be seen in Figure 2 above as suggested by leaving behind massive destruction of infrastructure, life and 
[23]:  economies of the hit areas [24]. 
 
 Conversion of fuel based nitrogen to NOX is 
practically complete for fuel lean flames operating on Smog and Visibility 
low nitrogen concentrations of < 0.5% by weight 
Smog is a combination of smoke and fog and represents cloud 
ratio. 
formation from photochemical reactions of the sun with the 
 Conversion increases slowly with increase in flame 
hydrocarbons and oxides of nitrogen emission effluents from 
temperatures as indicated by all routes and 
automobiles and stationary engines. Besides automobiles, 
mechanisms of NOX formation in preceding sections 
smog can also be derived from other sources like coal 
and Figure 2. 
emissions, industrial emissions, frost, agricultural fires and 
 The composition of the nitrogen bearing compounds 
natural causes. Smog irritates eyes and the throat, and causes 
do not have any effect on the nitrogen bound fuel 
impairment to the lungs decreasing capacity, emphysema, 
conversion rate percentage. 
bronchitis, asthma, inflammation of the breathing passages, 
It is imperative to mention here that according to Figure 1, the shortness of breath, damages plants and crops, and destroys 
temperature range at which fuel NOX sets in is relatively higher rubber products by cracking them through deterioration. One of 
than the prompt NOX, but it is less than the thermal NOX the major impacts of smog has been poor visibility in cities such 
formation temperatures which kick in at temperatures as Beijing (China), New Delhi (India), and London (United 
exceedingly 1800 ⁰C. Table 1 shows that fuels with a high Kingdom) amongst others. Poor visibility is caused by nitrogen 
nitrogen content have a natural inclination to produce more oxide absorbing the full visible spectrum of the light energy, 
NOX emissions than those with a low nitrogen content. thus leading to poor visibility even in the absence of particulate 
However, it is important to note that in diesel propelled engines matter (PM) that would cause physical absorption of light. 
NOX formation from nitrogen fuel is strongly controlled and  
overlapped by thermal NOX formation, therefore there is no 
guarantee that low nitrogen fuel content would not cause high  
production of NOX emissions.   
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The Ozone  From manufacturing industries, oil refineries and 
other industrial manufacturing processes. 
The tremendous increase in the urbanization has seen an 
increase in energy consumption putting millions of cars on the There are two forms of acid deposition mechanisms – wet 
road [25, 24]. As a consequence, many cities are feeling deposition and dry deposition. The former is where sulphuric 
pressure from severe air pollution and air quality. The oxides acid and nitric acid in the atmosphere fall to the ground in the 
of nitrogen are key to the formation of ozone and aerosols form of rain, snow, fog or hail, and the latter is when acidic 
(𝑃𝑀2.5) in the atmosphere which have a negative effect on air particles and gases deposit on surfaces of water bodies, on 
quality, acid deposition and the balance of atmospheric vegetation and buildings in the absence of moisture but quickly 
radiation [24]. react during atmospheric transportation to form particles that 
can harm human health. The dry mechanism largely depends 
The consequence of the anthropogenic pollutants like NOX in 
on the amount of rainfall an area receives to cause maximum 
decreasing the thermal layering of the stratosphere and increase 
effect. 
of earth’s surface temperatures cannot be over-emphasized. 
Animals, plants and humans are known to be sensitive to UV-  
B and UV-C rays of radiation which are richer in radiation 
3  FACTORS OF NOX EMISSIONS AND CONTROL energy. Exposure to a concentration level of 2500 μm/m  of
NOX for one-hour decreases a human being’s lung volume and NOX or oxides of nitrogen emissions control is the most 
maximum breathing thus impairing breathing and can lead to difficult problem engineers and environmentalists have been 
death, while exposure to ozone causes pulmonary haemorrhage grappling with since the discovery of the automobile, since 
with symptoms such as a dry throat, severe headache, most control methods tend to present additional problems in 
disorientation and altered breathing patterns. their own right. Most control systems will either increase fuel 
consumption or introduce new forms of emissions. However it 
 
is important to note that NOX emissions can only be obtained 
Toxicity by carefully considering air-fuel ratios, combustion and 
exhaust temperatures, design features that will reduce quench 
Nitrogen dioxide is a very toxic gas which irritates the entire 
zones, fuel injection timing, fuel injection rate, compression 
pulmonary system. Though little is known by what mechanism 
ratios, catalytic reductants and convertors, exhaust gas 
it causes toxicity, high concentrations have been known to 
recirculation systems, and advanced combustion control 
cause pulmonary oedema which is an abnormal accumulation 
strategies. No single method can handle the issue of NO  
of fluid in the pulmonary tissues [26]. In humans, oxides of X
emissions efficiently, but by combining two or more of these 
nitrogen interact with vitamin B12 leading to selective 
methods, the objective of reducing NOX emissions can be inhibition of methione synthase, with long exposure leading to 
achieved. The following sections will endeavour to discuss 
megaloblastic bone marrow depression and other neurological 
some of the methods that have been developed to mitigate the 
diseases [27]. 
emissions of NOX and show that through a multiple approach 
 this objective is achievable. 
Acid Rain   
The accumulation of oxides of nitrogen in the atmosphere Injection Timing 
increases acid deposition which tends to decrease ecosystem 
Injection timing is one of the most influential factors in control 
stability. Acid rain is caused by a chemical reaction which starts 
of oxides of nitrogen emissions, especially in lean burn diesel 
first with sulphur dioxide compounds reacting with oxides of 
engines, also known as direct injection compression ignition 
nitrogen that have been released to the stratosphere [28]. 
(DICI) engines, or their variant, indirect injection compression 
Sulphur dioxide and oxides of nitrogen react and dissolve in 
ignition (IDICI) engines. By retarding ignition or injection 
water and can be transported by wind. Once they mix and react 
timing, significant reduction of oxides of nitrogen emissions 
in the presence of moisture they form more pollutants with the 
can be achieved [30, 31] and as shown in Figure 3. 
result being acid rain [29]. Human activities, which release 
various chemicals into the air, are the main cause of acid.  It is observed from Figure 3 that advancing the injection timing 
or the SOI causes an increase in NOX emissions for all the test The major contributors and sources of sulphur dioxide (SO2) 
fuels covered by the study under multiple injection schemes. 
and oxides of nitrogen (NOX) are:  
This can be explained by the fact that ignition and combustion 
 Burning fossil fuels to generate power and electricity, then occur much earlier leading to early peak pressure as the 
producing two thirds of the sulphur dioxide and a piston approaches top dead centre (TDC), thus causing a rapid 
quarter of oxides of nitrogen in the atmosphere. increase in combustion temperatures which promotes NOX via 
 Heavy duty equipment and vehicles. the thermal mechanism also known as the Zeldovich 
mechanism. 
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© Research India Publications.  http://www.ripublication.com 
 
Figure 3: Effects and influence of start of injection (SOI) timings and injection timing strategies on NOX and smoke emission 
reduction for various diesel fuel blends compared to conventional diesel [31]. 
 
Injection Rate lower NOX emission level due to the fact that their retarding is 
greater than direct injection (DICI) engines with common 
The rate of injection determines the level and quantity of 
values for BTDC being 2⁰ and as late as 5⁰ ATDC without an 
emissions of oxides of nitrogen. This is due to an increase in 
effect on fuel economy. Continued retardation is observed to 
the mixing time or ignition delay during the injection period 
continue reducing NOX emissions but creates more 
which gives a very hot flame when combustion occurs, thus a 
hydrocarbon emissions and particulate matter (PM) emissions. 
greater quantity of NOX emission are released to the 
environment [32]. Research conducted so far points to the The NOX emissions in Figure 4 were observed to significantly 
successful reduction in the NOX emissions by use of this reduce with the combined effect of EGR, injection pressure and 
strategy [33, 34]. injection timing, with reduced injection rate showing an 
increase in the NOX emissions as compared to high injection 
pressure [35]. 
 
The Air-Fuel Ratio 
Any change to the air-fuel ratio mixture affects the combustion 
temperature which in turn affects the formation of NOX because 
it is a temperature dependent function as aforementioned. 
Therefore, an increase in air-fuel ratio achieved by decreasing 
the air amount in the mixture increases the amount of oxides of 
nitrogen. At higher fuel-air ratios the additional fuel has a 
tendency to cool the intake charge mixture temperatures, 
resulting in a drop in local peak temperatures within the 
  
combustion area, which decreases the NOX concentrated 
      emissions. However, IDICIs behave differently due to the 
Figure 4: NO  emissions at different engine loads of 25%, presence of a pre-chamber unit where the pre-chamber tends to X
50%, 75% and 100%, with application of EGR of 15% and operate in rich fuel situations and conditions, thus producing 
injection pressure rate of 240 bar, with two diesel blends of less NOX emissions [36, 37]. 
B20, B20-CE and conventional diesel at normal operating In diesel engines both air mass flow rate ṁ𝑎 and the fuel mass 
conditions.  flow rate ṁ𝑓 are measured parameters, with most modern diesel 
 engines running on an air-fuel ratio of lambda 𝜆 between 1.65 
to 1.10, with the lowest fuel consumption occurring at 𝜆 =
Considering the effects of oxides of nitrogen on the 1.65. This ratio of the flow rate is vital when defining engine 
environment, it is important to note from Figure 4 above that operating conditions [18] as shown in these two equations: 
diesel engines working to move goods and services normally 
operate at injection timings slightly retarded from the best fuel 𝐴 ṁ⁄𝐹 =
𝑎 (air/fuel)    Equation 19 
ṁ
economy timing. Indirect injection (IDICI) engines have a 𝑓
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© Research India Publications.  http://www.ripublication.com 
𝐹 ṁ𝑓⁄𝐴 =  (fuel/air)    Equation 20 
Increasing the compression ratio increases the power created 
ṁ𝑎 by the higher thermal efficiency that accrues due to the increase 
 in the compression ratio. However, it is noticed that a continued 
increase in the compression ratio increases NOX emissions by 
The Compression Ratio increasing the cycle temperatures [39]. 
The compression ratio is defined as the ratio of the volume of 4.5 The Catalysts 
the cylinder and of the combustion chamber of the cylinder 
head when the cylinder piston is at bottom dead centre (BDC) As a means of reducing NOX emissions in automobiles (both 
to the volume of the cylinder and of the combustion chamber heavy and light duty) and stationary engines, several kinds of 
of the cylinder head when the cylinder piston is at top dead catalytic converters have been developed to reduce the NOX 
centre (TDC), as shown on Figure 5 below. emissions in the presence of oxygen within the normal 
operating conditions of diesel engines, achieving some very 
positive results especially when used in tandem with other NOX 
control systems or in combination with themselves. Currently 
the most commonly used catalytic systems in diesel engines are 
selective catalyst reduction (SCR), diesel oxidation catalyst 
(DOC), and diesel particulate filter (DPF) 
Their main function is to oxidize HC, CO and NOX emissions 
through chemical reactions in the presence of oxygen thus 
rendering NOX emissions ineffective and low or eliminating 
them altogether, or, in the case of the diesel particulate filter, 
clean particulate matter emissions. Figure 7 below shows how 
a diesel oxidation filter works in schematic diagram form. 
 
Figure 5: Schematic diagram of an engine theoretical cylinder 
and the cylinder head showing the definition of the 
compression ratio. Where TC is top dead centre, BC is bottom 
dead centre, 𝑉𝑐 is clearance volume, B is the cylinder bore, L is 
the stroke and 𝑉𝑑 is the swept or the displaced volume.   
  
The compression ratio (𝑟𝑐) thus can be expressed in equation Figure 7: Schematic diagram of a diesel oxidation catalyst 
form as: (DOC) and its operation in reducing emissions of CO and HC 
𝑉 +𝑉
𝑟 = 𝑑 𝑐 through the process of oxidation. 𝑐      Equation 2 𝑉𝑐
 
Exhaust Gas Recirculation (EGR) 
EGR is a method of NOX control that has increasingly become 
very effective and is widely used in the control and reduction 
of NOX emissions in tandem with other NOX control systems. 
The system works by recirculating a portion of the exhaust 
gases into the combustion chamber for reburning together 
with the fresh intake charge. Since NOX is a temperature 
dependent function, EGR’s main function therefore is to 
reduce the combustion temperatures gases for a given mass of 
fuel and oxygen that is burnt in the combustion chamber. 
 However, while it is effective as a measure to control NOX 
Figure 6: Effects of the compression ratio on NOX formation, emissions, it has demerits because it tends to increase CO and 
at different engine loads PM emissions, thus limiting the extent to which it can be 
applied and recirculated, to about 30%. 
 
 
The increase in the compression ratio in Figure 6 shows that the 
combustion duration also decreases by almost 2⁰CA to 3⁰CA,  
due to the decrease in ignition delay which is a crucial factor of  
compression ratio in compression ignition engines [38]. 
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© Research India Publications.  http://www.ripublication.com 
 
Figure 8: Two types of EGR systems in use today, where if P3-P2>0 the delivery arrangement pipe is shorter hence the name short 
part (HPL)EGR, the long part arrangement (LPL)EGR is where the exhaust gases are taken from down stream of the compressor 
turbine and the pressure values P3-P2<0 [40]. 
 
CONCLUSION More resources need to be allocated for research and 
development, so that continued studies can be conducted on 
This work points out that human activities are the major source 
those methods of NOX control that have proved useful and 
of oxides of nitrogen, especially from power generation and 
efficient in dealing with the problem of emission, at the same 
from power propulsion for the transport industry. 
time affording young researchers an opportunity to carry out 
Although the effects of air pollution and environmental experiments and collect data that can be useful in development 
degradation on material, plants and animals can be measured of these methods and protecting the environment. 
and quantified, the effects of oxides of nitrogen emissions can 
 
only be estimated in humans from the epidemiological 
evidence which comes on later after exposure to high levels of ACKNOWLEDGEMENT 
NOX concentrations. 
First and far most to the good Lord for His enabling grace and 
The main significant oxides of nitrogen responsible for the mercies, to my supervisor Prof FN Inambao for his tireless 
pollution of the air, environmental degradation, human and efforts in molding me, to my brother David.K. Marwa for his 
animal health are nitric oxide and nitrogen dioxide. vision and effort and tireless support, to my family members 
especially my sisters Prof M.B Mwita and Dr I.N Marwa. 
All evidence available on the effects of NOX emissions Lastly, to the UKZN fraternity for the enabling environment, 
pollution so far provided indicate that high levels of NOX and my fellow post-graduate students who helped me to 
concentrations are a threat to human and animal health, besides 
complete this work.  
posing an immediate danger to the natural environment, 
therefore if measures are not implemented and attitudes  
changed, damage will continue and will increase.  
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3209 
CHAPTER SEVEN 
Influence of Exhaust Gas Recirculation and Emission 
Characteristics of a Diesel Engine Using Pyrolyzed Waste Plastic 
Biodiesel and Blends 
Journal article 
Maroa, S., & Inambao, F. (2018). Influence of Exhaust Gas Recirculation and Emission 
Characteristics of a Diesel Engine Using Pyrolyzed Waste Plastic Biodiesel and Blends 
International Journal of Applied Engineering Research, 13(10), 8321-8335. 
http://www.ripublication.com 
 
 
133 
 
International Journal of Applied Engineering Research ISSN 0973-4562 Volume 13, Number 10 (2018) pp. 8321-8335 
© Research India Publications.  http://www.ripublication.com 
Influence of Exhaust Gas Recirculation and Emission Characteristics of a 
Diesel Engine Using Pyrolyzed Waste Plastic Biodiesel and Blends 
 
2
Semakula Maroa1and Freddie Inambao  
1, 2 University of KwaZulu-Natal, Durban South-Africa. 
ssemakulamaroa@gmail.com; inambaof@ukzn.ac.za 
 
Abstract Keywords: Biodiesel, EGR Flow Rate, NOX, Smoke 
Emissions, pyrolysis, waste plastic oil 
Waste plastic pyrolysis biodiesel fuels have characteristics and 
qualities which are identical to conventional petroleum fuels.  
This work studied the influence of exhaust gas recirculation 
ABBREVIATIONS  
(EGR) on performance and emission characteristics of a diesel 
engine using biodiesel from Pyrolyzed waste plastics and RX  Molar Gas Ratio 
blends compared to conventional diesel. There are a limited 
n   Fuel Molar Quantity 
number of experimental works on waste plastic pyrolysis, f
especially the interaction, influence and effects of EGR on Al2O3,   Aluminium Tri-Oxide 
diesel engines. The results show that the amount of carbon 
monoxide (CO) emissions seemed to decrease at low engine ASTM   American Society for Testing and Materials 
loads up to intermediate loads (50 %), thereafter continued to BP  Brake Power  
increase significantly and marginally. Among the blends 
WPPOB100 reported the highest BSFC with a value of 0.4751 BSFC   Brake Specific Fuel Consumption 
g/kW.hr at 0 % EGR flow rate compared to 0.7235 g/kW.hr at BTE   Brake Thermal Efficiency 
30 % EGR flow rate. The increase in blend ratio showed a 
direct decrease in brake power (BP) in a linear relationship. At CaO   Calcium Oxide 
30 % engine load conventional diesel (CD), WPPOB10, CD  Conventional Diesel 
WPPOB20, WPPOB30 and WPPOB40 had values of 2.125 
kW, 2.15 kW, 2.05 kW, 1.98 kW, 1.86 kW and 1.75 kW CO   Carbon Monoxide 
respectively. Regarding exhaust gas temperature (EGT), the CO2  Carbon Dioxide  
WPPOB10 blend at 30 % EGR flow rate had the greatest 
reduction in temperature compared to the other WPPO blends Cu  Copper  
with a temperature value of 320 ⁰C. For unburnt hydrocarbon DEA  Department of Environmental Affairs  
(UHC) emissions, EGR flow rates of CD at 5 %, 10 %, 15 %, 
20 %, 25 % and 30 %, had 43 ppm, 57 ppm, 70 ppm, 82 ppm EGR  Exhaust Gas Recirculation 
and 85 ppm respectively, compared to WPPOB10 whose values EGT  Exhaust Gas Temperature  
were 23 ppm, 35 ppm, 40 ppm, 48 ppm, 50 ppm, and 52 ppm 
respectively. All the fuels tested indicated a drop in nitrogen FBP  Final Boiling Point 
oxide (NOX) emissions with an increase in the application of G-20   Group of 20 Highly Industrialized Countries 
EGR % flow rate, at all engine load conditions. The WPPO 
blends were observed to produce a continuous increase in G-7     Group of 7 Developed and Industrialized Countries 
smoke emission to almost double the values with the 
GC-MS  Gas Chromatography Mass Spectrometry 
application of EGR flow rate. At 10 % CO emission values 
were 9.79 %, 10.46 %, 10.91 %, 11.25 % and 12.75 % for GVC  Gross Calorific Value 
WPPO10, WPPO20, WPPO30, WPPO40 and WPPO100 
HPDE   High-Density Polyethylene 
respectively. The higher the blend ratio and the higher the 
increase in EGR % flow rates, the higher the carbon dioxide HRR  Heat Release Rate 
(CO2) emission values and vice versa. At 30 % EGR flow rate 
IBP  Indicated Brake Power  
the CO2 emissions for CD were 10.95 % compared to 
WPPOB10 9.95 %, WPPOB20 9.65 %, WPPOB30 8.85 % with SiO2  Silica Dioxide 
WPPOB100 showing the highest value of 14.35 %. Increased 
kW   Kilowatt 
blend ratio and EGR % flow rate showed a steady increase in 
the smoke emissions within the test fuels with marginal LPG  Liquid Petroleum Gas  
decreases observed across all the blended fuels. At 15 % EGR 
LTC  Low Temperature Combustion 
flow the values were 7.53 %, 7.1 %, 6.72 %, 6.25 %, 6.0 % and 
5.4 % for CD, WWPO10, WPPO20, WPPO30, WPPO40 and MSW  Municipal Solid Waste  
WPPO100 respectively. 
NaAlSi2O6-H2O  Zeolite 
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International Journal of Applied Engineering Research ISSN 0973-4562 Volume 13, Number 10 (2018) pp. 8321-8335 
© Research India Publications.  http://www.ripublication.com 
NOX  Oxides Nitrogen waste wreaks havoc on the environment because plastics are 
petroleum based which makes them non-biodegradable [5]. 
PEHD  Polyethylene High-Density 
Plastic waste accounts for between 8 % to 12 % of global waste 
PM  Particulate Matter with a projected increase of between 9 % to 13 % annually by 
the year 2025 [6] and [7]. According to a study conducted in 
PVC  Poly Vinyl Chloride 
India, nearly 5.6 million metric tonnes of plastic waste are 
UHC  Unburnt Hydrocarbons generated every year out of which only a paltry 10 % is 
recirculated, while 80 % goes into landfills which causes 
WPPO  Waste Plastic Pyrolysis Oil 
pollution. In South Africa 24 115 402 metric tonnes of general 
𝑥  Exhaust Molar Gas Quantity waste is produced annually, of which 6 % (1 446 924 metric 
tonnes) was plastic waste with a national average waste 
𝑦  Inlet Intake Molar Gas Quantity production annual increment projection of 2 % to 3 % since 
𝑧  Remainder of the Intake Charge 2008 [8], as shown in Figure 1. These figures support the case 
for using technology to degrade waste plastic mass into energy. 
 Techniques such as pyrolysis can yield hydrocarbons similar in 
INTRODUCTION quality and characteristics to petroleum fuels [9] and [10].  
The importance of modern day transport systems cannot be 
overstated, especially the transportation of goods and services 
and people. The propulsion provided by internal combustion 
engines with diesel fuel as the primary source of energy forms 
the bulk of commercial use and personal transport, owing to 
their numerous advantages of this fuel. Diesel engines are 
inherently lean burn engines, and emit relatively low carbon 
dioxide (CO2) emissions compared to petrol propelled internal 
combustion engines. Other advantages offered by diesel 
engines include high thermal efficiencies, durability and  
construction robustness [1]. These advantages mean that the 
Figure 1 Analysis of available general waste data (from 
use of such engines is increasing as more countries move into 
municipalities) in South Africa [8] 
urbanization and industrialization and catch up with the highly 
industrialized countries. However, there is now a challenge to  
phase them out based on environmental and human health 
Pyrolysis is a word originally coined from two Greek words 
issues due to the high levels of nitrogen oxide (NOX), smoke 
pyro “fire” and lysis “decomposition” [11]. It is a chemical 
and particulate matter (PM) emissions.  
decomposition process which makes fuel from plastic waste by 
There has been an increase in stringent emission regulations heating and has been commended by [12] as one of the 
enacted by global industrial powers such as the United States solutions to management of the menace of plastic waste. 
of America, the European Union and the G-7 and G-20. Diesel During the pyrolysis process, assorted waste plastics are 
engines have been identified as a major source of pollution due introduced and fed into a reactor and subjected to high 
to their emission of NOX gases, so now alternative fuels are temperatures of 400 ⁰ C to 600 ⁰ C and sometimes up to 900 
being sought in the interest of reducing energy consumption as ⁰ C at atmospheric pressure in the absence of oxygen for 3 to 4 
well as reducing environmental degradation. The rapid hours [13] to produce oil and other by-products. In order to 
expansion of the road transport industry sector is fast eroding maintain and sustain high temperatures during the pyrolysis 
all the technological developments and improvements thus far reaction, [14] catalysts are therefore employed, including 
achieved in the war against pollution from diesel propelled calcium oxide (CaO), silica dioxide (SiO2), aluminium tri-
engines. Issues relating to climatic change, uncontrolled or oxide (Al2O3) and zeolite (NaAlSi2O6-H2O) [15]. This process 
erratic energy prices, uncertainty of future fossil fuel supplies breaks down large molecules of waste into minute molecules 
[2] and unending internal conflicts and war in major oil resulting in hydrocarbons with a smaller molecular mass. For 
producing countries create a very compelling case for example, the addition of ethane enables fractional distillation 
alternative fuels.    to be applied and to obtain fuels, chemicals and by-products 
from the process. Pyrolysis as a process gives yields with a 
Due to the increase in food insecurity the production of 
weight factor of 75 % of liquid hydrocarbons in a mixture of 
biodiesel fuel from plant based stocks is a less viable option 
petrol, diesel and kerosene, 5 % to 6 % as residue coke and the 
because of its impact on higher food prices [3]. Therefore, 
remaining balance as liquefied petroleum gas (LPG) [16]. 
waste plastic from municipal solid waste management sites is 
Pyrolysis as a method of producing waste plastic biodiesel is 
increasingly being considered as an alternative source of fuel 
highly recommended by most researchers and commercial 
and energy due to the widespread availability of plastics and 
entities because of its cost effectiveness and its high energy 
their cost of disposal and negative effect on the environment. 
conversion rate, besides the high yield when compared to any 
There are two types of plastics widely used today, namely: PVC 
other method of plastic waste energy extraction [17]. 
(poly-vinyl chloride) and HPDE (High-density polyethylene) 
also known as polyethylene high-density (PEHD) [4]. Plastic 
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The use of biodiesel calls for NOX reduction techniques such as NOX, UHC and CO emissions. In order to stabilize them and 
exhaust gas recirculation (EGR) due to the oxygen content the performance of higher blend ratios, injection timing has 
inherent in most biodiesel fuels which is the single most been proposed without the need for upgrading the fuel, engine 
important factor responsible for NOX formation as it reacts with modification or fuel alteration through addition of additives as 
high temperature combustion mixtures thereby increasing the observed by [30]. Injection timing was seen to affect 
availability of NOX [7]. Diesel fuels and biodiesel fuels both performance from WPPO in the study by Sharma et al. [31], 
require fuel additives to improve engine lubricity, better where Jatropa blends of 20 % tyre oil and 80 % Jatropa ester 
ignition qualities and better mixing. Oxygenates in biodiesels oil resulting into lower fuel consumption, CO, UHC and PM 
are providing a promising future to reduce PM emissions since although it was observed to increase NOX emissions. [32] 
the O2 content aids in better combustion, besides lowering Observed an increase in the BTE and NOX emissions, thus 
exhaust emissions with a clear trade-off between PM and NOX concurring with the findings of Sharma et al. on emissions of 
as suggested in the findings of [18], [19] and [20]. Most NOX, although on fuel consumption, CO, and UHC they noted 
researchers suggest modifications, for example thermal barrier decreased results. This work will endeavour to study the 
coating by [21], where efficiency is improved, and NOx influence of EGR on the performance of a diesel engine using 
emissions and smoke density are reduced with a minimal a biodiesel fuel derived from pyrolysis of waste plastic, and 
increase in brake thermal efficiency (BTE), although fuel how WPPO interacts with engine performance and emission 
economy may drop. characteristics when EGR is applied. 
[22] Observed that with the application of EGR % flow rate a  
further reduction for both NOX and soot emission can be 
EXPERIMENTAL  
achieved when n-pentanol is added. [23] observed a 
simultaneous reduction for both NOX and soot emissions using Experimental Apparatus and Equipment 
a low temperature combustion (LTC) strategy, with application 
Figure 1 is a schematic diagram for the experimental engine 
of EGR % flow rates coupled with late injection timing and 
set-up and its components, which were used in the experiment. 
addition of n-pentanol to the blended diesel biodiesel tested. 
Figure 2 is ano5her schematic diagram showing the 
However when [24] added n-pentanol to the diesel biodiesel the 
modification of the EGR loop that was adapted for the 
brake specific fuel consumption (BSFC) increased with no 
experimental engine in Figure1. Table 1 is the experimental 
increase in BTE. This seems to confirm n-pentanol as a better 
engine specification standards of the experimental engine that 
fuel additive for waste plastic pyrolysis oil (WPPO) compared 
was used throughout the study 
to n-butanol due to its high cetane number, better blend ratio 
stability and less hygroscopic nature as observed by [25].   
Other methods that have been suggested include reduced 
ignition delay, which reduces combustion temperatures, thus 
aiding in the reduction of NOX as its formation is temperature 
dependent. The use of cetane improvers is also an alternative 
for reducing NOX as the poor cetane index of WPPO fuel blends 
leads to poor ignition quality, particularly when biodiesel fuels 
are used such as glycol ether which reduce PM, unburnt 
hydrocarbons (UHC) and CO emissions especially in common 
rail direct injection diesel engines. Cetane improvers also 
decrease cylinder pressure, ignition delay, heat release rate, and 
engine knock or noise [26]. The inclusion of n-pentanol in 
diesel biodiesel blends was observed by [27] to shorten 
combustion duration and increase the heat release rate (HRR), 
while significantly reducing the NOX, CO and UHC emissions.  
Fuel additives have also been suggested and used, with the most Figure 2. Experimental test engine test rig set-up diagram 
commonly used being diethyl ether. As an organic compound  
diethyl ether has a high cetane number and is capable of 
boosting ether compounds’ cetane numbers as an improver Key: 1. Cylinder pressure sensor, 2. EGR control valve, 3. EGR 
[28]. In their study Devaraj et al. found that when used as an cooler, 4. Injection Control Unit, 5. Exhaust gas exit, 6. Air 
additive diethyl ether reduces ignition delay, cylinder peak box, 7. Signal amplifier, 8. Gas analyser, 9. air flow meter,10. 
pressure, HRR, CO, CO2 and NOX with a trade off in BTE Data acquisition system, 11. Crank position sensor, 12. 
which was observed to be increased. [29] on the other hand Dynamometer, 13. Engine, 14. Air flow rate meter, 15. Cooling 
found diethyl ether to reduce the ignition delay period, UHC, water exit to the cooling tower, 16. Dynamometer drive 
and NOX whereas BTE seemed to increased, but [14] using coupling. 
WPPO fuel observed and reported ignition delay and higher 
HRR with diethyl ether.  
Diesel engines have been shown to run stably on most medium 
blended ratios of waste plastic oil, although they produce more 
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of 25.4 mm to 50 mm, suitable for the pyrolysis reactor. The 
waste material is then loaded into the pyrolysis reactor by 
means of an automatic feeding machine, although there is 
provision for manual loading in case there is a system failure. 
Before starting the pyrolysis, the airtight reactor door is locked. 
The system is started from a control panel by a machine 
operator with the preceding processes subsequently running 
automatically as indicated by the flow chart in Figure 3. The 
first stage is to heat up and dry the waste plastic materials and 
increase the reactor temperature to the required value of 400 
⁰ C to 500 ⁰ C. The heavy dense gas oil falls into the oil tank 
 while the light oils rise up into the condenser where they are 
cooled and taken back into the oil tank for collection. 
Figure 2. EGR schematic loop modification 
The liquefied gases in small quantities which cannot be 
 
converted into oil are recycled and collected by the recycling 
Key; 1. The direction of EGR gases, 2. The 𝑥 subscript system to be burnt as fuel gas or cleaned through the after-
representing the exhaust molar gas quantity, 3. The direction of treatment system with the smoke and flue emitted by the plant 
inlet gases fresh charge, 4. The 𝑧 subscript representing being released to the atmosphere after the removal of sulphurets 
remainder of the intake charge, 5. 𝑛𝑓 Which is the fuel molar and black carbon. After the pyrolysis process is over the system 
quantity, 6. The EGR valve, 7. The Rx molar gas ratio, 8. The requires a cooling period of 4 hours to 5 hours through natural 
subscript 𝑦 representing the inlet intake molar gas quantity, 9. cooling, but nitrogen and carbon dioxide gases can be used as 
The direction of the exhaust gases exit, 10. The engine unit  cooling agents to reduce the cooling waiting time. This enables 
the removal of the carbon black compound without 
 contamination and pollution to the environment. The final 
Table 1. Engine specifications, position value and type operation is the removal of steel and other metals obtained in 
the pyrolysis reactor plant, as this requires the opening of the 
Parameters Position value reactor door in preparation for the next batch of the pyrolysis 
Ignition Type 4 (Stroke) DICI process. 
Number of Cylinders 1  
Model TV 1 
Cooling Medium Water 1 14 
Manufacturer Kirloskar 
Revolutions Per Minute 1500 4 
Brake Power 3.5 Kw 2 5 
Cylinder Bore 87.5 mm 
Piston Stroke 110 mm 6 
3 7 
Compression Ratio 18.5:1 
Connecting-Rod Length 234 
8 
Engine Capacity 661cc 
11 9 
Dynamometer Make 234 
Injection Timing 23.4  ֯  bTDC 
Maximum Torque 28 N-M@1500 
13 12 10 
Injection Pressure 250 Bar 
 
 
Figure 3. The waste plastic pyrolysis oil processing plant 
Waste Plastic Preparation and Conversion Process flow chart. 
The waste plastic materials were acquired from a municipal  
waste management site in Durban. The waste plastic is first 
Key: 1. Pyrolysis reactor, 2. Carbon black discharge, 3. Carbon 
sorted in a section of the pyrolysis plant. The dust and other 
black deep processing, 4. Exhaust smoke discharge, 5. Gas 
fine waste from the sorted plastic waste materials is collected 
separator, 6. Smoke scrubber to take out colour and odour, 7. 
by the cyclone filter and disposed of through a vent with a 
Condenser, 8. Chimney, 9. Oil tank, 10. Synchronized gas 
particle size monitoring system within the plant. After the 
purification, 11. Synchronized gas recycling system, 12. Extra 
sorting and removal of unwanted materials and dust, the waste 
gas burning, 13. Heating furnace during operation, 14. Loading 
plastics are taken through a conveyor machine to a pressure 
of material section.  
cleaner for thorough cleaning, after which they are conveyed to 
a shredding machine which reduces them to the required size 
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Physical Properties of Waste Plastic Pyrolysis Oil (WPPO) Sample 
Table 2. The test fuel properties, their units of measurement, standard methods of testing and the values  
for conventional diesel in comparison to the values of waste plastic pyrolysis oil 
Property Unit CD WPPO STANDARD 
Appearance - Clear/brown Clear/amber Visual 
Density @ 20 kg/M3 838.8 788.9 ASTM D1298 
Kinematic viscosity @ 40 ⁰ C cSt 2.32 2.17 ASTM D445 
Flash point ⁰ C 56.0 20.0 ASTM D93 
Cetane index - 46 65 ASTM D4737 
Hydrogen  % 12.38 11.77 ASTM D7171 
Cu corrosion 3 hrs @ 100 - 1B ASTM D130 
⁰ C 
Carbon  % 74.99 79.60 ASTM D 7662 
Oxygen % 12.45 7.83 ASTM D5622 
Sulphur content  % < 0.0124 0.15 ASTM D4294 
IBP temperature ⁰ C 160 119 ASTM D86 
FBP temperature ⁰ C 353.5 353.5 ASTM D86 
Recovery %  98 - 
Residue and loss %  2.0 - 
Gross calorific value kJ/kg 44.84 40.15 ASTM D4868 
Table 2 shows the results of the physical properties of the gasoline and to some extent diesel oil in the tested samples. 
WPPO obtained through the pyrolysis process of the waste This leads to the observation and conclusion that it is possible 
plastics from municipal solid waste (MSW) management sites, for this oil to be used as a feedstock in future, if it is upgraded 
at optimized conditions and compared to the properties of CD into a lighter compound such as diesel fuel or any other liquid 
fuel oil. The appearance of the oil is yellowish in colour as fuel in the foreseeable future. 
shown by the photo in Figure 4 and the liquid distillate is free 
Table 3 shows the chemical composition of the WPPO from a 
of visible particulate sediments with a flash point of 20 ⁰ C, and 
pyrolysis plant process. 
a gross calorific value (GCV) OF 40.15 KJ/kg which almost 
compares to the range indicated by most petroleum fuels Table 3. Chemical composition of the WPPO obtained from a 
including CD, thus making it capable of giving the same GC-MS laboratory analysis report 
working performance results in internal combustion engines, 
especially diesel propelled ones. Molecular formula Percentage composition 
C10 66.32 
C10-C15 4.38 
C15-C20 12.66 
C20-C25 8.22 
C25-C30 8.42 
 
 
Figure 4. The distillate samples from the waste plastic Experimental Procedure 
pyrolysis oil samples The engine employed for this work was a Kirloskar 
 experimental variable compression engine, four stroke single 
cylinder, water cooled developing 3.75 kW of power at 1500 
The distillation report analysis shows the WPPO has an initial rpm. The schematic is provided and shown in Figure 1. The 
boiling point (IBP) of 119 ⁰ C to 353.5 ⁰ C, thus indicating technical specifications of the experimental variable 
some presence of other fuel oil components like kerosene, compression engine are shown and indicated in Table 1. A 
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dynamometer was used to provide the engine with load during Throughout this experiment blends will be referred to in this 
the experimentation, and to get measurements for the engine format with digits denoting the percentage blend ratio of WPPO 
intake air flow an air box was fitted to the engine intake by volume to CD supplied. To avoid experimental fuel from 
manifold system with a standard orifice mechanism. The fuel being contaminated from the previous test, each test was 
flow rate to the experimental engine was measured using a conducted after a thorough evacuation procedure was 
digital fuel gauge with a stop watch mounted to measure time conducted on the fuel lines and the fuel injection system 
taken for the fuel to be consumed.  mechanism of the test engine. This made it possible to conduct 
an experiment and collect good data and measurements with 
The exhaust gas temperatures were measured using mounted 
inputs from the test mode only, without fear of contamination 
temperature thermocouples of k-2 type, including the EGR 
and poor results from error. 
temperature which is measured before it mixes with the intake 
fresh air charge and the constituents of the combustion  
chamber, through the same k-2 thermocouples. A cylinder 
Analysis of Error and Percentage Inaccuracies 
pressure transducer was mounted on the engine cylinder head 
to monitor cylinder combustion pressure and collect data values This process was carried out for the purpose of performing and 
through a system charge amplifier connected to a computer data identifying the accuracy and precision of the measuring tools 
acquisition machine. The crankshaft position or the crank angle and instruments used in this experiment work, as errors can 
was monitored and measured through a mounted encoder near occur due to conditions outside of the experiment itself, such as 
the crankshaft pulley area. The emission gases during the calibration of the instruments, observational errors, 
experiment were monitored through a five gas exhaust gas manufacturing errors, errors associated with experimental set-
analyser, while to measure the smoke intensity an AVL 437C up and planning, besides environmental conditions existing 
smoke meter was employed. during the experiment [33]. The list of instruments used and 
their percentage error of analysis are provided in Table 4 
Since the engine develops maximum power at 1500 rpm and it 
together with the percentage inaccuracies of CO, CO2, UHC, 
is a variable compression engine, all the experiments were 
NOX, EGT and smoke opacity. 
conducted based on this nominal engine speed at part load and 
full load, but data could be obtained from different engine The percentage of error analyses are derived from the following 
loads. Part engine load is described as 50 % of engine load and formula, the root sum square method and expressed in equation 
full engine load is described as the engine running at 100 % form [1] as: 
load, with a fixed compression ratio of 18.5:1. The EGR system 
𝑛 2
is shown in Figure 3, modified to suit the experimental engine 𝑅 =  √∑𝑖=1 𝑋𝑖                                              Equation 1 
and enable data collection. The exhaust gases were tapped from Where R is the total uncertainty percentage, 𝑋𝑖 is the individual 
the exhaust pipe and joined to the intake manifold air intake uncertainty of all the calculated operating parameters, 𝑛 is the 
system via the air flow meter box through a manually 
total number of the parameters in the experiment and 𝑖 is the ith 
controlled gate valve which made it possible for the mixing of 
term of the computed parameters. The total percentage of the 
EGR gases and the fresh air intake. 
uncertainty is thus calculated based on Equation 2 as follows: 
The EGR % flow rate was divided into the following modes: 0 
2 2 2
%, 5 %. 10 %, 15 %, 20 %, 25 %, and 30 % spaced at intervals 𝑅 = √𝑋1 + 𝑋2 + 𝑋3 + ⋯ 𝑋
2
𝑖                           Equation 2 
of 5 %. The waste plastic pyrolysis oil fuel blends were Table 4 shows the instruments used for measurements and data 
prepared in the following percentage order and mixed with CD collection, their measuring range, accuracies and percentages 
fuel 10 %, 20 %, 30 %, 40 % and 100 %, where the WPPO10 of inaccuracies, as calculated from Equation 2.  
blend is 90 % CD fuel and 10 % WPPO fuel in that order. 
 
Table 4 Instruments used for measurements and data collection 
Instrument Accuracy Measuring Range Percentage inaccuracies 
AVL 437C smoke meter    
Smoke intensity ± 1 % 0-100 % ± 1 
AVL pressure  ± 0.01 bar 0-250 bar ± 0.01 
transducer GH14D 
AVL 365C Angle encoder ± 1⁰  - ± 0.2 
AVL Digas 444 (Five Gas Analyser)    
CO ± 0.03 % to  ± 5 % 0-10 % by vol ± 0.3 
CO2 ± 0.5 % to ± 5 % by vol 0-20 % vol ± 0.2 
O2 ± 5 % by vol 0-22 % by vol ± 0.3 
HC ± 0.1 % to  ± 5 % 0-20000 ppm by vol ± 0.2 
NOx  ± 10 % 0-5000 ppm by vol ± 0.2 
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Instrument Accuracy Measuring Range Percentage inaccuracies 
K-2 Thermocouple ± 1 ⁰ C 0-1250 ⁰ C ± 0.2 
Digital Stop Watch ± 0.2 s  ± 0.2 
Digital Fuel Gauge ± 1 mm  ± 2 
Burette ± 0.2 cc 1-30 cc ± 1.5 
 
3.0 RESULTS AND DISCUSSION observed by the findings of [35]. This scenario forces the 
engine to increase its fuel consumption in order to maintain 
3.1 Brake Specific Fuel Consumption (BSFC)  
constant speed and its increased load hence the increase in 
Figure 5 shows the variation of the BSFC under the effects of BSFC. 
EGR % flow rate with different fuel blends of WPPO and CD 
The WPPO biodiesel blends with EGR % flow rate applications 
operating at full engine load conditions. 
showed a better fuel economy, especially the lower blend ratios 
 of WPPOB10 and WPPOB20, compared to conventional diesel 
test fuels. However, as the EGR % flow rate increased there 
CD WPPOB10 WPPOB20 was a noticeable increase in the BSFC across all the test fuels 
used. At 0 % EGR CD was 0.4 g/kW.hr compared to 30 % EGR 
WPPOB30 WPPOB40 WPPOB100
flow rate which was 0.495 g/kW.hr. For the WPPO biodiesel 
0.8 WPPOB10 blend the value was 0.3225 g/kW.hr compared to 
0.5780 g/kW.hr at 30 % EGR flow rate. From Figure 5 it is 
0.7 evident that the test fuel that showed the highest BSFC among 
0.6 the blends of diesel and the CD test fuel was WPPOB100 which 
at 0 % EGR flow rate had a value of 0.4751 g/kW.hr compared 
0.5 to 0.7235 g/kW.hr at 30 % EGR % flow rate.  
0.4 An interesting observation during experimentation was that 
0.3 after the 10 % EGR flow rate the values for the BSFC across 
all the test fuel seemed to show and pick a lineal increment 
0.2 trend as can be shown in Figure 5 by the flattening of the graph 
0.1 curves and the close value trends. 
 
0
0 5 1 0 1 5 2 0 2 5 3 0 Brake Thermal Efficiency (BTE) 
BTE studies specifically help us to know the ability of the 
EGR %
 combustion system to accept the fuel provided and how 
Figure 5. Brake specific fuel consumption (BSFC) versus efficient it is in converting that fuel and turning it into a 
EGR % flow rate mechanical output, as observed by [28] and [27]. 
 Figure 6 shows the BTE % variations, under different blends of 
WPPO and conventional diesel fuel, with the application of 
As can be seen from Figure 5, the lower ratio blends of EGR % flow rate. 
WPPOB10 and WPPOB20 showed minimal reduction in BSFC 
at 0 % EGR flow rate compared to the values of CD and CD WPPOB10 WPPOB20
WPPOB100 which showed significantly higher values of 
BSFC at that mode. At 0 % EGR flow rate CD had a BSFC WPPOB30 WPPO40 WPPOB100
value of 0.4 g/kW.hr compared to WPPOB100 with a value of 30
0.4751 g/kW.hr, indicating that the fuel blend WPPOB100 had 25
a higher BFSC than CD and the other blends of WPPO at 20
0.3225 g/kW.hr, 0.3615 g/kW.hr, 0.3645 g/kW.hr and 0.3715 
15
g/kW.hr, for WPPOB10, WPPOB20, WPPOB30 and 
WPPOB40 respectively. 10
A similar trend was noted under the influence of EGR % flow 5
rate where for example at 20 % to 25 % EGR flow rate, the 0
BSFC showed increased tendencies. These findings concur 1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 0 1 0 0
with the findings of [34]. This phenomenon can be explained LOAD %
due to the effects of dilution of the fresh air intake as it mixes  
with exhaust gases that are being recirculated through the EGR Figure 6 brake thermal efficiency (BTE)  
system which leads to incomplete combustion of the inducted versus engine load % 
mixture, thus leading to a drop in power and engine torque as 
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BSFC ( g/kW.hr)
BTE % 
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Figure 7 shows the variation of BTE % with the application of The results in Figure 8 show that there was a lineal increase in 
EGR % flow rate using different blends of WPPO and CD. BP for all the test fuels with an increase in the engine load. CD 
had the highest increase in BP values compared to the blended 
CD WPPOB10 WPPOB20 fuels of WPPO. At 20 % engine load, the BP value for CD was 
1.45 kW while WPPOB10 had a value of 1.350 kW 
WPPOB30 WPPOB40 WPPOB100 representing a difference of 6.8 % in BTE when the two fuels 
are compared. 
14
12 The blended fuels in Figure 8 also show very close increments 
10 in BP with an increase in engine load conditions and a decrease 
8 in BP with an increase in the blend ratio for all the blended fuels 
6 tested. The increase in blend ratio showed a direct decrease in 
4 BP in a linearly incremental relationship. For example, at 30 % 
2 engine load CD, WPPOB10, WPPOB20, WPPOB30 and 
0 WPPOB40 had values of 2.125 kW, 2.15 kW, 2.05 kW, 1.98 
0 5 1 0 1 5 2 0 2 5 3 0
kW, 1.86 kW and 1.75 kW respectively, showing a decrease in 
EGR (%) the value of the BP throughout the experimentation period. The 
 WPPOB100 blend showed the lowest values for the BP 
Figure 7. Brake thermal efficiency (BTE) % versus EGR % compared to the blends of WPPOB10, WPPOB20, WPPOB30 
flow rate and WPPOB40 used in this study. These findings concur with 
the findings by [19] in relation to WPPO blends. 
 
The application of EGR % flow rate was observed to cause no 
From Figure 7 it is evident that there was a noticeable decrease 
significant change in BP. However, there was a negligible drop 
in the BTE especially with all high ratio fuel blends of 
in the engine brake power with the influence of EGR flow rate 
WPPOB40 and WPPOB100 compared to CD. In comparison 
except for the blend WPPOB10 which had almost identical 
to other WPPO blends WPPOB100 obtained the lowest 
value to CD as the curve of the two fuels indicate in Figure 8. 
decrease of BTE with a value of 7.05 % at 10 % EGR flow rate 
This leads to the conclusion that the blends of WPPO have 
and the least value of 2.35 % at 30 % EGR flow rate.  
identical BP values with CD. 
Although there was a reduction in BTE due to the application 
 
of EGR % flow rate as shown in Figure 7, the trends of 
decreased BTE continued to be observed. For example, at 0 % Exhaust Gas Temperature (EGT) 
EGR flow rate, the value of BTE for CD was 12.15 % compared 
Figure 9 shows the variation of exhaust gas temperature under 
to WPPOB10 and WPPOB20 with values of 13.25 % and 13.05 
different types of WPPO blend and CD with application of 
%. The WPPOB100 blend had the lowest value for BTE for all 
EGR % flow rate. 
EGR rate flow modes than any other test fuel.  
 CD WWPOB10 WPPOB20
Brake Power (BP) WPPOB30 WPPOB40 WPPOB100
Figure 8 shows engine brake power (BP) variations with EGR 600
% flow rate application with CD and different WPPO blends at 
full engine load conditions. 400
CD WPPOB10 WPPOB20 200
WPPOB30 WPPOB40 WPPOB100 0
0 5 1 0 1 5 2 0 2 5 3 0
5
EGR (%)
4
 
3
Figure 9. Exhaust gas temperature (EGT ⁰  C) versus EGR % 
2
flow rate 
1
 
0
1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 0 1 0 0 Temperature is one of the key factors in determining the 
LOAD (%) formation of engine exhaust emissions, besides providing or 
helping in the analysis and study of combustion processes in 
 
relation to fuel as observed by [36]. 
Figure 8. Engine brake power (BP) versus varying engine 
load % It is evident from the results in Figure 9 that there was a 
variation in EGT of WPPO and CD with the application of EGR 
 
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BP (KW) BTE (%)
EGT (⁰ C)
International Journal of Applied Engineering Research ISSN 0973-4562 Volume 13, Number 10 (2018) pp. 8321-8335 
© Research India Publications.  http://www.ripublication.com 
% flow rate. The results indicate that EGT decreased with 
different blends of WPPO compared to CD. The temperature CD WPPO10 WPPO20
difference between them was that WPPO blends had higher WPPO30 WPPO40 WPPO100
temperature increases in all the test conditions compared to CD. 
However, it should be mentioned that as the blend ratio 100
increased with EGR % flow rate application, the EGT reduced 80
significantly and marginally especially for WPPOB30 and 60
WPPOB40. The highest temperature value obtained for CD 40
was 456 ⁰ C at 0 % EGR flow rate, whereas the highest 
20
temperature value for WPPO blend was 490 ⁰ C obtained from 
WPPOB100 at 0 % EGR flow rate, although at 30 % EGR flow 0
rate this blend had the most reduction in temperature compared 0 5 10 15 20 25 30
to the other WPPO blends with a temperature value of 320 ⁰ C.  egr (%)
The application of EGR % flow rate in increasing modes  
brought further reduction in EGT with the highest value for CD Figure 10. Unburnt hydrocarbons (UHC) emissions versus 
obtained being 440 ⁰ C with an EGR % flow rate of 5 % while EGR % flow rate 
the minimum or lowest value of 340 ⁰ C was obtained at 30 % 
EGR flow rate. The WPPO blends showed a similar trend with  
decreasing temperatures with the application of EGR % flow As can be seen from Figure 10, when the EGR % flow rate 
rate with WPPOB10 showing the highest value of 467 ⁰ C and application was 0 %, there was no application effect. Figure 10 
lowest of 362 ⁰ C at 5 % and 30 % EGR rate flow respectively, shows that there were less UHC emissions for all the test fuels 
while WPPOB40 showed a highest value of 472 ⁰ C and a applied in this experiment, reporting values of 22 ppm, 23 ppm, 
lowest of 330 ⁰ C at 5 % and 30 % EGR flow rate respectively. 21 ppm, 20 ppm, 19 ppm and 17 ppm for WPPOB10, 
The cause of the reduction in EGT can be attributed to several WPPOB20, WPPOB30, WPPOB40 and WPPOB100 
respectively, compared to 20 % EGR percentage flow rates 
factors. The reduction in EGT among the different blends of 
with 77 ppm, 68 ppm, 52 ppm, 46 ppm, 44 ppm and 40 ppm 
WPPO is a result of low calorific value of the blends and the 
low exhaust loss, which concurs with the findings of [37]and respectively. 
[38]. The WPPO has a calorific value of 40.15 kJ/kg compared Although the application of EGR % flow rate reduces the 
to the calorific value of CD at 44.84 kJ/kg as shown in Table 2. amount of UHC emissions remitted by the applied test fuels 
Other causes are directly linked to the effects of EGR rate flow, across the board, CD fuel produced more UHC emissions from 
the dilution effect, chemical effects and thermal effects [39] the test engine compared to all the WPPO blends tested. Figure 
and [40].  10 shows that at EGR flow rates of 5 %, 10 %, 15 %, 20 %, 25 
% and 30 %, CD UHC emissions were 43 ppm, 57 ppm, 70 
 
ppm, 82 ppm and 85 ppm respectively, compared to WPPOB10 
Unburnt Hydrocarbon (UHC) Emissions whose values were 23 ppm, 35 ppm, 40 ppm, 48 ppm, 50 ppm, 
and 52 ppm respectively.  
Figure 10 shows the variation of UHC emissions in parts per 
million under full engine load with the application of various Therefore, the application of EGR % flow rate increased UHC 
EGR % flow rates, and different blends of WPPO and CD. The emissions as observed from the results and values presented 
UHC emissions were significantly higher across all the blended and obtained in Figure 10, with CD fuel producing the highest 
test fuels of WPPO, especially with higher engine load UHC emission values compared to all the test fuel blends of 
conditions as indicated by the values shown in Figure 10. WPPOB10, WPPOB20, WPPOB30, WPPOB40 and 
However, CD produced more and higher values of UHC WPPOB100. 
emissions compared to all the blends of WPPO across all the 
engine loading conditions and operating modes.  
Figure 10 shows UHC emissions (ppm) values under engine Nitrogen Oxide (NOX) Emissions 
load with different fuel blends of WPPO and CD under the Figure 11 shows variations of NOX emissions (ppm) and engine 
effects of EGR % flow rate application. loads with different blends of WPPO and CD fuel with EGR % 
 flow rate application. 
8329 
uhc (ppm) 
International Journal of Applied Engineering Research ISSN 0973-4562 Volume 13, Number 10 (2018) pp. 8321-8335 
© Research India Publications.  http://www.ripublication.com 
During study it was observed that at engine part load, as shown 
CD WPPO10 WPPO20 in Figure 12, the values for NOX emissions for all the fuels were 
WPPO30 WPPO40 WPPO100 lower compared to the values at full load engine conditions. 
The NOX emissions for CD at engine part load (50 %) was 635 
1400 ppm compared to full load at 1100 ppm, whereas the value for 
1200 WPPOB100 at engine part load (50 %) was 850 ppm compared 
to 1250 ppm at full engine load. This seems to indicate a 
1000 concurrence that at part engine load (50 %) the values of NOX 
800 emissions emitted by all the blends of WPPO except WPPO100 
were lower compared to the values at full engine load 
600 conditions.  
400  
200 Carbon Monoxide (CO) Emission 
0 Figure 13 shows CO emissions % variations versus varying 
0 5 1 0 1 5 2 0 2 5 3 0 engine load with different fuel blends of WPPO and CD test 
EGR (%) fuel with EGR % flow rate application. 
 
CD WPPO10 WPPO20
Figure 11. Variations of NOX emissions (ppm) versus EGR % 
flow rate WPPO30 WPPO40 WPPO100
 6
NOX emission formation is highly dependent on in-cylinder 5
temperature, the concentration of oxygen, and the residence 
time the fuel-air mixture spends in the combustion chamber 4
during the pre-mixing phase [41]. All the fuels tested indicated 3
a drop in the NOX emissions with an increase in the application 
of EGR % flow rate, at all engine load conditions. Figure 11 2
shows this was due to the rise in the total heat capacity of the 
1
working gases that increased with increasing EGR % flow rate, 
thus concurring with the studies and findings of [42], [43] and 0
[44]. 0 2 0 4 0 6 0 8 0 1 0 0 1 2 0
The NOX emissions value for the CD in Figure 11 at full load LOAD (%)
was 920 ppm with application of EGR % flow rate, while  
WPPOB100 was 1270 ppm, compared to the reduced values 
Figure 13. Carbon monoxide (CO) emissions % versus 
with application of EGR % flow rate of 30 % in Figure 11, 
varying engine load 
where CD NOX emissions decreased to 432 ppm compared to 
536 ppm before application of EGR % flow rate.  
Figure 12 shows variations of NOX emissions (ppm) under CO is a toxic gas that requires substantial control to acceptable 
varying engine load with different blends of WPPO and CD test levels. It is caused by poor combustion of hydrocarbon fuels as 
fuel without EGR % flow rate application. a result of dependency on the air-fuel ratio relative to the 
stoichiometric proportions [42].  
CD WPPOB10 WPPOB20 For all the test fuels the amount of CO emissions seemed to 
WPPOB30 WPPOB40 WPPOB100 decrease at lower engine loads up to part load percentages or 
intermediate loads of (50 %), after which the CO emissions 
1500 continued to increase significantly and marginally as in Figure 
1000 13. At 0 % engine load the value of conventional diesel was 
0.051 % compared to 50 % engine load when the value was 
500 reduced to 0.03 % by volume. However, as the engine load 
increased from 50 % there was a significant continuous and 
0
marginal increase in the percentage of carbon emissions by 
0 2 0 4 0 6 0 8 0 1 0 0 1 2 0
volume across all the test fuels irrespective of the EGR % flow 
LOAD (%) rate. For example, at 80 % engine load the value for 
 WPPOB100 was 0.02 % up from 0.0165 % by volume. The 
other WPPO biodiesel blends also showed a similar trend and 
Figure 12. Variations of NOX emissions (ppm) versus varying 
concurrency. WPPOB20 and WPPO30 test fuels at 50 % 
engine load % without application of EGR flow rate 
engine load condition had values of 2.25 % and 2.15 % 
8330 
NOX (PPM) NOX (PPM)
CO (%)
International Journal of Applied Engineering Research ISSN 0973-4562 Volume 13, Number 10 (2018) pp. 8321-8335 
© Research India Publications.  http://www.ripublication.com 
compared to 3.36 % and 2.95 % respectively at 80 % engine 
CD WPPO10 WPPO20
load. 14 WPPO30 WPPO40 WPPO100
Figure 14 shows the variation of CO emissions with EGR % 12
flow rate application under different blends of WPPO and CD.  
10
CD WPPO10 WPPO20 8
WPPO30 WPPO40 WPPO100 6
70 4
60
2
50
40 0
30 1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 0 1 0 0
20 LOAD (%)
 
10
0 Figure 15. Variation of carbon dioxide (CO2) % emissions 
0 5 1 0 1 5 2 0 2 5 3 0 versus engine Load %,  with different types of fuel blends of 
WPPO and CD 
EGR (%)
  
Figure 14. Carbon monoxide (CO) versus EGR % flow rate CO2 is the principal composition of EGR gases, and is a core 
application indicator of the quality of combustion and the temperature [6] 
of combustion within the combustion chamber. CO2 gas has a 
 
higher heat capacity which makes it act like a thermal heat sink 
As shown in Figure 14, the WPPO blends are observed to especially during the combustion process, making it possible to 
produce a continuous increase in smoke emissions to almost reduce peak cylinder temperatures, hence is of particular 
double the values with the application of EGR % flow rate as importance in the process leading to the reduction in NOX 
in Figure 14. For example, at 10 % EGR flow rate, CO emission emissions. 
values were 9.79 %, 10.46 %, 10.91 %, 11.25 % and 12.75 % 
Without EGR % flow rate and at lower engine loads, the value 
for WPPO10, WPPO20, WPPO30, WPPO40 and WPPO100 
of CO2 was considerably high for all the test fuels used. For 
respectively. CD reported the lowest carbon emissions with the 
example, at 20 % engine load WPPOB100 had the highest CO2 
application of EGR % flow rate with a value of 7.65 %.  
emission value of 4.65 % compared to the other test fuels used, 
There was also a correlation between the blend ratio and EGR namely CD, WPPO10, WPPO20, WPPO30 and WPPO40 with 
% flow rate on the amount of CO emissions produced. During 3%, 2.50 %, 1.5 %, 1.85 % respectively, as shown in Figure 15. 
experimentation it was observed that as the blend ratio 
However, it can also be observed from Figure 15 that the 
increased the CO emissions increased as the EGR % flow rate 
amount of CO2 increased with an increase in the engine load. 
increased as seen in Figure 10. At 20 % EGR flow rate CO 
As the engine load increased to 40 % the value of WPPOB40 
emission values were 18.25 %, 21.35 %, 22.65 %, 24.55 %, 
was 2.75 % compared to WPPOB30 at 3.25 %, while at 70 % 
26.95 % and 28.85 % respectively for CD, WPPO10, WPPO20, 
engine load the values 4.5 % and 5.25 % respectively. This 
WPPO30, WPPO40 and WPPO100. However, WPPO30 
leads us to the observation that as the engine load is increased 
reported values of 4.85 %, 7.28 %, 10.91 %, 16.37 %, 24.55 %, 
with an increase in the blend ratios, lower ratio blends are 
35.75 % and 52.69 % as the EGR flow rate increased from 0 % 
observed to emit more CO2 compared to higher ratio blends 
to 30 % respectively. This may be due to the effects of EGR % 
except blend WPPOB100 which released more CO2 than any 
flow rate application and for dilution, thermal and chemical 
test fuel as mentioned earlier. At full engine load the value of 
reasons some of the oxygen in the inlet charge was replaced 
CO2 emissions were at their highest values across all the test 
with recirculated exhaust gas which caused incomplete 
fuels. The values were 12.75 %, 10.85 %, 9.65 %, 8.75 %, 8.35 
combustion.  
% and 8 % for WPPOB100, CD, WPPOB10, WPPOB20, 
 WPPOB30 and WPPOB40 respectively.   
Carbon Dioxide (CO2) Emissions  Figure 16 shows CO2 % variations with different blends of 
WPPO and conventional diesel test fuel under engine load, with 
Figure 15 shows the CO2 % variations under varying engine application of EGR % flow rate. 
load with different blends of WPPO and CD test fuels.  
 
8331 
CO (%)
CO2 (%)
International Journal of Applied Engineering Research ISSN 0973-4562 Volume 13, Number 10 (2018) pp. 8321-8335 
© Research India Publications.  http://www.ripublication.com 
CD WPPOB10 WPPOB20
CD WPPOB10 WPPOB20
WPPOB30 WPPOB40 WPPOB100
16 WPPOB30 WPPOB40 WPPOB100
14 1.2
12 1
10 0.8
8
0.6
6
0.4
4
0.2
2
0 0 EGR(%)
0 5 1 0 1 5 2 0 2 5 3 0 0 5 1 0 1 5 2 0 2 5 3 0
 
EGR (%)
 Figure 17. Variation of smoke emissions or opacity % versus  
Figure 16. Variations of carbon dioxide (CO2) % versus EGR EGR % flow rate, with different WPPO blends and CD 
% flow rate, with different types of WPPO fuel blends and  
CD 
The steady increase in smoke emissions can be explained by 
 the fact that WPPO blends of fuel have a high kinematic of 
The application of EGR % flow rate increased the CO  viscosity compared to CD fuel, and they have low volatility 2
emissions exponentially by almost doubling the values as can values. Other possible explanations for this phenomenon are 
be seen in Figure 16. For example, at 10 % EGR flow rate the the poor injection and spray characteristics observed for WPPO 
value of CD was 5.35 % compared to WPPOB100 at 7.25 %, blends of fuels compared to those of CD fuel which has better 
WPPOB10 was at 4.75 %, WPPOB20 at 4.25 %, WPPOB30 at spray qualities. Another likely cause is associated with the high 
3.95 %, and WPPOB40 at 3.65 % respectively. These results aromatic compounds found in most WPPO blends of fuel 
reinforce the observation that CO  emissions with the compared to CD fuel. 2
application of EGR % flow rate have a correlation with The application of EGR % flow rate showed that there was a 
blending i.e., the lower the blend ratio, the higher the emission significant increase in the values of smoke emissions and 
values and vice versa. The values for 30 % EGR flow rate for particulate matter emissions across all the test fuels. Smoke 
all the test fuels contained high emissions for CO2, for example emissions with WPPOB10 blend were 7.2 % lower compared 
CD was 10.95 %, WPPOB10 9.95 %, WPPOB20 9.65 %, to CD at 0 % EGR flow rate, with CD being 11.5 % higher than 
WPPOB30 8.85 %, and WPPOB100 had the highest value of WPPOB100 blend fuel at 30 % EGR flow rate. This result 
14.35 %.  seems to concur with the study findings of [46].   
 The WPPOB10 blend used in this study emitted the highest 
Smoke Emissions (Opacity) levels of smoke emissions of the blended fuels followed by 
WPPOB100. However, it should be mentioned here that as the 
This can be defined as the solid hydrocarbon soot particles that blend ratio and the EGR % flow rate increased there was a 
are found in exhaust system gases [45] and are directly linked steady increase in the smoke emissions across all the test fuels 
to smoke emissions formation. For all the blends of WPPO compared to when the EGR % flow rate was at 0 %. 
there was a noted increase in the level of smoke emissions, 
although the levels and values were considerably lower  
compared to CD. CONCLUSION  
Figure 17 shows smoke emissions or opacity % variations at  As the percentage of the blends of WPPO increased 
full engine load, with different blends of WPPO and CD under there was a marked decrease in the engine BP of the 
the effect of EGR % flow rate application. blended fuels compared to CD. This is true 
considering that the energy content (Table 2) for the 
WPPO test fuel was lower compared to the CD test 
fuel. 
 Peak power produced using WPPO biodiesel blends 
failed to match the peak power produced by CD, being 
in the region of 5% to 8% less compared to the peak 
power produced from the CD test fuel. 
 Data presented in this work provides more reasons to 
support the widespread use of WPPO as an alternative 
fuel for all types of compression ignition engines with 
8332 
CO2 (%)
SMOKE OPACITY (%)
International Journal of Applied Engineering Research ISSN 0973-4562 Volume 13, Number 10 (2018) pp. 8321-8335 
© Research India Publications.  http://www.ripublication.com 
or without modifications to the engines, especially mould me, to my brothers D.K. Marwa, Dr J.M. Marwa and 
when blends of WPPOB10 and WPPOB20 are used F.S. Marwa for their moral, financial, visionary leadership and 
because in this study their peak BP and BSFC values support. Lastly to the UKZN and staff fraternity for the suitable 
were identical to CD. environment, especially the LAN manager Shaun, the lab 
 Results (Figure 12) indicate that less NOX emissions technicians and my fellow post-graduate students, who enabled 
occurred during part and intermediate engine load for me to complete this work. 
the WPPO test fuels and CD test fuel compared to full 
 
engine load. 
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8335 
CHAPTER EIGHT 
THE EXPERIMENTAL SET-UP AND EQUIPMENT 
8.1 Introduction  
As discussed in section 2.1 in the literature review, one of the grey areas is the paucity of 
literature related to plastic pyrolysis oil biodiesel. This work is motivated by the lack of 
experimental work on waste plastic pyrolysis oil biodiesel in relation to EGR. Considering this 
trend and the level of experimental work thus far achieved, it is important that an experimental 
work be attempted to increase and deepen understanding in this area. Using a mathematical 
model and an experimental approach was deemed to be the best way to achieve the study aims 
and objectives as set forth to investigate the effects of exhaust gas recirculation (EGR) on the 
performance parameters of a diesel engine. 
8.2 Experimental Apparatus and Equipment 
The engine experimental set up is shown on Figure 8.1 followed by its position names. 
  
Figure 8.1: Experimental test engine schematic set-up diagram 
Figure 8.1 shows the schematic diagram of the experimental engine test rig set up and its 
nomenclature. 1. Cylinder pressure sensor, 2. EGR control valve, 3. EGR cooler, 4. Injection 
Control Unit, 5. Exhaust gas exit, 6. Air box, 7. Signal amplifier, 8. Gas analyzer, 9. air flow 
meter,10. Data acquisition system, 11. Crank position sensor, 12. Dynamometer, 13. Engine, 
14. Air flow rate meter, 15. Cooling water exit to the cooling tower, 16. Dynamometer drive 
coupling. 
149 
 
Figure 8.2: schematic of the lab set-up of the fabricated and modified exhaust recirculation mechanism to 
conform to the engine experimental set-up in Figure 8.1 
Figure 8.2 nomenclature. 1. Engine, 2. EGR cooler system, 3. EGR control valve, 4. Inlet intake 
manifold, 5. Exhaust manifold. 
Table 8.1: The engine specifications, position value and their type 
Parameter Position value 
Ignition Type 4 (Stroke) 
Number of Cylinders DICI 1 
Model TV 1 
Cooling Medium Water 
Manufacturer Kirloskar 1500 
Revolutions per Minute 3.5 Kw 
Brake Power 87.5 mm 
Cylinder Bore 110 mm
Piston Stroke 18.5:1 
Compression Ratio 234 
Connecting-Rod Length 661cc 
Engine Capacity 234 
Dynamometer Make 23.4֯ bTDC 28⁰ 
Injection Timing N-M@1500 
Maximum Torque  250 Bar 
Injection Pressure 
150 
8.3 Waste Plastic Preparation and the Conversion Process 
a. The waste plastic materials were acquired from a municipal waste management site and 
offloaded into a waste plastic sorting section of the pyrolysis plant. The dust and other fine 
wastes collected from the sorted plastic waste materials were collected by the cyclone filter 
and disposed of through a vent with a particle size monitoring system within the plant. 
b. After the sorting and removal of unwanted materials and dust, the waste plastics were taken 
through a conveyor machine to a pressure cleaner for thorough cleaning, as they were being 
conveyed to the shredding machine which reduced them to the required size of 25.4 mm to 50 
mm for the pyrolysis reactor. 
c. The waste material was then loaded into the pyrolysis reactor through an automatic feeding 
machine, although there was provision for manual loading in case there is system failure. 
Thereafter the reactor door was air tight locked ready to begin the pyrolysis process. 
d. The system was started from a control panel by a machine operator, with the preceding 
processes subsequently running automatically as indicated by the flow chart in Figure 8.3.  
e. The first stage was to heat up and dry the waste plastic materials and increase the reactors 
temperature to the required value of 400 ⁰C to 500 ⁰C. The heavy dense gas oil fell into the oil 
tank while the light oils rose up into the condenser where they were cooled and taken back into 
the oil tank for collection. 
f. The small quantity of liquefied gases which could not be converted into oil were recycled by 
the recycling system to be burned as fuel gas or cleaned through the after-treatment system 
with smoke and flue emitted by the plant released to the atmosphere after the removal of 
sulphurets and black carbon. 
g. After the pyrolysis process was over the system required a cooling period of 4 hours to 5 
hours through natural cooling, but faster cooling to reduce cooling waiting time can be done 
using nitrogen and carbon dioxide gases as cooling agents. This enables the removal of the 
carbon black compound without contamination and pollution to the environment. 
h. The final operation was the removal of steel and other metals obtained in the pyrolysis 
reactor plant, as this required the opening of the reactor door in preparation for the next batch 
of the pyrolysis process. 
 
151 
 
Figure 8.3 illustrates the waste plastic pyrolysis oil processing plant flow chart. 
 
1 14 
 
4 
2 5 
 
 6 
 3 7 
 8 
11 9 
 
 13 12 10 
Figure 8.3: Pyrolysis flow chart 
 
The nomenclature of the pyrolysis plant flow chart illustrated in Figure 7.3 is as follows: 1. 
Pyrolysis reactor, 2. Carbon black discharge, 3. Carbon black deep processing, 4. Exhaust 
smoke discharge, 5. Gas separator, 6. Smoke scrubber to take out color and odor, 7. Condenser, 
8. Chimney, 9. Oil tank, 10. Synchronized gas purification, 11. Synchronized gas recycling 
system, 12. Extra gas burning, 13. Heating furnace during operation, 14. Loading of material.  
8.4 Physical Properties of Waste Plastic Pyrolysis Oil (WPPO) Sample 
Table 8.2 shows the results of the physical properties of the waste plastic pyrolysis oil obtained 
through the pyrolysis process of the waste plastics from municipal solid waste (MSW) 
management sites, at optimized conditions and compared to the properties of conventional 
diesel fuel oil.  
The appearance of the oil was yellowish in color as shown by photograph in Figure 8.4. The 
liquid distillate is free of visible particulate sediments with a flash point of 20 ⁰C, and a gross 
calorific value (GCV) of 40.15 KJ/kg which almost compares to the range indicated by most 
petroleum fuels including conventional diesel, thus making it capable of giving the same 
working performance in internal combustion engines, especially diesel propelled ones. 
 
 
152 
 
Table 8.2: The test fuel properties, their units of measurement, standard methods of testing and the 
values for conventional diesel in comparison to the values of waste plastic pyrolysis oil 
       
Property Unit CD WPPO ASTM 
Standard 
Appearance - Clear/brown Clear/amber Visual 
Density @20 kg/M3 838.8 788.9 ASTM D1298 
Kinematic viscosity cSt 2.32 2.17 ASTM D445 
@40⁰C 
Flash point ⁰C 56.0 20.0 ASTM D93 
Cetane index - 46 65 ASTM D4737 
Hydrogen  % 12.38 11.77 ASTM D7171 
Cu corrosion 3hrs@100⁰C - 1B ASTM D130 
Carbon  % 74.99 79.60 ASTM D 7662 
Oxygen % 12.45 7.83 ASTM D5622 
Sulphur content  % <0.0124 0.15 ASTM D4294 
IBP temperature ⁰C 160 119 ASTM D86 
FBP temperature ⁰C 353.5 353.5 ASTM D86 
Recovery %  98 - 
Residue and loss %  2.0 - 
Gross calorific value kJ/kg 44.84 40.15 ASTM D4868 
 
 
Figure 8.4: The distillate samples from the waste plastic pyrolysis oil samples 
 
The distillation report analysis (Table 8.3) shows that the WPPO had an initial boiling point 
(IBP) of 119 ⁰C to 353.5 ⁰C, thus indicating some presence of other fuel oil components like 
kerosene, gasoline and to some extent diesel oil in the tested samples. This leads to the 
observation and conclusion that it is possible for this oil to be used as feedstock in future if it 
is upgraded into a lighter compound such as diesel fuel or any liquid.  
153 
 
Table 8.3: The chemical composition of the waste plastic oil from a pyrolysis plant process: the carbon 
chains and percentage compositions in the test fuel, obtained from a GC-MS laboratory analysis report 
Molecular formula Percentage composition  
C10 66.32 
C10-C15 4.38  
C15-C20 12.66  
C20-C25 8.22 
C25-C30 8.42  
 
8.5 Experimental Procedure 
a. The engine that was employed for this work was a Kirloskar experimental variable 
compression engine, four stroke single cylinder, water cooled developing 3.725 kW of power 
at 1500 rpm. The schematic of the experimental set-up being as in Figure 8.1. 
b. The technical specifications of the experimental variable compression engine are shown and 
indicated in Table 8.1. A dynamometer was used to provide the engine with load during the 
experimentation. To get measurements for the EGR system and engine intake air flow, an air 
box was fitted to the engine intake manifold system with a standard orifice mechanism, and a 
fabricated EGR valve and cooler exchanger as provided for in Figure 8.3 and by the equations 
on orifice systems as presented in Chapter 3 sub-sections 3.4 to 3.53.  
c. The fuel flow rate to the experimental engine was measured using a digital fuel gauge with 
a stop watch mounted to measure time taken for the fuel to be consumed. The calculations were 
as per the equations in Chapter 3 sub-section 3.3.5 and 3.3.7. 
d. The EGR % flow rate was calculated according to the equation on air fraction provided in 
Chapter 3 subsection 3.4.3. 
e. The BTE from the experimental engine was calculated according to the equations in Chapter 
3 subsection 3.3.5. 
f. The BP in this experiment was given as per the equation in Chapter 3 subsection 3.3.6. 
g. The EGTs were measured using mounted temperature thermocouples of k-2 type, including 
the EGR temperature which was measured before it mixed with the intake air fresh charge and 
the constituents of the combustion chamber, through the same k-2 thermocouples. 
154 
 
h. A cylinder pressure transducer was mounted on the engine cylinder head to monitor cylinder 
combustion pressure and collect data values through a system charge amplifier connected to a 
computer data acquisition machine.  
i. The crankshaft position or the crank angle was monitored and measured through a mounted 
encoder near the crankshaft pulley area and data transferred to the data processor. 
j. The emission gases during the experiment were monitored through a five gas exhaust gas 
analyzer, and the smoke intensity was measured by means of an AVL 437C smoke meter. 
k. Since the engine developed maximum power at 1500 rpm and was a variable compression 
engine, all the experiments were conducted based on this nominal engine speed at part load 
and full load, but data could be obtained from different engine loads. Part engine load is 
described as 50 % of engine load and engine full load is described as engine running at 100 % 
load, with a fixed CR of 18.5:1.  
l. The EGR system was modified to suit the experimental engine and to enable the study to be 
conducted, with exhaust gases being tapped from the exhaust pipe and joined to the intake 
manifold air intake system via the air flow meter box through a manually controlled gate valve 
which made it possible for the mixing of EGR gases and the fresh air intake.  
m. The EGR % flow rate was divided into the following modes 0 %, 5 %. 10 %, 15 %, 20 %, 
25 %, and 30 % spaced at intervals of 5 %. 
n. The WPPO fuel blends were prepared in the following percentages order and mixed with 
diesel fuel in 10 %, 20 %, 30 %, 40 % and 100 %, where a 10 % blend was 90 % conventional 
diesel fuel and 10 % WPPO fuel in that order, referred to as WPPOB10 with 10 denoting the 
percentage blend of WPPO by volume supplied.  
o. To avoid contamination of the experimental fuel, each test was conducted after a thorough 
evacuation procedure was conducted on the previous preceding experiment on the fuel lines 
and the fuel injection system mechanism of the test engine. This made it possible to conduct 
an experiment and collect good data and measurements with inputs from the test mode only, 
without fear of contamination and poor results from error. 
8.6 Performance Measurements 
The following are the tables for the various performance parameters and the values obtained 
during the experiment and study of the interactions between test fuels (conventional diesel and 
155 
 
blends of WPPO diesel), EGR application percentages, engine load percentages and EGT 
values in relation to BSFC, BTE and BP. 
Table 8.4: BSFC values in relation to engine load and various test fuel blends of WPPO and conventional 
diesel (CD) 
LOAD % CD WPPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
10 0,612 0,575 0,585 0,595 0,605 0,675 
20 0,560 0,515 0,535 0,545 0,560 0,595 
30 0,525 0,465 0,485 0,540 0,550 0,565 
40 0,465 0,375 0,425 0,535 0,542 0,555 
50 0,375 0,300 0,325 0,475 0,440 0,550 
60 0,355 0,275 0,315 0,450 0,540 0,548 
70 0,350 0,300 0,325 0,400 0,440 0,540 
80 0,356 0,325 0,330 0,385 0,420 0,520 
90 0,360 0,315 0,335 0,380 0,415 0,510 
100 0,380 0,310 0,338 0,375 0,410 0,500 
 
Table 8.5: BSFC kg/kW. hr. values under the effects of EGR % flow rate, full engine load with different 
fuel blends of WPPO and conventional diesel (CD) 
EGR % CD WPPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
0 0,4000 0,3225 0,3615 0,3645 0,3715 0,4751 
5 0,4345 0,4070 0,4915 0,4995 0,5125 0,5522 
10 0,450 0,5350 0,5715 0,6125 0,6483 0,6751 
15 0,464 0,5500 0,5885 0,6225 0,6525 0,6862 
20 0,4755 0,5650 0,5995 0,6452 0,6671 0,6695 
25 0,487 0,5700 0,6100 0,6491 0,6697 0,7116 
30 0,495 0,5780 0,6135 0,6573 0,6682 0,7235 
 
  
156 
 
Table 8.6: BTE % values, at full engine load with blends of WPPO and conventional diesel (CD) 
LOAD % CD WPPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
10 6,5 7,05 6,8 6,75 6,15 5,6 
20 13,25 15,25 15,25 15 13,45 11,95 
30 15,45 20 19 18,15 16,85 13,75 
40 18,15 23 22,25 22,45 21 16,85 
50 20,25 24,7 23,65 23,15 22,1 17,45 
60 20,95 26,15 25 25,25 23,25 18,45 
70 23 26,45 25,35 25,2 24,15 20,25 
80 24,15 26,65 25,75 25,1 24,95 21,45 
90 23 26,85 26,15 25 24,75 20,96 
100 21 27 25,75 24,75 23,25 18,95 
 
Table 8.7: BTE % values, under load with EGR % flow rate application with blends of WPPO and 
conventional diesel (CD) 
EGR % CD WPPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
0 12,15 13,25 13,05 12,75 11,55 10,65 
5 10,25 11,85 11,35 10,6 9,65 9,15 
10 8,15 10,15 9,65 8,85 7,5 7,05 
15 6,35 8,35 8,15 6,95 5,85 5,25 
20 5,15 6,85 6,45 5,65 4,85 4,15 
25 4,35 5,4 4,85 4,2 3,85 3,15 
30 3,95 4,25 4,15 3,85 3,15 2,35 
 
Table 8.8: BP values under engine load with blends of WPPO and conventional diesel (CD) 
LOAD % CD WPPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
10 0,875 0,87 0,868 0,865 0,862 0,855 
20 1,45 1,35 1,255 1,215 1,1 1 
30 2,125 2,15 2,05 1,98 1,86 1,75 
40 2,65 2,55 2,45 2,36 2,33 2,25 
50 3 2,98 2,9 2,86 2,75 2,65 
60 3,25 3,2 3,1 3,05 3 2,9 
70 3,5 3,45 3,35 3,22 3,18 3,05 
80 3,75 3,68 3,6 3,48 3,4 3,25 
90 3,95 3,85 3,76 3,65 3,58 3,45 
100 4 3,95 3,85 3,74 3,7 3,55 
 
157 
 
Table 8.9: Temperature values for exhaust gas under different WPPO blends and conventional diesel (CD) 
with the application of EGR % flow rate at full engine load  
EGR % CD WWPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
0 456 478 478 476 468 490 
5 440 464 395 456 460 478 
10 392 424 378 416 422 415 
15 369 394 370 406 410 386 
20 365 390 360 386 382 375 
25 363 383 358 353 368 364 
30 340 362 345 338 331 320 
 
8.7 Emission Measurements 
This section deals with emission measurements and the data obtained from emissions of NOX, 
CO2, CO, O2 and UHC and smoke opacity. The emissions for these gases were measured by 
the five gas analyzer while the smoke emissions or smoke opacity were measured with a smoke 
meter as mentioned in section 8.5 on experimental procedure. The data obtained is tabulated in 
Tables 8.10 to 8.17. 
Table 8.10: NOX emissions (ppm) values versus engine load with different WPPO blended fuels and 
conventional diesel (CD) under the influence of EGR % flow rate application 
EGR % CD WPPO10 WPPO20 WPPO30 WPPO40 WPPO100 
0 920 1000 1060 1120 1240 1270 
5 812 890 1016 1076 1196 1226 
10 760 830 971 1031 1151 1116 
15 700 696 784 887 947 972 
20 623 591 645 750 778 851 
25 533 501 555 640 633 706 
30 432 401 455 483 475 536 
 
 
 
158 
 
Table 8.11: NOX emissions (ppm) values versus full engine load with different blends of WPPO and 
conventional diesel (CD) 
LOAD % CD WPPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
10 250 280 315 325 345 420 
20 355 375 385 395 405 500 
30 432 415 445 465 485 555 
40 515 535 565 575 600 685 
50 635 655 686 698 745 795 
60 700 685 700 745 792 865 
70 745 705 735 798 865 945 
80 860 775 815 905 975 1095 
90 1000 895 955 1045 1115 1215 
100 1100 995 1065 1145 1225 1325 
 
Table 8.12: UHC emissions (ppm) values under engine load with different fuel blends of WPPO and 
conventional diesel (CD) under the effects of EGR % flow rate application 
EGR % CD WPPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
0 22 23 21 20 19 17 
5 43 35 32 27 23 24 
10 57 46 38 34 29 26 
15 70 58 45 40 37 35 
20 77 68 52 46 44 40 
25 82 72 55 48 43 41 
30 85 75 60 49 46 43 
 
Table 8.13: CO emissions % values versus engine load with the different fuel blends of WPPO and 
conventional diesel CD) 
LOAD % CD WPPO10 WPPO20 WPPO30 WPPO40 WPPO100 
10 4,75 3,95 3,76 3,45 3,35 2,65 
20 4,35 3,55 3,36 3,1 2,97 2,35 
30 4,0 3, 2,85 2,72 2,65 2, 
40 3,5 2,5 2,45 2,35 2,25 1,85 
50 3, 2,25 2,15 2,15 2,15 1,65 
60 3,27 2,55 2,55 2,45 2,25 1,68 
70 35 3, 3, 2,65 2,5 1,7 
80 4,1 3,35 3,36 2,95 2,88 2, 
90 4,65 4,25 4,15 3,75 3,55 2,65 
100 5,1 4,99 4,9 4,6 4,45 3,1 
 
159 
 
Table 8.14: CO emissions % values versus EGR % flow rate with different fuel blends of WPPO and 
conventional diesel (CD) 
EGR % CD WPPO10 WPPO20 WPPO30 WPPO40 WPPO100 
0 4,2 4,35 4,65 4,85 5 5,25 
5 5,25 6,52 6,98 7,28 7,5 8,75 
10 7,65 9,79 10,46 10,91 11,25 12,75 
15 11,68 13,55 15,7 16,37 16,88 18,95 
20 18,25 21,35 22,65 24,55 26,95 28,85 
25 26,56 30,56 33,25 35,75 38,94 42,15 
30 36,25 43,65 49,58 52,69 56,95 59,79 
 
Table 8.15: CO2 emissions % values for various blends of WPPO and conventional diesel (CD) under engine 
load 
LOAD % CD WPPO10 WPPO20 WPPO30 WPPO40  WPPO100 
10 2 1,95 1,25 1,15 1,2  3,35 
20 3 2,5 2,25 1,5 1,85  4,65 
30 3,5 3,25 3 2 2,25  5,45 
40 4,45 4 3,65 3,25 2,75  6,15 
50 5,5 4,98 4,5 3,85 3  6,95 
60 6,35 5,65 5 4,45 3,5  8,05 
70 6,75 6,25 5,85 5,25 4,5  9,3 
80 8,5 7,65 7 6,65 5,98  10,45 
90 9,85 8,55 8 7,75 7,15  11,65 
100 10,85 9,65 8,75 8,35 8  12,75 
 
Table 8.16: CO2 % values for various blends of WPPO and conventional diesel fuel under engine load with 
the application of EGR % flow rate 
EGR % CD WPPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
0 3,85 3,65 3,25 2,75 2,65 6,25 
5 4,45 4,25 3,65 3,25 3 6,55 
10 5,35 4,75 4,25 3,95 3,65 7,25 
15 6,35 5,75 5,15 4,5 4,15 8,75 
20 7,35 6,65 5,95 5,65 5,15 9,95 
25 8,5 7,95 7,45 7,05 6,52 11,95 
30 10,95 9,95 9,65 8,85 8,55 14,35 
160 
 
Table 8.17: Smoke emissions or opacity % values at full engine load, with different blends of WPPO and 
conventional diesel (CD) under the effects of EGR % flow rate application 
EGR % CD WPPOB10 WPPOB20 WPPOB30 WPPOB40 WPPOB100 
0 5,5 5,1 5 5,15 5,05 4,8 
5 6,65 6,15 5,8 5,45 5,15 4,65 
10 6,95 6,6 6,25 5,85 5,55 5 
15 7,53 7,1 6,72 6,25 6 5,4 
20 8 7,65 7,15 6,95 6,55 5,65 
25 9,1 8,45 7,75 7,33 6,9 5,95 
30 9,85 9,25 8,65 8,45 7,85 7,15 
 
8.8 Analysis of Error and Percentage Uncertainties 
This process is carried out for the purpose of performing and identifying the accuracy and 
precision of the measuring tools and instruments used in this experiment work, as errors can 
occur due to other conditions outside of the experiment itself [1], such as calibration of the 
instruments, observational errors [2], manufacturing errors, errors associated with experimental 
set-up and planning, as well as environmental conditions existing during the experiment [3]. 
The list of instruments used and their percentage error of analysis are provided in Table 8.18 
together with the uncertainties of CO, CO2, UHC, NOX, EGT and smoke opacity. 
This percentages of error analysis is derived from the following formula, the root sum square 
method and expressed in equation form [4] as: 
𝑅𝑅 =  �∑𝑛𝑛 2𝑖𝑖=1𝑋𝑋𝑖𝑖          Equation 1 
Where R is the total uncertainty percentage, 𝑋𝑋𝑖𝑖 is the individual uncertainty of all the calculated 
operating parameters, 𝑛𝑛 is the total number of the parameters in the experiment and 𝑖𝑖 is the ith 
term of the computed parameters. The total percentage of the uncertainty is thus calculated 
based on Equation 2 as follows [5]. 
𝑅𝑅 = �𝑋𝑋21 + 𝑋𝑋22 + 𝑋𝑋2 23 + ⋯𝑋𝑋𝑖𝑖         Equation 2 
 
 
161 
 
Table 8.18: Instruments used for measurements and data collection, their measuring range, accuracies 
and percentages of inaccuracies, as calculated from Equation 2  
Instrument Accuracy Measuring Percentage 
Range inaccuracies 
AVL 437C (smoke ±1% 0-100% ±1 
meter) Smoke intensity 
AVL pressure  ±0.01 bar 0-250 bar ±0.01 
transducer GH14D 
VL 365C Angle encoder ±1⁰  ±0.2 
AVL digas 444 (five gas    
analyzer) 
CO ±0.03% to ±5% 0-10% by vol ±0.3 
CO2 ±0.5% to ±5% by vol 0-20% vol ±0.2 
O2 ±5% by vol 0-22% by vol ±0.3 
HC ±0.1% to ±5% 0-20000ppm ±0.2 
by vol 
NOx  ±10% 0-5000 ppm by ±0.2 
vol 
K-2 thermocouple ±1⁰C 0-1250⁰C ±0.2 
Digital stop watch ±0.2s  ±0.2 
U-tube Manometer ±1mm  ±2 
Burette ±0.2cc 1-30cc ±1.5 
 
 
 
 
  
162 
 
References 
 
[1] M. Kanoğlu, "Cryogenic turbine efficiencies," Exergy, An International Journal, vol. 
1, pp. 202-208, 2001. 
[2] K. Krisnangkura, "A simple method for estimation of cetane index of vegetable oil 
methyl esters," Journal of the American Oil Chemists Society, vol. 63, pp. 552-553, 
1986. 
[3] P. Senthilkumar and G. Sankaranarayanan, "Production of waste polyethylene bags in 
to oil and studies performance, emission and combustion characteristics in di diesel 
engine," International journal of humanities, arts, medicine and science, vol. 3, pp. 
149-158, 2015. 
[4] V. K. Kaimal and P. Vijayabalan, "A detailed study of combustion characteristics of a 
DI diesel engine using waste plastic oil and its blends," Energy conversion and 
Management, vol. 105, pp. 951-956, 2015. 
[5] D. Damodharan, A. P. Sathiyagnanam, B. R. Kumar, and K. C. Ganesh, "Cleaner 
emissions from a DI diesel engine fueled with waste plastic oil derived from municipal 
solid waste under the influence of n-pentanol addition, cold EGR, and injection timing," 
Environmental Science and Pollution Research, pp. 1-15, 2018. 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
  
163 
 
CHAPTER NINE 
RESULTS AND DISCUSSION 
9.1 Brake Specific Fuel Consumption (BSFC)  
Figure 9.1 shows the variation of the BSFC under the effects of EGR % flow rate with different 
fuel blends of WPPO and conventional diesel operating at full engine load conditions. 
CD WPPOB10 WPPOB20
WPPOB30 WPPOB40 WPPOB100
0,8
0,6
0,4
0,2
0
0 5 1 0 1 5 2 0 2 5 3 0
EGR %  
Figure 9.1: BSFC versus EGR % rate flow under full engine load conditions 
 
As can be seen from Figure 9.1, The lower ratio blends of WPPOB10 and WPPOB20 show 
low and reduced values of BSFC at 0 % EGR flow rate as in Figure 9.1 and Table 8.5 in Chapter 
8 compared to the values of conventional diesel. However, WPPOB100 shows a significant 
high value of 0.4751 g/kW.hr for BSFC at 0 % EGR flow rate compared to all other test fuels 
even with application of any percentage rate of EGR flow. This is an indication that 
WPPOB100 has a higher BSFC than diesel and the other blends of WPPO which have values 
of 0.3225 g/kW.hr, 0.3615 g/kW.hr, 0.3645 g/kW.hr and 0.3715 g/kW.hr, for WPPOB10, 
WPPOB20, WPPOB30, WPPOB40 and conventional diesel with 0.4 g/kW.hr respectively. 
From the graph in Figure 8.1, it can be seen that as the EGR % flow rate increases from 0 % to 
10 % all the test fuels report an increased BSFC except for the conventional diesel test fuel, 
which continues to have lower values of BSFC irrespective of the EGR flow rate applied. For 
example, at 10 % its value is 0.45 g/kw.hr compared to 0.5715 g/kW.hr, 0.5885 g/kW.hr, 
0.6483 g/kW.hr, 0.6483 g/kW.hr and 0.6751 g/kW.hr respectively for WPPO10, WPPO20, 
WPPO30, WPPO40 and WPPO100. 
At 10 % EGR flow rate the values for the BSFC across all the test fuel except conventional 
diesel seem to show and pick up a sharp increasing curve below the 10 % EGR flow rate, as in 
164 
 
BSFC ( g/kW.hr)
Figure 8.1 but as the EGR flow rate increases from 10 % there is a flattening of the graph 
curves and close value trends observed. 
The WPPO biodiesel blends with EGR % flow rate application showed better fuel economy 
especially the lower blend ratios of WPPOB10 and WPPOB20 as compared to high ratio 
blended test fuels. However, as the EGR % flow rate is increased there is a noticeable increase 
in the BSFC across all the test fuels used. At 0 % EGR conventional diesel is 0.4 g/kW.hr 
compared to 30 % EGR flow rate at 0.495 g/kW.hr, while for the WPPO biodiesel blends 
WPPOB10 is 0.3225 g/kW. as compared to 0.5780 g/kW.hr at 30 % EGR. From Figure 9.1 it 
is evident that the test fuel that showed the highest BSFC value among the blends of diesel and 
conventional diesel test fuel is WPPOB100 which at 0 % EGR % flow rate had a value of 
0.4751 g/kW.hr compared to 0.7235 g/kW.hr at 30 % EGR % flow rate. 
A similar trend is noted regarding the influence of EGR % flow rate where, for example at 
between 20 % to 25 % EGR flow rate, the BSFC continued to show increasing tendencies. This 
finding concurs with the findings of [1]. This phenomenon can be explained due to the effects 
of dilution of the fresh air intake as it mixes with exhaust gases that are being recirculated 
through the EGR system which leads to incomplete combustion of the inducted mixture, thus 
leading to a drop in power and engine torque as reported by the findings of [2]. This scenario 
forces the engine to increase its fuel consumption in order to maintain momentum and constant 
speed, these changes lead to increased load hence the increase in BSFC. 
9.2 Brake Thermal Efficiency (BTE) 
BTE studies specifically help us to know the ability of the combustion system to accept the 
fuel provided as a comparable means of assessing how efficiently fuel conversion was carried 
in turning that fuel into a mechanical output able to do work, as observed by [3, 4]. 
Figure 9.2 shows the BTE % variations under engine load with different blends of WPPO and 
conventional diesel test fuels. 
165 
 
CD WPPOB10 WPPOB20
WPPOB30 WPPO40 WPPOB100
30
25
20
15
10
5
0
1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 0 1 0 0
LOAD %  
Figure 9.2: Variation of BTE versus engine load % 
 
Figure 9.3 shows the BTE % variations under engine loads with different blends of WPPO and 
conventional diesel fuel, with application of EGR % flow rate. 
CD WPPOB10 WPPOB20
WPPOB30 WPPOB40 WPPOB100
15
10
5
0
0 5 1 0 1 5 2 0 2 5 3 0
EGR (%)  
Figure 8.3: BTE % versus EGR % flow rate 
 
The BTE is as provided in Tables 8.6 and 8.7 in Chapter 8 and as provided for in Chapter 3 
sub-section 3.3.5. Figure 9.2 is the variation of the brake thermal efficiency under load with 
different blends of WPPO. There is a noticeable increase in the BTE especially with all low 
ratio fuel blends of WPPOB10 and WPPOB20 as compared to conventional diesel fuel. 
However, in comparison to other WPPO blends WPPOB100 obtained the lowest increment of 
BTE with a value of 5.6 % at 10 % engine load and a maximum increment of 18.95 % at 100 
% engine load. The other blends of WPPO as aforementioned exhibited higher BTE values, 
especially for the low ratio blends compared to the high ratio blends. For example, at 10 % 
engine load the value of the BTE for WPPOB10 is 7.05 % while WPPOB30 and WPPO40 
have values of 6.75 % and 6.15 % respectively. The other blends of WPPO continued to show 
166 
 
BTE (%) BTE % 
this same trend because at 100 % engine load the value for WPPOB10 and WPPO20 is 
observed to be 27.0 % and 25.75 % as compared to WPPOB30 and WPPOB40 which have 
values of 24.75 % and 23.35 % respectively.  
The high BTE values observed for WPPO blends in Figure 8.2 may be due to the high viscosity 
of the WPPO blends as compared to conventional diesel during the combustion pre-mixing 
phase. It is observed that at high engine load the brake thermal efficiency for conventional 
diesel is 21.0 % but to the values for the WPPO blends are 27 %. 25.75 %, 24.75 %, 23. 25 and 
18.95 for WPPOB10, WPPOB20, WPPOB30, WPPOB40 and WPPOB100 respectively.  
Although there is a reduction in BTE due to the application of EGR % flow rate as in Figure 
8.3 here and in Table 8.7 in Chapter 8, the trends of decreased BTE continue to be observed, 
despite showing higher values than conventional diesel except for WPPO100. For example, at 
5 % EGR flow rate, the value of BTE for conventional diesel is 10.25 % compared to 
WPPOB10 and WPPOB20 with values of 11.85 % and 11.35 %. The WPPOB100 blend has 
the lowest value for BTE of all EGR flow rate modes out of all test fuels. At an EGR flow rate 
of 0 % WPPOB100 has a maximum value of 10.65 % compared to 2.35 % at 30 % EGR flow 
rate which was the minimum value obtained. 
There is a correlation between the blend ratio and the rate of EGR flow that is observed. As the 
blend ratio increases there is a general decrease in the BTE across and within all the blended 
test fuels, with an increase in EGR flow rate. For example, all the blends of WPPO at 10 % 
EGR flow rate report values of 10.15 kW.hr, 9.65 kW.hr 8.85 kW.hr, 7.5 kW.hr and 7.05 
kW.hr. However, WPPO20 reports decreasing values of 13.05 kW.hr, 11.35 kW.hr, 9.65 
kW.hr, 8.15 kW.hr, 6.45 kW.hr, 4.85 kW.hr and 4.15 kW.hr as the rate of EGR % flow is 
increasing from 0 % to 30 % respectively. 
9.3 Brake Power (BP) 
Figure 9.4 shows engine BP variations with exhaust gas recirculation (EGR) % flow rate 
application with conventional diesel fuel and different blends of WPPO at full engine load 
conditions. 
167 
 
CD WPPOB10
WPPOB20 WPPOB30
WPPOB40 WPPOB100
5
4
3
2
1
0
1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 0 1 0 0
LOAD (%)
 
Figure 9.4: Variation of engine brake power versus varying engine load % 
 
As can be seen from Figure 9.4 there is a lineal increase in the BP for all the test fuels applied 
with an increase in the engine load. Conventional diesel fuel has the highest increase in BP 
values compared to the blended fuels of WPPO. At 20 % engine load WPPOB10 has a value 
of 1.350 kW compared to conventional diesel at 1.45 kW, representing a difference of 6.8 % 
when the two fuels are compared. 
The blended fuels in Figure 9.4 also show very close increments with increase in engine load 
conditions and a decrease in BP with an increase in the blend ratio for all the blended fuels 
tested when compared to conventional diesel fuel. The increase in the blend ratio showed a 
direct decrease in BP in a linear incremental relationship across the blends. For example, at 60 
% engine load WPPOB10, WPPOB20, WPPOB30 WPPOB40 and WPPO100 have values of 
3.2 kW, 3.1kW, 3.05 kW, 3 kW and 2.9 kW respectively showing a decrease in the value of 
the engine BP as the load increased across the blends throughout the experimentation period. 
These findings concur with the findings by another researcher in relation to WPPO blends [5]. 
WPPOB100 blend showed the lowest values for the engine BP compared to the blends of 
WPPOB10, WPPOB20, WPPOB30 and WPPOB40. At 30 % engine load WPPOB100 has a 
value of 1.75 kW compared to WPPOB10 at 2.15 kW, WPPOB20 at 2.05 kW, WPPOB30 at 
1.98 kW and WPPOB40 at 1.86 kW respectively. From this result, it can be established that all 
the blends of WPPO produce less engine power compared to the power produced by the 
conventional diesel test fuel. This result is attributed to their differences in GCV as outlined in 
Table 8.2 in Chapter 8. 
168 
 
BP (KW)
The application of EGR % flow rate is observed to cause no significant change in BP. However, 
there was a negligible drop in the engine BP with the influence of EGR % flow rate except for 
the blend WPPOB10 which had almost identical values to conventional diesel as the curve of 
the two fuels indicate in Figure 9.4 and Table 8.8 in Chapter 8.  
9.4 Exhaust Gas Temperature (EGT) 
Temperature is one of the key factors in determining the formation of engine exhaust emissions, 
besides providing or helping in the analysis and study of combustion processes in relation to 
fuel as observed by [6]. 
Figure 9. 5 shows the variation of EGT under different types of WPPO blends and conventional 
diesel test fuel with application of EGR % flow rate. 
CD WWPOB10 WPPOB20
WPPOB30 WPPOB40 WPPOB100
600
400
200
0
0 5 1 0 1 5 2 0 2 5 3 0
EGR (%)  
Figure 9.5: Variation of EGT ⁰C versus EGR % flow rate 
 
The graph in Figure 9.5 and Table 8.9 in Chapter 8 show a variation in EGT with different fuel 
blends and conventional diesel with the application of EGR % flow rate. The result indicates 
that EGT decreases with different blends of WPPO compared to conventional diesel, although 
the temperature difference between them is that WPPO blends have higher temperature 
increases in all the test conditions compared to conventional diesel fuel. However, it should be 
mentioned that as the blend ratio increases with increase in application of the EGR % flow rate, 
the EGT reduced significantly and marginally especially for WPPOB30 and WPPOB40. For 
example, at 10 % EGR flow rate the values are 416 ⁰C and 422 ⁰C. However, when the rate of 
EGR flow is increased to 20 % the values drop to 386 ⁰C and 382 ⁰C respectively. The highest 
temperature obtained for conventional diesel is 456 ⁰C at 0 % EGR flow rate, whereas the 
highest temperature value for a WPPO blend is 490 ⁰C obtained from WPPOB100 at 0 % EGR 
flow rate, although at 30 % EGR flow rate this blend was most reduced in temperature 
169 
 
EGT (⁰C)
compared to the other WPPO blends with a temperature value of 320 ⁰C, while WWPO10, 
WPPO20, WPPO30 and WPPO40 had values of 362 ⁰C, 345 ⁰C, 338 ⁰C and 331 ⁰C 
respectively. 
The application of EGR % flow rate in increasing modes brings further reduction in EGT with 
the highest value for conventional diesel test fuel obtained being 440 ⁰C when EGR % flow 
rate was at 5 %, while the lowest value of 340 ⁰C was obtained at 30 % EGR flow rate. The 
WPPO blends show a similar trend with decreasing temperatures with the increasing 
application of EGR % flow rate. WPPOB10 reports the highest values to be 467 ⁰C and the 
lowest to be 362 ⁰C at 5 % and 30 % EGR flow rates respectively, while WPPOB40 shows its 
highest value to be 472 ⁰C and the lowest to be 330 ⁰C at 5 % and 30 % EGR flow rates 
respectively. This concludes that all the WPPO blends report low EGT values compared to the 
conventional diesel test fuel. 
It is imperative to note here that a decrease in EGT is noticed irrespective of the continuous 
increase in EGR % flow rate even after the 15 % point EGR flow rate is reached. The main 
cause for reduction in the EGT can be attributed to several factors: The reduction in EGT 
among the different blends of WPPO could result from low CV of the blends and low exhaust 
loss which concurs with the findings of  [7, 8].The WPPO has a CV of 40.15 kJ/kg compared 
to the CV of conventional diesel at 44.84 kJ/kg as seen in Table 8.2.The third cause is directly 
linked to the effects of exhaust gas recirculation percentage flow rate, the dilution effect, 
chemical effects and thermal effects [9, 10] which are extensively discussed in Chapter 2. 
9.5 NOX Emissions 
Figure 8.6 shows the variations of NOX emissions (ppm) and engine load with different blends 
of WPPO and conventional diesel fuel with EGR % flow rate application. 
CD WPPO10 WPPO20
WPPO30 WPPO40 WPPO100
1500
1000
500
0
0 5 1 0 1 5 2 0 2 5 3 0
EGR (%)
 
Figure 9.6: Variations of NOX emissions (ppm) versus EGR % flow rate 
170 
 
NOX (PPM)
 
NOX emissions formation is highly dependent on in-cylinder temperature, the concentration of 
oxygen and the residence time the fuel-air mixture spends in the combustion chamber during 
the pre-mixing phase [11]. All the fuels tested indicated a drop in NOX emissions with an 
increase in the application of EGR % flow rate, at all engine load conditions. Table 8.10 in 
Chapter 8 and Figure 9.6 shows that this could be due to the rise in the total heat capacity of 
the working gases that are increasing with increasing EGR % flow rate thus concurring with 
the studies and findings of [12, 13]. 
The NOX emissions values for conventional diesel in Figure 9.6 and Table 8.10 is 920 ppm for 
conventional diesel test fuel, while for WPPOB100 it is 1325 ppm at full engine load of 100 
%, compared to the reduced values with application of EGR flow rate of 30 % as in Figure 9.7 
and Table 8.11, where conventional diesel decreases its NOX emissions to 432 ppm compared 
to 1100 ppm at full load of 100 % before application of EGR % flow rate. 
As the blend ratio increases with increased EGR % flow rate there is an increase in NOX 
emissions across CD and the WPPO blends but with reduced NOX emissions within the blends 
and within the CD test fuel as in Figure 9.6 and Table 8.10. For example, at 20 % EGR flow 
rate, blends WPPO10, WPPO2O, WPPO30, WPPO40 and WPPO100 had 591 ppm, 645 ppm, 
750 ppm, 778 ppm and 851 ppm respectively. However, WPPO20 has reduced values of 1060 
ppm, 1016 ppm, 971 ppm, 784 ppm, 645 ppm, 555 ppm and 455 ppm as the rate of EGR flow 
increases from 0 % to 30 %. 
Figure 9.7 shows variations of NOX emissions (ppm) under engine load with different blends 
of WPPO and conventional diesel test fuel. 
CD WPPOB10 WPPOB20
WPPOB30 WPPOB40 WPPOB100
1500
1000
500
0
0 5 0 1 0 0 1 5 0
LOAD (%)   
Figure 9.7: Variations of NOX emissions (ppm) versus varying engine load % 
171 
 
NOX (PPM)
 
During experimentation it was observed that at part engine load as in Table 8.11 in Chapter 8 
and Figure 9.7 the values for NOX emissions for the same fuels were lower compared to the 
full load engine conditions. The NOX emission for conventional diesel at engine part load (50 
%) is 635 ppm compared to full load at 1100 ppm, whereas the value for WPPOB100 at engine 
part load (50 %) is 795 ppm compared to 1325 ppm at full engine load. This seems to indicate 
a concurrence that at part engine load (50 %) the values of NOX emissions emitted by all the 
blends of WPPO except WPPO100 are lower as compared to the values reported at full engine 
load conditions.  
9.6 Hydrocarbon (HC) Emissions 
Figure 9.8 shows the variation of HC emissions in parts per million under full engine load with 
the application of various EGR % flow rates, different blends of WPPO and conventional 
diesel. The HC emissions were significantly higher across all the blended test fuels of WPPO, 
especially with higher engine load conditions as indicated by the values shown in Figure 9.8 
and Table 8.12 in Chapter 8. However conventional diesel still produced higher values of HC 
emissions compared to all blends of WPPO across all the engine loading conditions and 
operating modes. 
Figure 9.8 shows UHC emission (ppm) values under engine loads with different fuel blends of 
WPPO and conventional diesel under the effects of EGR % flow rate application. 
CD WPPO10 WPPO20
WPPO30 WPPO40 WPPO100
100
50
0
0 5 1 0 1 5 2 0 2 5 3 0
EGR (%)  
Figure 9.8: Unburnt hydrocarbons emissions versus EGR % flow rate 
 
When the EGR % flow rate application is 0 % there is no application effect. Figure 9.8 shows 
that there are less HC emissions for all the test fuels applied in this experiment, compared to 
172 
 
UHC (PPM) 
20 % or 30 % EGR percentage flow rates. The value for the conventional diesel test fuel was 
22 ppm at 0 % EGR flow rate and for the blends of WPPO the values were 23 ppm, 21 ppm, 
20 ppm, and 17 ppm for WPPOB10, WPPOB20, WPPOB30, WPPOB40 and WPPOB100 
respectively. However, with EGR flow rate of 10 % was applied the values are 57 ppm for 
conventional diesel and the blends of WPPO 46 ppm, 38 ppm, 34 ppm, 29 ppm and 26 ppm 
respectively. 
Although the application of EGR % flow rate reduces the amount of HC emissions from the 
applied test fuels across the board, conventional diesel fuel still produced more HC emissions 
from the test engine compared to all the WPPO blends tested. Figure 8.8 shows that at EGR 
flows rate of 5 %, 10 %, 15 %, 20 %, 25 % and 30 %, conventional diesel had 43 ppm, 57 ppm, 
70 ppm, 82 ppm and 85 ppm respectively, as compared to WPPOB10 whose values are 23 
ppm, 35 ppm, 40 ppm, 48 ppm, 50 ppm, and 52 ppm respectively.  
Therefore, the application of EGR % flow rate increased HC emissions as observed and 
indicated from the results and values presented in Figure 9.8 and Table 8.12 in chapter 8, with 
conventional diesel fuel producing the highest HC emission values as compared to all the test 
fuel blends of WPPOB10, WPPOB20, WPPOB30, WPPOB40 and WPPOB100. 
There was a correlation established between blend ratio, EGR flow rate and the amount of UHC 
emissions. As the EGR flow rate increases there is an increase in total UHC emissions produced 
within a blend and vice versa for all test fuel. However, there is a decrease in UHC as the blend 
ratio increased with different blends of WPPO as the EGR flow rate increased and vice versa. 
For example, the values of WPPO10 are 23 ppm, 35 ppm, 46 ppm, 58 ppm, 68 ppm, 72 ppm 
and 75 ppm for EGR flow rates of 0 %, 5 %, 10 %, 15 %, 20 %, 25 % and 30 % respectively, 
while the values for conventional diesel and the other blends of WPPO at 10 % EGR flow rate 
are 57 ppm, 46 ppm, 38 ppm, 34 ppm, 29 ppm and 26 ppm. 
9.7 Carbon Monoxide (CO) Emission 
Figure 9.9 shows the variations of CO emission % with load under the effects of EGR % rate 
flow application, with different fuel blends of WPPO and conventional diesel fuel. CO is a 
toxic gas that requires substantial control to maintain acceptable levels. It is caused by poor 
combustion of hydrocarbon fuels as a result of dependency on the AFR relative to the 
stoichiometric proportions [12, 14].  
173 
 
Figure 9.9 shows CO emission % variations versus varying engine load with different fuel 
blends of WPPO and conventional diesel test fuel.  
CD WPPO10 WPPO20
WPPO30 WPPO40 WPPO100
6
4
2
0
0 2 0 4 0 6 0 8 0 1 0 0 1 2 0
LOAD (%)
  
Figure 9.9: CO emission % versus varying engine load 
 
For all the test fuels the amount of CO emissions seemed to decrease at lower engine loads up 
to part load percentages or intermediate loads of (50 %), thereafter the CO emissions continued 
to increase significantly and marginally as in Table 8.13 in Chapter 8and Figure 9.9. At 10 % 
engine load the value of conventional diesel is 4.75 % compared to 50 % engine load when the 
value is reduced to 3.0 % by volume. However, as the engine load is increased from 50 % 
upwards there is a significant continuous and marginal increase in the percentage of CO 
emissions by volume as the load is increased across all the test fuels. For example, at 80 % 
engine load the value for WPPOB100 is 2.0 % up from 1.65 % by volume, at 50 % part engine 
load the value of conventional diesel is 4.1 % compared to 2.95 % for WPPOB100 by volume 
at 50 % part engine load. The other WPPO biodiesel blends also show a similar trend and 
concurrency. WPPOB20 test fuel at 50 % part engine load condition has a value of 2.15 % by 
volume as compared to 3.36 % by volume at 80 % engine load. 
Figure 9.10 shows the variation of CO emissions with EGR % flow rate application under 
different blends of WPPO and conventional diesel.  
174 
 
CO (%)
CD WPPO10 WPPO20
WPPO30 WPPO40 WPPO100
100
50
0
0 5 1 0 1 5 2 0 2 5 3 0
EGR (%)
 
Figure 9.10: Variation of CO VS EGR % flow rate application 
As can be seen from Figure 9.10, the WPPO blends are observed to produce a continuous 
increase in the smoke emissions almost to double the values with the application of EGR % 
flow rate as in Figure 9.10 and Table 8.14 in Chapter 8. For example, at 10 % EGR flow rate, 
CO emission values are 9.79 %, 10.46 %, 10.91 %, 11.25 % and 12.75 % for WPPO10, 
WPPO20, WPPO30, WPPO40 and WPPO100 respectively. Conventional diesel test fuel 
reported the lowest CO emissions with the application of EGR % flow rate with a value of 7.65 
%.  
There is also a correlation between the blend ratio and EGR % flow rate and the amount of CO 
emissions produced. During experimentation it was observed that as the blend ratio increases 
the CO emissions increase within the blend as the EGR % flow rate increases as in Figure 9.10 
and Table 8.14 in chapter 8. At 20 % EGR flow rate CO emission values are 18.25 %, 21.35 
%, 22.65 %, 24.55 %, 26.95 % and 28.85 % respectively for conventional diesel, WPPO10, 
WPPO20, WPPO30, WPPO40 and WPPO100. However, WPPO30 reports values of 4.85 %, 
7.28 %, 10.91 %, 16.37 %, 24.55 %, 35.75 % and 52.69 % as the EGR flow rate increases from 
0 % to 30 %. This may be due to the effects of EGR % flow rate in terms of dilution, and 
thermal and chemical effects as a result of some of the oxygen in the inlet charge being replaced 
with recirculated exhaust gas causing incomplete combustion.  
9.8 Carbon Dioxide (CO2) Emissions  
CO2 is the principal gas composition of the recirculating exhaust gas; however it is a core 
indicator of the quality of combustion and the existing in-cylinder temperatures [15] of 
combustion within the combustion chamber. CO2 gas with a higher heat capacity makes it act 
like a thermal heat sink especially during the combustion process, leading to a possible 
175 
 
CO (%)
reduction of peak cylinder temperatures and of particular importance is the reduction in the 
NOX emissions. 
Without EGR % flow rate and at lower engine loads, the value of CO2 is considerably high for 
all the test fuels used. For example, at 20 % engine load WPPOB100 has the highest CO2 
emissions compared to all the other test fuels used. CO2 emission values for conventional diesel 
fuel is 3.0 % by volume compared to WPPOB100 which is 4.6 % by volume as seen in Figure 
9.11 and Table 8.15 in Chapter 8.  
However, it can also be observed from Figure 9.11 and Table 8.15 in Chapter 8 that the amount 
of CO2 increases with increase in engine load. As the engine load increases to 40 % the value 
of WPPOB40 is 2.75 % compared to 60 % when the value is 3.5 %. The test blend WPPOB30 
at 40 % is 3.25 % by volume compared to 4.5 % and 5.25 % for WPPOB40 and for WPPOB30 
respectively at 70 % engine load. This leads to the observation that as the engine load increases 
and the blend ratios increase, lower blends ratios are observed to emit more CO2 emissions 
compared to those blends with high ratios except for blend WPPOB100 which released more 
carbon emissions than any of the WPPO blended test fuels. At full engine load the value of 
CO2 emissions is at its highest values as seen in Figure 9.11 and Table 8.15 in Chapter 8 across 
all the test fuels. The value for conventional diesel is 10.85 % compared to the blend 
WPPOB100 at 12.75 %, WPPOB10 at 9.65 %, WPPOB20 at 8.75 %, WPPOB30 at 8.35 % and 
WPPOB40 at 8.0 %. 
Figure 9.11 shows CO2 % variation under engine loads with different blends of WPPO and 
conventional diesel test fuel.  
CD WPPO10 WPPO20
WPPO30 WPPO40 WPPO100
15
10
5
0
1 0 2 0 3 0 4 0 5 0 6 0 7 0 8 0 9 0 1 0 0
LOAD (%)  
Figure 9.11: Variation of CO2 % emissions versus engine load %, with different types of fuel blends of 
WPPO and conventional diesel 
 
176 
 
CO2 (%)
Figure 9.12 shows CO2 % variations with different blends of WPPO and conventional diesel 
test fuel under varying engine loads, with application of EGR % flow rate. 
CD WPPOB10 WPPOB20
WPPOB30 WPPOB40 WPPOB100
20
10
0
0 5 1 0 1 5 2 0 2 5 3 0
EGR (%)
 
Figure 9.12: Variations of CO2 % versus EGR % flow rate 
 
The results in Figure 9.12 and Table 8.16 in Chapter 8 reinforce the observation that under 
EGR % flow rate CO2 emissions increase with increase in the blend ratio and EGR % flow 
rate. The higher the blend ratio and the higher the increases in EGR % flow rates the higher the 
CO2 emission values and vice versa. The values for 30 % EGR flow rate for all the test fuels 
contain high emissions CO2, for example conventional diesel is 10.95 %, WPPOB10 9.95 %, 
WPPOB20 9.65 %, WPPOB30 8.85 % and with WPPOB100 showing the highest value of 
14.35 % for carbon dioxide emissions. 
9.9 Smoke Emissions (Opacity) 
This can be defined as the solid hydrocarbon soot particles that are found in the exhaust gas 
system [16] and directly linked to smoke emissions formation. For all the blends of WPPO 
there is a noted increase in the level of smoke emissions, although compared to conventional 
diesel the levels and values are considerably lower.  
The application of EGR % flow rate as seen in Figure 9.13 and Table 8.17 in Chapter 8 shows 
that there is a significant increase in the values of smoke emissions and particulate matter 
emissions across all the test fuels employed. Smoke emissions with WPPOB10 blend are seen 
to be 5.1 % lower compared to conventional diesel at 0 % EGR flow rate with 5.5 %. On the 
other hand, conventional diesel is higher than WPPOB100 blend of fuel at 30 % EGR flow rate 
with a value 9.85 % compared to WPPO100 with a value of 7.3 %. This result seems to concur 
with the findings and study of [17]. 
177 
 
CO2 (%)
The WPPOB10 blend during experimentation is seen to emit the highest levels of smoke 
emissions for the blended fuels compared to the other WPPO blends. However, it should be 
mentioned here that as the blend ratio and the EGR % flow rate increased there is consistent 
increase in the smoke emissions within the test fuels with marginal decreases observed across 
all the blended fuel with values at 15 % EGR flow rate reporting 7.53 %, 7.1 %, 6.72 %, 6.25 
%, 6.0 % and 5.4 % for conventional diesel, WWPO10, WPPO20, WPPO30, WPPO40 and 
WPPO100 respectively in decreasing marginal values, compared to increasing values within 
the blend of WPPO20 with values of 5.0 %, 5.8 %, 6.35 %, 6.72 %, 7.15 %, 7.55 % and 8.65 
%, for increasing EGR flow rates of 0 % to 30 %. 
Figure 9.13 shows smoke emissions or opacity % variations at full engine load, with different 
blends of WPPO and conventional diesel under the effects of EGR % flow rate application. 
CD WPPOB10 WPPOB20
WPPOB30 WPPOB40 WPPOB100
12
10
8
6
4
2
0
0 5 1 0 1 5 2 0 2 5 3 0
EGR (%)  
Figure 9.13: Variation of smoke emissions or opacity % versus EGR % flow rate 
 
The constant increase in smoke emissions can be explained by the fact that WPPO blends of 
fuel have a high kinematic of viscosity compared to conventional diesel test fuel. Besides the 
low volatility values of the WPPO tested blends, this affects their spray characteristics, mixing 
qualities and combustion patterns. Other possible explanations for this phenomenon is the poor 
injection and spray characteristics observed with most WPPO blends of fuels compared to the 
spray and injection characteristics of conventional diesel fuel, which has better spray and 
mixing qualities. Another likely cause associated with this phenomenon is the high aromatic 
compounds found in most WPPO blends of fuel compared to conventional diesel test fuel. 
  
178 
 
SMOKE EMISSIONS (%)
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180 
 
CHAPTER TEN 
CONCLUSION AND RECOMMENDATIONS 
10.1 Conclusion  
• As the percentage of the blends of WPPO increased there was a marked decrease in the 
engine BP from the blended fuels of WPPO compared to conventional diesel. This 
makes sense considering that the energy content (as seen in the table of properties of 
the WPPO test fuel) is lower compared to the energy content of conventional diesel. 
• Peak power produced using WPPO biodiesel blends failed to match the peak power 
produced by conventional diesel test fuel. The difference in peak power ranged between 
5 % to 8 % less compared to the peak power produced from conventional diesel test 
fuel. 
• Data presented in this work provides more reasons to support the widespread use of 
WPPO as an alternative fuel for all types of compression ignition engines with or 
without modifications to the engine, especially when blends of WPPOB10 and 
WPPOB20 are used because their peak BP and BSFC values have identical values to 
conventional diesel fuel. 
• As the percentage of the blend ratio of the biodiesel WPPO increased there was a 
significant increase in the BSFC as in Figure 9.1 for the blends of WPPO compared to 
the brake specific fuel consumption (BSFC) values of conventional diesel test fuel. 
• During experimentation it was noted that after the 10 % EGR flow rate the values for 
the BSFC across all the test fuel except conventional diesel seem to show and pick a 
sharp increasing curve below the 10 % EGR flow rate, as in Figure 9.1 and Table 8.5 
in Chapter 8, but as the EGR flow rate increases from 10 % there is a flattening of the 
graph curves leading to close value trends. 
• After the 10 % EGR flow rate the values for the BSFC across all the test fuel seem to 
show and pick a lineal increment trend as can be shown in Figure 9.1 by flattening of 
the graph and close value trends. 
• Another observation reported during this experiment is that as the blends ratio is 
increased with increasing load there is a reduction in the percentage amount of CO 
emissions released by the test engine as seen in Table 8.13 in Chapter 8 and Figure 9.9 
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in Chapter 8. This can be attributed to a high oxygen content in WPPO blended fuels 
compared to conventional diesel, which aided in complete combustion. 
• The results values of CO emissions by two blends of WPPOB10 and WPPOB20 are 
observed to be close with minimal differences in terms of their volume percentage 
emissions of CO produced by the test engine as seen in Figure 9.9 in Chapter 9. 
• There is a correlation between the blend ratio and EGR % flow rate on the amount of 
CO emissions produced. During experimentation it was observed that as the blend ratio 
increases the CO emissions increase within the blend as the EGR % flow rate increased 
as evident in Figure 9.10 in Chapter 9 and Table 8.14 in Chapter 8. 
• Figure 9.11 in Chapter 9 and Table 8.15 in chapter 8 show that the amount of CO2 
increases with the increase in the engine load; as the engine load increases to 40 % the 
value of WPPOB40 is 2.75 % compared to 60 % where the value is 3.5 %. 
• Under the influence of EGR % flow rate, CO2 emissions increase with increases in the 
blend ratio and EGR % flow rate. The higher the blend ratio and the higher the increases 
in EGR % flow rates the higher the CO2 emission values and vice versa.  
• There was an observed constant increase in smoke emissions during experimentation 
for all the blends of WPPO with or without EGR % flow rate application, with test fuel 
blend WPPOB10 producing the highest values of smoke emissions followed by the 
WPPOB 100 test fuel blend. 
• The EGTs during experimentation were seen to constantly decrease with application of 
EGR % flow rate. It is imperative to note here that decrease in EGT is noticed 
irrespective of the continuous increase in EGR % flow rate even when the EGR % flow 
rate went above the 15 % point. 
• There was a concurrence that at part engine load (50 %) the values of NOX emissions 
emitted by all the blends of WPPO except WPPO100 were lower compared to the 
values at full engine load conditions. 
• During part load and intermediate engine load it was observed that the engine operating 
under different blends of WPPO and conventional diesel fuel emitted less NOX 
emissions compared to the full load mode condition as demonstrated in the results 
shown in Figure 9.7 in Chapter 9 and Table 8.11 in Chapter 8. 
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• The application of EGR % flow rate increased HC emissions as observed from the 
results and values presented and obtained in Figure 9.8 in Chapter 9 and Table 8.12 in 
Chapter 8, with conventional diesel fuel producing the highest HC emission values 
compared to all the test fuel blends of WPPOB10, WPPOB20, WPPOB30, WPPOB40 
and WPPOB100. 
• There was a correlation established between blend ratio, EGR flow rate and the amount 
of UHC emissions. As the blend ratio increased there was a reduction in total UHC 
emissions produced across all blends. For example, at 10 % EGR flow rate the values 
of WPPO10, WPPO20, WWPO30, WPPO40 and WPPO100 were 46 ppm, 38 ppm, 34 
ppm, 29 ppm and 26 ppm respectively. 
10.2 Future Recommendations 
• EGR as a method of controlling emissions, although seeming to contribute to other 
emission problems, thus making the use of compression ignition engines undesirable, 
is an effective technique in the control of NOX emissions in diesel engine emissions. 
This technique in combination with other NOX emission control systems is one of the 
most efficient technique in combustion control strategies in not only controlling 
compression ignition engines harmful emissions but also in reducing emissions to 
acceptable standards.  
• It can be safely concluded that EGR can be used as a technique for controlling emissions 
of NOX, CO, CO2, HC and SOF in conjunction with other techniques. However, more 
needs to be done to bridge the gap between research and industry, for example increase 
the use of EGR cooling particularly in stationery engine systems and in marine 
applications and spread to a wider area in industrial production where carbon emissions 
are more pronounced. In this regard, the use of high boost pressure coupled with EGR 
application should be encouraged.  
• EGR cooling will allow us to retain the benefits of low NOX emissions without 
compromising engine efficiency. EGR cooling is necessary as it reduces soot and SOF 
emissions, for example when used with other techniques like variable gas turbine 
technology, which recirculates all exhaust gases together with the inlet intake mixture, 
thus increasing the mixture heat capacity leading to lowered in-cylinder combustion 
temperatures and consequently a reduction in NOX emissions.  
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• During the review it was observed that the preheating of the inlet air reduced CO 
emissions. However, the introduction of EGR % flow rate into the system is observed 
to increase CO emissions linearly as the EGR % flow rate was being increased, which 
calls for the question of harmonization of the two systems to achieve low carbon 
emissions without compromising engine performance and emission control. 
• From the combustion control strategies that were discussed, what implementation 
strategies can we employ to harness improving the combustion characteristics of the 
alternative fuels available, especially their chemical and physical properties? The 
chemical and physical properties of biodiesels lack proper testing and analysis facilities 
because this is still a grey area in terms of industrial commercial development, as 
industry regulators impose stringent control on the use of alternative fuels.  
• More work needs to be done to develop high quality alternative biodiesel fuels suitable 
for future combustion engines and systems, especially projects involving waste to 
energy alternatives such as using waste plastic to generate propulsion fuel. This will 
address two issues – finding an answer to an alternative fuel and reducing pollution to 
natural environment from municipal solid waste management landfill sites. 
• Although EGR mitigates formation of the NOX gases, it promotes smoke emissions 
especially during higher engine loads, i.e. it requires to be used in combination with 
other emissions control techniques and combustion control strategies like LPL EGR, 
HPL, selective catalytic reduction (SCR), diesel particulate filter (DPF), diesel 
oxidation catalyst (DOC) and retarded timing. Another area that needs new studies and 
research is the development of accurate measurement of different rates of EGR in 
transient engine conditions for implementation and control of the exhaust gases when 
hot EGR or cold EGR is used.  
• The personal modern day transport system is a major contributor to traffic related 
pollution, followed by the mass transport vehicles, buses, the rail and air transport. 
These systems impact negatively on the health of the world population. More work 
needs to be done to improve air pollution control strategies to include control of 
additional key pollutants. There is a need to do improve measuring technology, increase 
observatory sites and increase satellite information recovery and retrieval time beside 
the development of more advanced simulation models.  
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• During the literature review it became evident that more information is needed 
regarding the toxicity of some of the toxic gases emitted by diesel engines. The real 
impact of exhaust gas emissions is still limited according to the studies that have been 
conducted so far. Therefore, more impact assessment studies need to be conducted. 
• Although tremendous progress has been made in reducing air pollution and thus 
improving public health, there is still a need for more isolation of the individual 
pollutants species and source categorization to increase specific understanding on the 
effects produced on human health, making it possible institute control actions on 
different sources and better prioritization of mitigation measures.  
• The differences in the disciplinary backgrounds of most air pollution research groups 
has to be bridged. The current pollution problem requires a multi-disciplinary and inter-
disciplinary approach to increase interaction between the disciplines of science and 
engineering technology going forward, to better incorporate new findings, capabilities, 
limitations, and research priorities as team work instead of individual fields with 
individual discipline solutions. 
• Funding of new and young researchers, better educational opportunities for young 
researchers and encouragement of the inter-disciplinary conferences and meetings to 
break barriers of traditional disciplinary boundaries should be encouraged. This is 
necessary considering the importance and urgency of the air pollution problem globally.  
 
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