examination of egr cooler fouling and engine …

The Pennsylvania State University

The Graduate School

Department of Energy and Mineral Engineering

EXAMINATION OF EGR COOLER FOULING AND ENGINE EFFICIENCY

IMPROVEMENT IN COMPRESSION IGNITION ENGINES

A Dissertation in

Energy and Mineral Engineering

Fuel Science Option

by

Bhaskar Prabhakar

2013 Bhaskar Prabhakar

Submitted in Partial Fulfillment

of the Requirements

for the Degree of

Doctor of Philosophy

May 2013

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The dissertation of Bhaskar Prabhakar was reviewed and approved* by the following:

André L. Boehman

Adjunct Professor of Energy and Mineral Engineering, Adjunct Professor

of Mechanical Engineering

Dissertation Advisor

Co-Chair of Committee

Randy L. Vander Wal

Professor of Energy and Mineral Engineering and Materials Science and

Engineering

Co-Chair of Committee

Jonathan P. Mathews

Assistant Professor of Energy and Mineral Engineering

Daniel C. Haworth

Professor of Mechanical Engineering

Luis F. Ayala H.

Associate Professor of Petroleum and Natural Gas Engineering

Graduate Program Officer of Energy and Mineral Engineering

*Signatures are on file in the Graduate School

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ABSTRACT

The scope of this investigation is to understand the challenges associated with

achieving high engine efficiency and low emissions in ‘Clean Diesel’ technology. The

topics addressed in this study are: 1) understanding the challenges in reducing NOx

emissions due to fouling of EGR coolers, 2) exploring high efficiency dual fuel

combustion with fumigation of liquefied gases into engine air intake, and 3)

understanding the effect of diesel fuel formulation on engine efficiency and emissions.

In the first study, the performance of a model EGR cooler attached to a 6.4L

turbodiesel engine was investigated by analyzing the microstructure and chemical

composition of the deposits on the fouled heat exchanger surfaces at two engine loads:

medium and low, and at two coolant temperatures: 85°C and 40°C. Results indicated that

the medium load condition resulted in greater thermal effectiveness loss and mass gain

inside the EGR cooler, mostly due to increased thermophoresis, producing smaller (grain

size) and coarser deposits. In contrast, the low load condition resulted in lower

effectiveness loss, but produced large-sized deposits mostly due to increased hydrocarbon

(HC) condensation. Regardless of the engine load, effectiveness and deposit mass gain

plateaued in about 9 hours. Coolant temperature played a significant role in altering the

deposit microstructure and in increasing the amount of condensed HCs. Deposit mass

increased for the 40°C coolant condition due to an increase in both HC condensation and

thermophoresis. For most conditions, the deposits were comprised of some aromatics and

mostly heavy aliphatics (C17-C25 paraffins) which arise due to incomplete combustion

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of the heavy-end long chain compounds present in the fuel and lubricating oil. Low

coolant temperatures promoted higher effectiveness recovery during engine start-up

suggesting an influence of condensed water vapor on deposit layer removal. Use of an

oxidation catalyst upstream of the EGR cooler to reduce hydrocarbon condensation was

not effective at the engine conditions tested in the study due to low operating

temperatures across the catalyst.

In the second study, the role of ignition quality of a fumigated fuel on combustion

phasing and brake thermal efficiency (BTE) was investigated on a 2.5L turbocharged

common rail light-duty diesel engine, in a process similar to dual fuel combustion.

Different combinations of DME and propane were fumigated into the intake air via a

specially designed manifold assembly, each combination representing a percentage of the

energy supplied to the engine, with rest of the fuel being ultra-low sulfur diesel (ULSD).

Fumigation of DME and propane significantly increased BTE and reduced brake specific

energy consumption (BSEC) compared to the baseline diesel condition with no

fumigation. A mixture of 20% DME with 30% propane provided the maximum BTE,

with 24% reduction in BSEC, however, at the expense of increasing peak cylinder

pressure by 6 bar, which was even higher at greater DME%. Fumigated DME auto-

ignited early, ahead of top dead center (TDC), showing the typical low temperature and

high temperature heat release events and propane addition suppressed the early low

temperature heat release (LTHR), shifting more of the DME heat release closer to TDC.

Total hydrocarbon emissions decreased with DME substitution, and increased with

propane substitution. NOx emissions reduced with increasing DME and propane

v

substitution, and were 45% lower at the peak BTE in comparison to the baseline diesel

condition. CO emissions increased with increasing propane and DME substitution until

20% DME (455 ppm at baseline to 3700 ppm at the peak BTE), while CO2 emissions

decreased mainly with increasing propane while they remained more or less constant with

increasing DME substitution. It was concluded that DME and propane fumigation into

the air intake offered a pathway to high efficiency combustion.

In the final work, the sensitivity of fuel conversion efficiency and engine

performance to fuel formulation was evaluated at different engine operating conditions

under a range of ultra-low sulfur diesel fuels. The fuels were comprised of a commercial

baseline ULSD, along with six other diesel fuels varying in derived cetane number

(DCN), total aromatic percentage, and distillation temperature (T90). Increasing the DCN

improved BSFC and BTE, and reduced PM, HC and CO emissions, while no conclusive

trend was observed in NOx emissions. Increasing the aromatic content reduced BTE and

increased BSFC and other regulated emissions. Increasing the T90 temperature did not

have much effect on BSFC, BTE, and NOx emissions, however, PM emissions increased

while HC and CO emissions marginally decreased. Overall, a fuel with a high DCN, a

low aromatic content, and a high T90 resulted in high engine efficiency, low fuel

consumption, low NOx and PM emissions. A fuel with a low DCN, a low T90 and a

high aromatic content was a less desirable combination for engine performance. These

results confirmed that fuel formulation of ultra-low sulfur diesels played an important

role in achieving high efficiency and low emissions.

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TABLE OF CONTENTS

List of Figures .................................................................................................................... vi

List of Tables ................................................................................................................... xiii

Acknowledgements ......................................................................................................... xvii

Introduction ........................................................................................................ 1 Chapter 1

Literature Review of EGR Cooler Fouling ........................................................ 7 Chapter 2

2.1 NOx Formation Chemistry ..................................................................................... 7

2.1.1 Reduction of NOx Emissions with EGR ................................................ 8

2.2 Heat Exchanger Fundamentals ............................................................................. 10

2.2.1 Overall Heat Transfer Coefficient ........................................................ 10

2.2.2 Measurement of Heat Exchanger Performance .................................... 12

2.3 Heat Exchanger Fouling ....................................................................................... 13

2.4 Sintering versus Fouling ....................................................................................... 14

2.5 Effect of Fouling on EGR Cooler Performance .................................................... 15

2.6 Factors Affecting EGR Cooler Fouling ................................................................ 17

2.6.1 Effect of Exhaust Gas Velocity on EGR Cooler Fouling .................... 18

2.6.2 Effect of Exhaust Gas Particle Size Distribution on EGR Cooler

Fouling ................................................................................................. 18

2.6.3 Effect of Fuel on EGR Cooler Deposit Properties and Fouling

Rate ...................................................................................................... 20

2.6.4 Effect of EGR Cooler Design and Material Selection on Fouling ....... 21

2.7 What Constitutes the EGR Cooler Deposits? ....................................................... 23

2.8 Deposition Mechanisms ........................................................................................ 25

2.8.1 Thermophoresis .................................................................................... 25

2.8.2 Hydrocarbon Condensation .................................................................. 28

2.8.3 Eddy Diffusion ..................................................................................... 28

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2.8.4 Turbulent Impaction ............................................................................. 29

2.8.5 Gravitational Forces ............................................................................. 30

2.9 Deposit Removal Mechanisms ............................................................................. 31

2.9.1 EGR System Fouling Control .............................................................. 33

2.9.2 Diesel Oxidation Catalysts ................................................................... 34

Literature Review of High Efficiency Combustion ......................................... 36 Chapter 3

3.1 Introduction ........................................................................................................... 36

3.2 Conventional Diesel Combustion ......................................................................... 37

3.3 Homogeneous Charge Compression Ignition Combustion................................... 39

3.4 Premixed Charge Compression Ignition Combustion .......................................... 40

3.5 Partially Premixed Combustion ............................................................................ 41

3.6 Mixed Mode Combustion ..................................................................................... 41

3.7 Reactivity Controlled Compression Ignition Combustion .................................... 42

3.8 Fumigated Fuels for Dual Fuel Combustion ......................................................... 43

3.8.1 Dimethyl Ether (DME) ........................................................................ 43

3.8.2 Propane ................................................................................................. 46

3.8.3 Propane versus Methane ...................................................................... 47

Effect of Engine Operating Conditions, Coolant Temperature and Chapter 4

Oxidation Catalyst on Morphology and Composition of Deposits from a

Fouled Automotive Exhaust Gas Recirculation Cooler ...................................... 49

4.1 Introduction ........................................................................................................... 49

4.2 Objectives for Study of EGR Cooler Fouling ....................................................... 53

4.3 Experimental ......................................................................................................... 54

4.3.1 Engine .................................................................................................. 54

4.3.2 EGR Cooler Test Rig ........................................................................... 54

4.3.3 Modifications for Catalyst Study ......................................................... 58

4.3.4 Fuel ....................................................................................................... 59

4.3.5 Emissions Measurement....................................................................... 60

4.3.6 Analytical Techniques .......................................................................... 60

4.3.7 Engine Test Conditions ........................................................................ 63

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4.4 Results and Discussion ......................................................................................... 66

4.4.1 Effect of Engine Cruise and Near-Idle Conditions on EGR Cooler

Fouling ................................................................................................. 66

4.4.2 Effect of Coolant Temperature on EGR Cooler Fouling ..................... 79

4.4.3 Role of Water Vapor Condensation on EGR Cooler Recovery ........... 87

4.4.4 Effect of Engine Startup and Shutdown on EGR Cooler Recovery .... 90

4.4.5 Effect of EGR Oxidation Catalyst on Temperature and

Effectiveness Change for Cruise and Near-idle Conditions ................ 95

4.5 Conclusions ......................................................................................................... 105

Experimental Studies of High-Efficiency Combustion with Fumigation Chapter 5

of Liquefied Fuels into Diesel Engine Air Intake ............................................. 107

5.1 Introduction ......................................................................................................... 107

5.2 Hypothesis for High Efficiency Diesel Combustion .......................................... 110

5.3 Objectives for High Efficiency Diesel Combustion ........................................... 110

5.4 Experimental ....................................................................................................... 111

5.4.1 Engine ................................................................................................ 111

5.4.2 Fuel ..................................................................................................... 112

5.4.3 Emissions ........................................................................................... 113

5.4.4 Test Matrix ......................................................................................... 114

5.5 Results and Discussion ....................................................................................... 115

5.5.1 Effect of DME and Propane Fumigation on BTE and BSEC ............ 115

5.5.2 Effect of DME and Propane Fumigation on Emissions ..................... 129

5.6 Conclusions ......................................................................................................... 137

Performance Evaluation of Ultra-Low Sulfur Diesels ................................... 139 Chapter 6

6.1 Introduction ......................................................................................................... 139

6.2 Objective ............................................................................................................. 145

6.3 Experimental ....................................................................................................... 147

6.3.1 Engine ................................................................................................ 147

6.3.2 Particle Size and Distribution Measurements .................................... 147

6.3.3 Test Conditions .................................................................................. 148

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6.3.4 Notes on Experimental Conditions and Results ................................. 149

6.4 Results and Discussion ....................................................................................... 149

6.4.1 Effect of Fuel Properties on BSFC and BTE ..................................... 149

6.4.2 Effect of Fuel Properties on Emissions .............................................. 154

6.4.3 Effect of Fuel Properties on Particle Size and Distribution ............... 161

6.4.4 Effect of Fuel Properties on Apparent Heat Release Rate ................. 163

6.5 Conclusions ......................................................................................................... 165

Conclusions and Suggestions for Future Work .............................................. 167 Chapter 7

7.1 Summary ............................................................................................................. 167

7.1.1 Conclusions from EGR Cooler Fouling Study .................................. 167

7.1.2 Conclusions from High Efficiency Combustion Study ...................... 168

7.1.3 Conclusions from Fuel Impacts on Engine Performance ................... 170

7.2 Suggestions for Future Work .............................................................................. 170

7.2.1 Suggestions for Future Work on EGR Cooler Fouling ...................... 170

7.2.2 Suggestions for Future Work on Dual Fuel High Efficiency

Combustion ........................................................................................ 171

References ....................................................................................................................... 173

Appendix A: Fuel Specifications .................................................................................... 193

Appendix B: Preliminary Calculations, Repeatability Studies, and Error Bars in

Measurements .................................................................................................... 194

Appendix C: Calibration of High Temperature Flowmeter and Matheson

Flowmeter .......................................................................................................... 202

Appendix D: Calculation of Heat Release Profiles from Pressure Traces ...................... 205

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LIST OF FIGURES

Figure 2.1: Effect of EGR rate on NOx emissions under different engine loads [22] ........ 9

Figure 2.2: Effect of EGR cooling on NOx and PM emissions [23] .................................. 9

Figure 2.3: Fouled EGR cooler tubes at the inlet.............................................................. 13

Figure 2.4: Change in the a) heat exchanger effectiveness, b) thermal resistance, c)

pressure drop, d) mass flow rate with time; (mass flow rate of exhaust =

8.2kg/hr, inlet temperature = 250°C, PM concentration = 130 mg/m3) [35] ..... 15

Figure 2.5: Performance degradation of EGR cooler with n-dodecane injection at

various coolant temperatures [51]...................................................................... 19

Figure 2.6: Particle size distribution upstream and downstream of the EGR cooler

at various coolant temperatures [52] .................................................................. 20

Figure 2.7: Effect of tube surface coating on EGR cooler effectiveness loss and

deposit mass gain [59] ....................................................................................... 23

Figure 2.8: Variation of thermophoretic coefficient with Knudsen number for

Brock-Talbot and MCMW correlations; Brock-Talbot correlation over

predicts the thermophoretic coefficient after Kn >2 [74] .................................. 27

Figure 2.9: Comparison of deposition velocities for submicron particles [74] ................ 31

Figure 3.1: Conventional diesel heat release profile [18] ................................................. 37

Figure 4.1: Degradation of EGR cooler performance as a function of time [142] ........... 51

Figure 4.2: Conventional EGR cooler used in Ford “Powerstroke” engine ..................... 55

Figure 4.3: Model EGR cooler with 6 surrogate tubes ..................................................... 56

Figure 4.4: Test rig with in-house EGR cooler, flowmeter, high temperature valve

and recirculating chiller. Image is not to scale. ................................................. 57

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Figure 4.5: Engine test cell of EGR cooler test rig, showing the EGR cooler, high

temperature flowmeter and valve, pressure and thermocouple

instrumentation. Recirculating chiller is not shown in this image. .................... 58

Figure 4.6: Ford 6.4L engine drawing showing the EGR oxidation catalyst and dual

EGR coolers [85] ............................................................................................... 59

Figure 4.7: Effect of engine operating condition on temperature profiles, 2150

rpm, 203 Nm, exhaust, 2150 rpm, 203 Nm, EGR inlet, 2150 rpm,

203 Nm, EGR outlet, ●1400 rpm, 81 Nm, exhaust, 1400 rpm, 81 Nm,

EGR inlet, ▲ 1400 rpm, 81 Nm, EGR outlet ................................................... 67

Figure 4.8: Time varying effect of engine operating conditions on EGR cooler

effectiveness change, 2150 rpm, 203 Nm, 1400 rpm, 81 Nm ................... 68

Figure 4.9: Time varying effect of engine operating conditions on deposit mass,

2150 rpm, 203 Nm, 1400 rpm, 81 Nm........................................................... 69

Figure 4.10: Variation of deposit microstructure (low magnification) as a function

of time, a) 1.5 hours, b) 3.0 hours, c) 4.5 hours, d) 6.0 hours, and e) 7.5

hours................................................................................................................... 70

Figure 4.11: Variation of deposit microstructure (high magnification) as a function

of time, a) 1.5 hours, b) 3.0 hours, c) 4.5 hours, d) 6.0 hours, and e) 7.5

hours, and f) 9.0 hours ....................................................................................... 72

Figure 4.12: Py-GC chromatographs of species eluted as a function of time, a) 1.5

hours, b) 3.0 hours, c) 4.5 hours, d) 6.0 hours, e) 7.5 hours, and f) 9.0

hours................................................................................................................... 73

Figure 4.13: Variation of aromatics and aliphatics percentage as a function of time

for 2150 rpm, 203 Nm engine condition, , 1.5 hours, 3.0 hours, 4.5

hours, 6.0 hours, 9.0 hours, and data unavailable for 7.5 hours

condition ............................................................................................................ 74

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Figure 4.14: Effect of engine operating condition on EGR cooler deposit

microstructure, a) 2150 rpm, 203 Nm, b) 1400 rpm, 81 Nm ............................. 76

Figure 4.15: Py-GC chromatographs of species eluted as a function of engine

operating condition, a) 2150 rpm, 203 Nm, b) 1400 rpm, 81 Nm ..................... 77

Figure 4.16: Variation of aromatics and aliphatics percentage as a function of

engine operating condition, 2150 rpm, 203 Nm, 1400 rpm, 81 Nm ...... 77

Figure 4.17: Effect of engine operating condition on volatile organic fraction of

deposits after 9 hours test, 2150 rpm, 203 Nm, 1400 rpm, 81 Nm ........... 79

Figure 4.18: Effect of engine operating condition on temperature profiles, 85°C,

exhaust, 85°C, EGR inlet, 85°C, EGR outlet, ● 40°C Nm,

exhaust, 40°C, EGR inlet, ▲ 40°C, EGR outlet ......................................... 81

Figure 4.19: Time varying effect of coolant temperature on EGR cooler

effectiveness change, 85°C, 40°C ............................................................. 82

Figure 4.20: Time varying effect of coolant temperature on the mass of deposits,

85°C, 40°C ..................................................................................................... 83

Figure 4.21: Effect of coolant temperature on deposit microstructure, a) 85°C

coolant, b) 40°C coolant .................................................................................... 84

Figure 4.22: Py-GC chromatographs of species eluted as a function of coolant

temperature, a) 85°C coolant, b) 40°C coolant .................................................. 85


coolant temperature, 85°C coolant, 40°C coolant .................................... 85

Figure 4.24: Effect of coolant temperature on volatile organic fraction from

deposits, 85°C, 40°C ................................................................................. 87

Figure 4.25: 5 minute time interval snapshots of deposit layer removal due to water

vapor condensation; shiny regions reveal the metal surface [80] ...................... 89

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Figure 4.26: 5 minute time interval snapshots showing water vapor condensate

forming droplets below the deposit layer and the subsequent removal of

the deposit layer [80] ......................................................................................... 90

Figure 4.27: Temperature profiles for EGR cooler recovery test, Engine exhaust

temperature, EGR inlet temperature, EGR outlet temperature ................. 92

Figure 4.28: Effectiveness change versus coolant temperature at engine start-up ........... 93

Figure 4.29: Effect of EGR oxidation catalyst on temperature profiles at 2150 rpm,

203 Nm, without catalyst: engine exhaust, EGR inlet, EGR

outlet, with catalyst: engine exhaust, EGR inlet, ▲ EGR outlet ............... 96


81 Nm load, without catalyst: engine exhaust, EGR inlet, , EGR

outlet, with catalyst: engine exhaust, EGR inlet, ▲ EGR outlet ............... 96


effectiveness change at 2150 rpm, 203 Nm, without ECAT, with

ECAT ................................................................................................................. 97


effectiveness change at 1400 rpm, 81 Nm, without ECAT, with

ECAT ................................................................................................................. 98

Figure 4.33: Effect of ECAT on effectiveness change at high speed condition [166] ..... 99

Figure 4.34: Effect of engine operating condition on average deposit mass,

without ECAT, with ECAT ......................................................................... 100

Figure 4.35: Effect of engine operating condition on deposit microstructure, a)

2150 rpm, 203 Nm without ECAT, b) 2150 rpm, 203 Nm with ECAT, c)

1400 rpm, 81Nm without ECAT, d) 1400 rpm, 81 Nm with ECAT ............... 101

Figure 4.36: Variation of aromatics and aliphatics percentage at 2150 rpm, 203 Nm,

without ECAT, with ECAT .................................................................... 102

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without ECAT, with ECAT .................................................................... 103

Figure 4.38: Total ion count with and without ECAT [148] .......................................... 103

Figure 4.39: Catalyst removal efficiency as a function of aliphatic chain length

[166] ................................................................................................................. 104

Figure 5.1: Φ – T map showing soot and NOx formation zones, with advanced

combustion modes, adapted from Dec [170] ................................................... 108

Figure 5.2: Custom intake air manifold system for DME and propane fumigation ....... 113

Figure 5.3: Brake thermal efficiency at varying DME and propane substitution

levels (D=DME, P=Propane) ........................................................................... 117

Figure 5.4: Brake specific energy consumption at varying DME and propane

substitution levels (D=DME, P=Propane) ....................................................... 118

Figure 5.5: Indicated thermal efficiency at varying DME and propane substitution


Figure 5.6: Frictional power at varying DME and propane substitution levels

(D=DME, P=Propane) ..................................................................................... 119

Figure 5.7: Volumetric efficiency at varying DME and propane substitution levels

(D=DME, P=Propane) ..................................................................................... 119

Figure 5.8: Cylinder pressure vs. crank angle for 0% DME substitution and 0 –

40% propane substitution (D=DME, P=Propane) ........................................... 121







xv

Figure 5.12: Heat release rate vs. crank angle for 0% DME substitution and 0 –








Figure 5.16: Bulk-averaged cylinder temperature vs. crank angle for 0% DME

substitution and 0 – 40% propane substitution (D=DME, P=Propane) ........... 127







Figure 5.20: Total hydrocarbon emissions at varying DME and propane substitution


Figure 5.21: NOx emissions at varying DME and propane substitution levels

(D=DME, P=Propane) ..................................................................................... 133

Figure 5.22: CO emissions at varying DME and propane substitution levels ................ 135

Figure 5.23: Air-Fuel ratio at varying DME and propane substitution levels ................ 136

Figure 5.24: CO2 emissions at varying DME and propane substitution levels ............... 136

Figure 6.1: Thermodynamic comparisons of available fuel energy [186] ...................... 141

Figure 6.2: Layout of Matrix 1 fuels ............................................................................... 145

Figure 6.3: Engine operating conditions ......................................................................... 148

xvi

Figure 6.4: Effect of fuel properties on BSFC at a) 1400 rpm, 20% load, b) 1400

rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20% load .................... 150

Figure 6.5: Effect of fuel properties on BTE at a) 1400 rpm, 20% load, b) 1400

rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20% load .................... 151

Figure 6.6: Effect of fuel properties on fuel consumption (L/min) at a) 1400 rpm,

20% load, b) 1400 rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm,

20% load .......................................................................................................... 153

Figure 6.7: Effect of fuel properties on BSNOx emissions at a) 1400 rpm, 20%

load, b) 1400 rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20%

load................................................................................................................... 156

Figure 6.8: Effect of fuel properties on BSPM emissions at a) 1400 rpm, 20% load,

b) 1400 rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20% load ....... 158

Figure 6.9: Effect of fuel properties on BSHC emissions at a) 1400 rpm, 20% load,

b) 1400 rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20% load ....... 160

Figure 6.10: Effect of fuel properties on BSCO emissions at a) 1400 rpm, 20%

load, b) 1400 rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20%

load................................................................................................................... 161

Figure 6.11: Effect of fuel properties on particle size and distribution at a) 1400

rpm, 20% load, b) 1400 rpm, 60% load, c) 2000 rpm, 45% load, d) 2800

rpm, 20% load .................................................................................................. 163

Figure 6.12: Effect of fuel properties on apparent heat release at a) 1400 rpm, 20%

load, b) 1400 rpm, 60% load, c) 2000 rpm, 45% load ..................................... 164

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LIST OF TABLES

Table 3.1: Physical properties of diesel, DME, propane and methane [112,113] ............ 44

Table 4.1: Engine specifications ....................................................................................... 54

Table 4.2: ChevronPhillips Chemical ULS 2007 diesel fuel properties ........................... 59

Table 4.3: TGA procedure to determine VOF [155] ........................................................ 63

Table 4.4: Cruise and near-idle operating conditions ....................................................... 64

Table 4.5: Effect of time on elemental composition of deposits ...................................... 75

Table 4.6: Effect of engine condition on elemental composition of deposits ................... 78

Table 4.7: Effect of coolant temperature on elemental composition of deposits .............. 86

Table 4.8: Test procedure for EGR cooler recovery monitoring ...................................... 91

Table 4.9: Starting and ending EGR cooler effectiveness ................................................ 94

Table 5.1: Engine specifications ..................................................................................... 112

Table 5.2: Specifications of ultra-low sulfur diesel fuel ................................................. 113

Table 5.3: Percentage of DME and propane energy fumigated into engine air intake ... 115

Table 6.1: Fuel Matrix .................................................................................................... 146

Table 6.2: Summary of the effect of fuel properties on engine performance ................. 166

xviii

Abbreviations

BSFC Brake specific fuel consumption

BSEC Brake specific energy consumption

BTE Brake thermal efficiency

CAFÉ Corporate average fuel economy

CEGR Cooled exhaust gas recirculation

CI Compression ignition

CRC Coordinating research council

DBTP di-tert-butyl peroxide

DME Dimethyl Ether

DOC Diesel oxidation catalyst

DPF Diesel particulate filter

EOC/ECAT EGR oxidation catalyst

EGR Exhaust gas recirculation

FACE Fuels for advanced combustion engines

GC Gas chromatography

HC Hydrocarbons

HCCI Homogenous charge compression ignition

HTHR High temperature heat release

LPG Liquefied Petroleum Gas/Propane

LTHR Low temperature heat release

MS Mass spectroscopy

xix

NOx Nitrogen oxide emissions

PCCI Premixed charge compression ignition

PM Particulate matter

PPC Partially premixed combustion

Py-GC Pyrolysis gas chromatography

RCCI Reactivity controlled compression ignition

SEM Scanning electron microscope

SI Spark ignited

SOF Soluble organic fraction

T90 Temperature at 90 volume % distilled

TDC Top dead center

THC Total hydrocarbon emissions

TWC Three-way catalyst

UHC Unburned hydrocarbons

ULEV Ultra-low emissions vehicle

ULSD Ultra-low sulfur diesel

VOF Volatile organic fraction

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Nomenclature

P Pressure (bar)

Scp Particle Schmidt number

T Gas Temperature (K)

U Overall heat transfer coefficient (W/m2K)

dp Particle diameter (nm)

g gravity (N/kg)

ℰ Thermal effectiveness (%)

λ Mean free path (nm)

kg Gas thermal conductivity (W/mK)

kp Particle thermal conductivity (W/mK)

ν kinematic viscosity (m2/s)

τ particle relaxation time

u friction velocity (m/s)

xxi

ACKNOWLEDGEMENTS

Doctoral program at any graduate school is not just about academics, but about

one’s journey to self-realization and commitment to find solutions for even the simplest

of problems. This dissertation reflects my scholarly activity and the hard work for the

past several years. This however, would not have been possible without the guidance and

support of so many people I have been involved with.

First and foremost, I thank my adviser Dr. André Boehman, for giving me an

opportunity to work under him and mentoring me through-out my graduate program at

PSU. I have admired him for his optimism even under the most stressful situations. It is

through him that I have learned to be more patient and productive.

I would like to specially thank Dr. Randy Vander Wal for agreeing to be my co-

advisor due to certain change of events at PSU and also thank my committee members

Dr. Jonathan Mathews and Dr. Daniel Haworth. Their guidance over the years has been

excellent.

Special thanks to Vincent Zello at the Diesel Combustions Lab, for keeping the

engines running at all times. I have spent countless number of hours working and

learning from him so many fundamental concepts of engineering, which has given me the

confidence to do experimental research. From the same lab, I would like to thank the

support from Dr. Stephen Kirby, Dr. Kuen Yehliu, Dr. Hee Je Seong, Dr. Peng Ye, Dr.

Gregory Lilik, Vickey, Eduardo, Claire and Dongil. At the Energy Institute, I also thank

Dr. Dania Fonseca and Ronald Wasco for help with the deposit analysis. Additionally,

xxii

the staff on the EI Bridge (Kelly, Nicky, Cindy, Danielle, Erin) was very helpful in

keeping up with orders and tracking official requirements.

I would also like to thank GE Global Research Center and GE Transportation, in

particular, David Walker and David Watson, for supporting my work on EGR cooler

fouling. I would like to thank Amar Pascal and Samuel McLaughin from Volvo Truck

Technology for providing the support for the work on high efficiency combustion. I

would also like to thank Krystal Wrigley from ExxonMobil for providing fuel for some

of the experiments.

Through my life as a graduate researcher, I made several friends who carried me

through to the finish line. In particular, I would like to thank Dr. Venkatesh Iyer for his

constant support and motivation and wish him the best in his career.

I would like to thank my sister Kavita Prabhakar, my brother-in-law Karthik and

their kid Gowri for all the encouragement. My niece, Gowri, has been the reason for

constant smile and happiness. Last but not the least, I would like to thank my parents,

L.D Prabhakar and Gayathri Prabhakar, who believed in me, and motivated me at every

instance of life. They have instilled in me the confidence to be a successful researcher,

and more than anything, molded me into a humble and modest individual ready to accept

life with passion and enthusiasm.

xxiii

DEDICATION

To amma and daddy, you know how much this means to me!

1

Chapter 1

Introduction

Automobiles have been an indispensable means of transportation for our modern

society. Most on-road vehicles have reciprocating engines which are either compression-

ignited (CI) diesel engines or the spark-ignited (SI) gasoline engines. Diesel engines

typically offer higher thermodynamic efficiency and increased fuel economy compared to

gasoline engines, and hence continues to gain interest from consumers. The United States

which largely relies on petroleum based fuels for its transportation needs (15 million

barrels oil consumed per day), is among the highest emitters of greenhouse gases (GHG),

and has the highest per capita emissions of GHG [1]. With increasing concern about

rising fuel prices, limited petroleum fuel supplies, and gaseous emissions, the call for

advanced, efficient and cleaner engines is louder than ever.

Exhaust emissions from both diesel and gasoline engines contain a wide variety of

components including carbonaceous particulate matter (PM), oxides of nitrogen (NOx)

and unburned hydrocarbons (UHC). These emissions have an adverse effect on health

and the environment [2–5]. For the heavy-duty diesel engines, the US 2010 emissions

regulations required NOx and PM emissions to be under 0.2g/bhp-hr and 0.01 g/bhp-hr,

respectively [6], which very few of the engines were able to meet without extensive

exhaust aftertreatment systems.

2

“Clean Diesel” which was once considered an oxymoron, is now more practical

and realistic. Clean Diesel technology is a system of three key parts: clean diesel fuel

(focus on ultra-low sulfur diesel), efficient engines (technologically advanced), and

effective emissions control technology (aftertreatment). Each of these parts plays a major

role in achieving high efficiency and low emissions from diesel engines.

From a fuel standpoint, introduction of advanced diesel technology in 2007 relied

on ultra-low sulfur diesel fuel (97% reduction in fuel sulfur content) which reduced PM

and NOx emissions by over 98% in the heavy-duty engine segment compared to the year

2000 engine models [7]. Even though the composition of diesel fuel traditionally exerted

a modest influence on engine efficiency compared to the engine design, recent research

has shown that advanced combustion engines, which show promise for high efficiency

and low emissions with modest or no aftertreatment systems, are sensitive to fuel

properties. It is widely agreed that the most important fuel properties in this regard are

cetane number, aromatic content, and volatility [8–11].

From an engine standpoint, there has been extensive research to develop advanced

combustion strategies to move away from the stratified diesel combustion which

inherently has a low efficiency [12]. This has been made possible with the use of

electronic control and advanced fuel injection systems. Some of the high efficiency

combustion modes commonly discussed in the literature are the homogenous charge

compression ignition (HCCI), premixed charge compression ignition (PCCI), partially

premixed combustion (PPC), and reactivity controlled compression ignition (RCCI).

These advanced combustion modes include fuel injection at high pressures, combustion

3

at low temperatures, varying degrees of fuel/air mixing, exhaust gas recirculation (EGR),

and/or multiple fuels.

From an aftertreatment standpoint, several systems such as diesel oxidation

catalysts (DOC), selective catalytic reduction (SCR), lean NOx traps (LNT), NOx

absorbers, diesel particulate filters (DPF), etc. have been developed to mitigate engine-

out UHC, NOx and PM emissions, since in-cylinder emissions control is challenging

while burning diesel fuel. These systems have become more effective ever since the

switch to ULSDs has been made, as it was found that sulfur had a tendency to poison the

catalyst’s activity [7].

Even with these merits, Clean Diesel technology is plagued with several

challenges. For one, there is a heavy dependence on crude oil for diesel fuel production.

The United States Energy Information Administration predicted in 2006 that world

consumption of oil will increase to 98.3 million barrels per day (mbd) in 2015 and 118

mbd in 2030 [13]. Hence from an energy security standpoint, there is an imminent need

to identify alternate fuels which will burn well in the existing systems (i.e., drop-in fuels)

so that there is no need for an infrastructure overhaul.

Secondly, the use of aftertreatment systems to reduce emissions has been found to

rob the engine’s fuel economy and overall efficiency. For example, an engine relying on

LNT for NOx reduction must periodically operate rich to reduce the stored NOx; thus

reducing the fuel economy [14,15]. Similarly, DPF’s for particulate control require

periodic regeneration, which often leads to an increase in the exhaust back-pressure and

fuel consumption [16,17]. Hence, it’s clear that for maximizing overall engine efficiency,

4

an engine should minimize the need for aftertreatment systems. Advanced combustion

strategies shows promises of in-cylinder NOx and PM reduction. Most of these strategies

rely on high EGR rates to reduce NOx emissions and long ignition delay times to allow

adequate fuel-air mixing prior to combustion to reduce PM emissions. Modern day

engines, which are equipped with EGR coolers to reduce NOx emissions further, are

subjected to particulate deposition inside the coolers which lead to fouling and plugging.

Fouled EGR coolers reduce the heat transfer effectiveness, typically on the order of 30-

40%, and increase the pressure drop which forces the engine to do more work to obtain

the same cooling, leading to a fuel economy penalty. EGR cooler fouling has been

recognized as a significant problem in the auto industry that needs immediate attention.

Finally, engine efficiency is not entirely governed by the engine design and the

combustion strategy. It is now being understood that in modern engines, fuel formulation

plays an important role in achieving the high efficiencies and mandated emissions [8].

One such example was the formation of the FACE (fuels for advanced combustion

engines) working group [9]. The fuels were designed to vary independently in the

properties of cetane number, aromatic content, and distillation temperature (T90). Results

suggested that lower cetane number fuels were desirable for improved fuel efficiency, but

resulted in higher NOx emissions. Fuels with a high T90 resulted in high CO and HC

emissions. Aromatic content did not appear to have any effect on fuel economy and

emissions. These observations have triggered research to determine if fuels should evolve

to meet future requirements for the advanced combustion strategies.

5

The goal of this dissertation is to understand three crucial challenges faced in the

Clean Diesel technology: 1) Understanding the phenomenon of fouling in EGR coolers to

develop fouling mitigation strategies, 2) Achieving high efficiency combustion by

utilizing alternate fuels (DME and propane) along with conventional diesel fuel and 3)

Addressing the synergistic role of fuel formulation (cetane number, aromatic content, and

distillation temperature) on engine efficiency and emissions.

Chapters 2 and 3 present a comprehensive literature review about EGR cooler

fouling and advanced combustion strategies for diesel engines, respectively.

Experimental setup, hypotheses and objectives for each piece of work are discussed in

their respective chapters.

In Chapter 4, the effect of engine operating conditions and coolant temperature on

EGR cooler fouling are discussed for two engine conditions: cruise and near-idle. Two

coolant temperatures, 85°C and 40°C, were selected based on the current on-road

standards. Preliminary results from the experiments suggested that water condensation

plays an important role in the recovery of EGR coolers, and hence experiments were

performed to understand its role especially during engine start and stop cycles. This

chapter concludes with a study of oxidation catalysts for reducing EGR cooler fouling.

Chapter 5 explores the role of ignition quality of fumigated fuels on improving

combustion phasing and efficiency in a diesel engine. Two fuels: DME and propane were

fumigated into the diesel engine air intake, to simulate a process similar to dual fuel

combustion. The engine was operated at a single speed and load condition under fixed

start of injection and without EGR. The concentrations of DME and propane were varied

6

over a span of 0 to 60% diesel energy equivalent. The engine performance was evaluated

on the basis of brake thermal energy (BTE), brake-specific energy consumption (BSEC),

pressure rise rate, heat release rate, and emissions.

Chapter 6 explores the role of fuel formulation on engine efficiency, emissions,

and particle size distribution. Three key fuel properties viz. cetane number, aromatic

content and distillation temperature were investigated. Seven different ULSD fuels, each

varying in the three mentioned fuel parameters are tested, and the best performing fuel

was suggested.

Chapter 7 provides conclusions and recommendations for future work based on

all the experimental results observed in different studies.

7

Chapter 2

Literature Review of EGR Cooler Fouling

2.1 NOx Formation Chemistry

While nitric oxide (NO) and nitrogen dioxide (NO2) are usually grouped together

and referred to as NOx emissions, NO is the predominant oxide produced inside the

engine cylinder [18]. Current petroleum fuels do not contain significant levels of nitrogen

and hence the principal source of NO is the oxidation of atmospheric nitrogen. Although

a large number of reactions and reaction pathways have been found to participate in NOx

formation process [19,20], the key reactions at the pressure, temperature, and time scales

of engines are identified together as the extended Zeldovich Mechanism [21].

where,

→ (2.1)

→ (2.2)

→ (2.3)

[

( )] (2.4)

[

( )] (2.5)

[

( )] (2.6)

8

If the relevant time scales are sufficiently long, one can assume that the N2, O2, O,

and OH concentrations are at their equilibrium values and N atoms are in steady state.

This yields the following rather simple rate expression as seen in Equation (2.7). As

observed, NO formation has a large dependence on temperature (since k1f has large

activation energy) and the availability of oxygen at equilibrium.

2.1.1 Reduction of NOx Emissions with EGR

The previous section outlined that NOx formation is governed by the availability

of oxygen for combustion and the mean temperature inside the cylinder. Hence ways to

reduce NOx emissions in-cylinder would be to lower the oxygen concentration and the

mean flame temperature. Exhaust gas recirculation (EGR) is one such technique which

achieves both without major engine modifications. An example of NOx emissions

reduction with EGR at three engine loads is shown in Figure 2.1. From this figure, it can

be observed that NOx emissions increase with engine load due to high in-cylinder

temperatures, and decreases with increase in EGR rate. Reduction of NOx emissions with

EGR happens in two ways.

Dilution mechanism: Addition of burned combustion gases into the air intake

potentially increases the mixing time and burn duration due to the dilution of intake

air. In addition, EGR decreases the concentration of oxygen in the intake air which

reduces the flame temperature leading to lower NOx emissions.

[ ]

[ ] [ ] (2.7)

9

Thermal mechanism: The CO2 and H2O content of the exhaust gas acts as a heat sink

and reduces the adiabatic flame temperature which in turn reduces NOx emissions.

Figure 2.1: Effect of EGR rate on NOx emissions under different engine loads [22]

Cooling the exhaust gas prior to mixing with fresh intake air reduces NOx

emissions further, as shown in Figure 2.2. EGR cooling is achieved through heat

exchangers known as EGR coolers. A simple ‘back-of-the-envelope’ thermodynamic

calculation can explain the reduction in the adiabatic temperature of the gas (hence NOx

emissions) on cooling, but is not discussed here.

Figure 2.2: Effect of EGR cooling on NOx and PM emissions [23]

Increasing load

10

2.2 Heat Exchanger Fundamentals

The process of heat exchange between two fluids that are at different temperatures

and separated by a solid wall occurs in many engineering applications. The device used

to implement this exchange is termed a heat exchanger [24]. They are usually categorized

based on the flow arrangement and type of construction. Common types of heat

exchanger flow configurations include parallel flow, counter flow, and cross flow. In

parallel flow, both fluids move in the same direction while transferring heat; in counter

flow, the fluids move in opposite directions; and in cross flow, the fluids move at right

angles to each other. Heat exchanger designs include shell and tube, double pipe,

extruded finned pipe, spiral fin pipe, u-tube, and stacked plate. Shell and tube heat

exchangers operating under counter flow arrangement are commonly used as EGR

coolers in automotive applications to lower the temperature of the exhaust gas before it

mixes with fresh intake air. Counter flow arrangements have a higher heat transfer rate

compared to parallel flow arrangements for the same size and heat transfer surface area.

EGR coolers are typically cooled by the engine coolant, whose temperature is around 85-

90°C. The size of the EGR cooler depends on several factors such as the cooling capacity

required, operating pressure, and the space available for packaging.

2.2.1 Overall Heat Transfer Coefficient

An important parameter in heat exchanger analysis is the determination of the

overall heat transfer coefficient (U), which is a measure of the overall ability of a series

11

of conductive and convective barriers to transfer heat. For a heat exchanger, U can be

used to determine the total heat transfer between the two non-mixing streams.

where,

q: heat transfer rate (W)

A: heat transfer surface area (m2)

U: overall heat transfer coefficient (W/m2K)

ΔTLM: logarithmic mean temperature difference (K)

The overall heat transfer coefficient takes into account the individual heat transfer

coefficients of each stream and the resistance of the material. It can be calculated as the

reciprocal of the sum of a series of thermal resistances as shown in Equations (2.9) and

(2.10).

where,

R: resistance to heat flow (K/W)

x: length (m), k: thermal conductivity of the material (W/mK)

h: heat transfer coefficient (W/m2K)

(2.8)

(2.9)

(2.10)

12

2.2.2 Measurement of Heat Exchanger Performance

Thermal effectiveness ( ) is defined as the ratio of the actual heat transfer to the

maximum possible heat transfer [24], as shown in Equation (2.11). A heat exchanger is

100% effective if the gas temperature at the outlet is equal to the coolant temperature at

the inlet.

(2.11)

Pressure differential (Δp) is the difference between the pressures upstream and

downstream of the EGR cooler. As deposits build inside the EGR cooler, the effective

diameter of the cooler tubes reduces and the pressure differential increases due to

physical blockage of the gas flow passages. The pressure differential across the cooler is

an important design variable, as it influences the energy required for a given amount of

EGR flow, which in turn has an influence on the overall engine efficiency, and is

calculated as shown in Equation (2.12).

where,

ρ: fluid density (kg/m3), f: friction factor

L: tube length (m), D: tube diameter (m)

V: average velocity in the tube (m/s)

(

) (

) (2.12)

13

2.3 Heat Exchanger Fouling

The accumulation of unwanted deposits on the heat transfer surfaces of a heat

exchanger is referred to as fouling. These deposits add extra thermal resistance to heat

flow and may noticeably decrease the overall heat transfer coefficient and performance,

in addition to other problems like plugging [24]. Fouling may be due to biological

material, the products of chemical reaction including corrosion, or particulate matter.

Fouling can occur as a result of fluids being handled and their constituents in

combination with the operating conditions such as temperature and velocity.

Automotive heat exchangers are exposed to exhaust gas comprised of particulate

matter, unburned hydrocarbons and water vapor. PM and HCs from the exhaust gas

migrate to the walls of the heat exchanger and form an insulating layer resulting in the

loss of thermal effectiveness. Additional deposits can get caught on the existing layer

leading to a growth of the insulating layer, and plug the heat exchanger tubes in the

extreme case. Plugging leads to a catastrophic engine failure. One example of a fouled

EGR cooler is shown in Figure 2.3.

Figure 2.3: Fouled EGR cooler tubes at the inlet

14

The additional thermal resistance due to fouling can be found by comparing the

overall heat transfer coefficient determined from laboratory readings with calculations

based on theoretical correlations as shown in Equation (2.13).

where,

U: Overall heat transfer coefficient for heat exchanger (W/m2K)

Rf: thermal resistance due to fouling (W/m2K)

2.4 Sintering versus Fouling

Fouling of heat exchange surfaces is common in gasifiers, boilers, incinerators,

etc. [25–28]. Sintering is defined as a process of forming a coherent mass from powders

by heating the powder without melting it. Sintering takes place if the surface temperature

of the fouling layer exceeds a certain limit, known as the minimum sintering temperature

[29], and this temperature is usually below the melting point of the fouling layer material

[30,31]. Sintering leads to the reduction of the void volume and reinforcement of the

contact bridges between the particles of the fouling layer, and is therefore responsible for

strengthening of the fouling layer [32,33]. The reduction in porosity also results in

increased thermal conductivity as observed while burning coal in a gasifier [34]. Typical

melting point of carbon (soot) is around 3800K. Sintering is not observed in EGR coolers

as the gas side temperature does not exceed 350-400°C in the heat exchanger.

(2.13)

15

2.5 Effect of Fouling on EGR Cooler Performance

Zhang et al. [35] investigated the effect of diesel soot deposition on the

performance of a small 6-tube shell and tube heat exchanger by operating the engine at

medium load which produced an exhaust gas temperature of around 250 °C. They found

that fouling increased the thermal resistance and pressure drop by 150% during 12 hours

of exposure. The rate of increase of the thermal resistance decreased over this period and

approached an asymptotic value as shown in Figure 2.4.

Figure 2.4: Change in the a) heat exchanger effectiveness, b) thermal resistance, c)

pressure drop, d) mass flow rate with time; (mass flow rate of exhaust = 8.2kg/hr,

inlet temperature = 250°C, PM concentration = 130 mg/m3) [35]

In on-road testing performed at freeway cruise condition (2280 rpm and 8.2 brake

mean effective pressure), Mulenga et al. observed that the effectiveness of the EGR

cooler reduced by about 10-30% in the first 5 hours of operation due to fouling [36].

16

Additionally, after a brief shutdown period of 30 minutes, the effectiveness improved by

about 10%, but dropped again in the following 5 hours. The reason for effectiveness

improvement was not clear and was assumed to be a combination of shut-down cooling,

rapid heat-up and abrasive high gas speeds. It has been observed that deposit

accumulation is worse under low speed and low load conditions, where the flow rate and

temperature of the gas is low [37].

Ismail et al. [38] used a non-destructive neutron radiography technique to measure

the thickness of diesel soot deposited in the EGR cooler tubes as a function of tube length

and Reynolds number. The analysis of the neutron images revealed that soot deposition in

the tube occurred at a faster rate for turbulent flow than for laminar flow, and the deposit

thickness decreased along the length of the tube for both the flow regimes. This is a direct

consequence of the temperature difference at the inlet and outlet of the tubes. The exhaust

gas is the hottest and coolant temperature is the lowest at the EGR inlet, while the

exhaust gas cools down and the coolant heats up at the exit of the EGR cooler. These

differences result in varying thermophoretic gradients leading to differences in deposition

inside the tubes. On the contrary, Stolz et al. [37] did not observe any differences in the

deposit layer thickness across the length of the tube. This could be because Ismail et al.

ran the test for a short duration (~5 hours) while Stolz et al. ran the tests for over 200

hours. If deposits are heavier initially, the deposit layer will create an insulating effect,

and thereby reduces the deposition rate. And thus the gas arriving downstream will be

hotter and dirtier, and the deposition profile will even out over time.

17

Future emission standards require use of high EGR rates and oversizing the EGR

cooler to obtain required cooling efficiency when dirty is not a viable option considering

the constraints in package space available in engines.

2.6 Factors Affecting EGR Cooler Fouling

Hoard et al. [39] provided a comprehensive report of factors affecting EGR cooler

fouling. For a given design of an EGR cooler, the key factors affecting fouling are

feedgas temperature, coolant temperature, gas flow rate, particulate matter and

hydrocarbon concentrations. Changes in the engine operating condition would alter the

temperature and the composition of the exhaust gas [40–44]. For example, increasing the

engine load increases the exhaust gas temperature and concentration of the soot particles

but decreases the unburned hydrocarbon concentration [18,45]. On the contrary,

operating the engine on idle results in an excess of unburned hydrocarbon emissions due

to incomplete combustion of the fuel because of low in-cylinder temperatures [46,47].

The temperature gradient imposed by the difference between the exhaust gas and

coolant temperatures produces a thermophoretic driving force that influences PM

deposition on the cooler walls. The coolant temperature influences the condensation of

gas-phase HCs on the relatively cold walls of the cooler. PM and HC concentrations are

important because they influence the deposition rate inside the EGR cooler. These

deposition mechanisms are further discussed in section 2.8.

18

2.6.1 Effect of Exhaust Gas Velocity on EGR Cooler Fouling

Bravo et al. [48] measured the variation of fouling resistance as a function of gas

velocity (Reynolds number varying between laminar to turbulent flow regimes) and

found that the fouling resistance decreases with increase in Reynolds number, which

differs from what was reported by Ismail et al. [38]. This could be a result of competition

between particle deposition and removal, as turbulent flow can promote deposition due to

diffusion in the flow and can also induce shear stresses on the deposits which can aid

deposit removal. Which of these processes is dominant is still not understood. Some

literature reported that particulate fouling can be avoided if the gas velocity is above a

critical flow velocity [49,50], which is defined as the main stream velocity above which

rolling of the deposited particles occurs. This value is a strong function of the particle

size and is found to decrease with increasing particle size [49]. The typical size of diesel

particles is in the nanometer range of 100-500nm, which makes it difficult to control the

critical flow velocity for different engine operating conditions.

2.6.2 Effect of Exhaust Gas Particle Size Distribution on EGR Cooler Fouling

Hong et al. [51] performed a parametric study on the effect of particle size and

soluble organic content on EGR cooler fouling, using a soot generator to independently

control the size of the particles (diameter range between 41 to 190 nm) and injecting

vaporized n-dodecane (0, 2, 4 mL/h) into the exhaust stream to vary the percentage of

soluble organics. They found that the deposition fractions were inversely proportional to

the particle size, indicating that smaller soot particles are more likely to cause fouling

19

than the larger ones due to greater mobility. Fouling increased with an increase in n-

dodecane injection rate, and this effect was magnified at lower coolant temperatures, as

shown in Figure 2.5, suggesting that ‘wet soot’, which is comprised of a higher soluble

organic fraction (SOF), is more likely to increase fouling than ‘dry soot’.

Figure 2.5: Performance degradation of EGR cooler with n-dodecane injection at

various coolant temperatures [51]

In a similar experiment, Bika et al. [52] analyzed the impact of nucleation and

accumulation modes of exhaust particle size distributions1 over a range of EGR coolant

temperatures and engine-out soot and HC concentrations. They observed a reduction in

the accumulation mode particle concentration at the EGR cooler outlet under high soot

concentrations as seen in Figure 2.6, which implied that soot deposition increased inside

the EGR cooler. At low coolant temperatures, the accumulation mode particles reduced

1 Nucleation mode refers to particle diameter Dp < 30 nm, accumulation mode 30 nm < Dp <500 nm, coarse

mode Dp > 500 nm [220,223]

20

further indicating a high deposition rate due to increased thermophoresis. A significant

increase in the nucleation mode particles was observed at the EGR cooler outlet for low

soot concentration exhaust gas flow. These experiments concluded that the particle size

distribution influenced the rate of fouling in the EGR coolers, and more importantly,

accumulation mode particles contributed to the increase in mass of the deposits due to

their inherent larger size.

Figure 2.6: Particle size distribution upstream and downstream of the EGR cooler

at various coolant temperatures [52]

2.6.3 Effect of Fuel on EGR Cooler Deposit Properties and Fouling Rate

Sluder and Storey [53] investigated the effect of fuel composition on EGR cooler

performance and degradation. The two fuels used were a conventional ultra-low sulfur

diesel (ULSD) and a 20% volume blend of soy bio-diesel in diesel (B20). Results showed

that B20 and ULSD fuel did not result in significantly different EGR cooler effectiveness

21

loss or deposit mass gain at the conditions studied. However, B20 deposits contained

higher fractions (~10%) of volatile components than ULSD. Additionally, higher HC

concentration in the exhaust gas resulted in higher levels of volatiles in the deposits,

regardless of the fuel used.

There seems to be no published work on the effect of gasoline or gasoline blends

on EGR cooler fouling in SI engines and this opens up a new area for research. Similarly,

the effect of renewable diesels and Fischer Tropsch fuels on diesel EGR cooler fouling

has not been investigated to date.

2.6.4 Effect of EGR Cooler Design and Material Selection on Fouling

Kim et al. [54] investigated the heat exchange effectiveness of EGR coolers with

shell and tube and stack type designs and found that stack type EGR cooler had gas outlet

temperatures of 15-35°C lower than the shell and tube type EGR cooler, which resulted

in a 25-30% improvement in thermal effectiveness. This was mainly due to increased

heat transfer surface area for the stack type design. Charles et al. [55] investigated the

heat exchanger performance for finned-plate type and shell and tube heat exchanger. The

former design has extended surfaces and is normally operated in laminar or transitional

flow regimes. Shell and tube heat exchangers are used typically in turbulent flow

regimes. They observed that the finned-plate type EGR cooler was less affected by soot

deposition compared to shell and tube design. Park et al. [56] reported on shell and tube

heat exchangers, with straight and spiral tubes and observed that the spiral tubes transfer

more heat when clean, however, they are more susceptible to fouling compared to

22

straight tubes. Usui et al. [57] investigated the effect of semi-circular micro ribs aligned

in the direction of the gas flow on particle deposition and found that the ribs could reduce

deposit accumulation by about 38%, compared to smooth and flat surfaces. The ribs

reduced surface friction and enhanced heat transfer rate. They did not observe any

additional deposition at the ribs’ valleys.

Several EGR cooler manufacturers are evaluating the role of corrugated tubes in

shell and tube heat exchangers, however, no published work exists which discusses its

effect on fouling. Corrugated tubes create a helical secondary flow that increases

turbulence and breaks-up the boundary layer, and the local recirculating flow vortices

near the walls assist in keeping the soot off the wall. Each manufacturer claims to have an

optimal set of geometries for best packaging size, effectiveness, deposit resistance, etc.

The choice of the material used in EGR coolers depends not only the material’s

thermal properties, but also on its ability to be molded into shape and the cost for large

scale production. Typical materials used in EGR coolers are stainless steel or aluminum

[58]. Sluder et al. [59] investigated the effect of tube surface treatments on soot

deposition in EGR coolers. The baseline material was grade-316 stainless steel, with the

following coatings: silica/silicone, nickel/Teflon coating, alumina-boron nitride coating,

and electro-polished 316SS. Some treatments resulted in the material being electrically

conducting and some were insulating. The authors observed no significant difference in

the effectiveness change or deposit mass gain for most surface treated tubes compared to

the base stainless steel tube, as shown in Figure 2.7. These results show that the deposit

accumulation inside the heat exchangers is not affected by the material used in the

23

construction of the EGR cooler. Considering these findings, it might be viable to go with

the cheapest option, which would be stainless steel. The detailed effect of various

geometries and material selection is still not clearly understood.

Figure 2.7: Effect of tube surface coating on EGR cooler effectiveness loss and

deposit mass gain [59]

2.7 What Constitutes the EGR Cooler Deposits?

It is well agreed that the deposits from the exhaust stream accumulate on the

surface of the EGR cooler. The deposits are black in color and can be either dry or wet

based on the relative proportions of soot, HCs, and acids that form as a function of

temperature and time [39]. In general, engine deposits (not necessarily EGR cooler

deposits) which form in a low temperature region (<200°C), are dark, and tar like

0

5

10

15

20

25

30

35

40 Effectiveness Loss, %

Depost Mass Gain, mg

24

portions are visible, as noted by Lepperhoff and Houben [60]. At temperatures between

250-300°C, the deposits are nearly dry-porous, and at temperatures >300°C, different

light colors are seen in a thin layer [39]. These are mainly due to inorganics from the fuel

and lube additives.

The deposits in the EGR cooler contain soot from the engine exhaust, which is

essentially elemental carbon with particle size in the range of 20-500 nm [61,62]. Diesel

exhaust also consists of a wide range of HCs which arise from incomplete combustion of

the fuel and the lubricating oil. The deposits can also contain traces of acids, depending

on the coolant temperature in the EGR cooler [39]. Presence of sulfates in soot particles

or direct condensation of sulfuric acid has the potential to significantly magnify corrosion

and fouling [63,64]. The dew point temperature of sulfuric acid is around 100°C, which is

in the typical operating range of the EGR cooler [63]. Girard et al. [65] collected sulfuric

acid condensates from the EGR cooler deposits, and found that the concentrations were

fairly low. The concentration of H2SO4 will be significantly lower while burning ultra-

low sulfur diesel.

Lance et al. [66] measured the specific heat (cp) and thermal diffusivity (α) of

EGR cooler deposits to calculate the deposit’s thermal conductivity (k). Using a primary

soot particle density of 1.77 g/cm3, the average density of the deposits was calculated to

be 0.035 g/cm3, which resulted in a deposit layer porosity or the void fraction (ϕ = 1- ρbulk

/ ρparticle) of 98%. The average thermal conductivity of the soot deposits was calculated to

be equal to 0.057 W/m-K, which is slightly above that for air, but much lower than

stainless steel which has a conductivity of 14.7 W/m-K. These results clearly indicate that

25

EGR cooler deposits have a very low thermal conductivity, resulting in a reduced heat

transfer effectiveness of the EGR cooler when fouled. The reason(s) why the deposit

layer is so porous is not understood, and needs further research.

2.8 Deposition Mechanisms

Deposition mechanisms relevant to automotive heat exchangers are

thermophoresis, condensation, eddy diffusion, turbulent impaction, and gravitational

forces. A summary of these mechanisms are described in the following section and the

extent to which each of these mechanisms contribute to fouling is also discussed.

2.8.1 Thermophoresis

Thermophoresis, a phenomenon driven by a thermal gradient between the hot

exhaust gas and the cold heat exchanger surface, appears to be the dominant mechanism

for particle deposition inside the EGR cooler [51,67,68]. Gas molecules around the

particles on the hot side move faster than the molecules on the cold side, and as a result a

net force is generated pushing the particle toward the cold wall. Particles reaching the

wall stick to it due to Van der Waals forces [69]. These forces arise due to warping of the

electron cloud as the particle approaches the wall surface [70]. The electrons on the wall

move away from the electrons on the particle causing a local positive charge on the wall.

When particles and tube walls have opposite charges, a net attractive force arises causing

26

the particles to stick to the wall. The Brock-Talbot correlation is commonly used to

calculate the thermophoretic velocity [71].

Brock-Talbot Correlation

The thermophoretic drift velocity of a particle in a pipe flow is defined as shown in

Equation (2.14).

where,

λ: Particle mean free path (nm), kg: Gas thermal conductivity (W/mK)

ν: kinematic viscosity (m2/s), kp: Particle thermal conductivity (W/mK)

dp: Particle diameter (nm), T: Gas temperature (K)

A, B, C, Cs, Cm, Ct are thermophoretic constants which are 1.257, 0.4, 1.1, 1.14, 1.17, and

2.18 respectively.

Knudsen number (Kn), shown in Equation (2.17), is a dimensionless number

defined as the ratio of the molecular mean free path length to a representative physical

length scale.

(2.14)

( )

( )

(2.15)

( ) (2.16)

27

When the particle diameter (dp) equals the mean free path (λ) of the gas

molecules, the Knudsen number is 2. The Brock-Talbot correlation is valid for Knudsen

number < 2. For Knudsen number > 2, it is common to use another correlation referred

to as modified Cha-McCoy-Wood (MCMW) [72] which gives reasonable results as

suggested by He and Ahmadi [73], as seen in Figure 2.8. When the particle diameter is

less than the mean free path, thermophoresis is dominated by the temperature gradient in

the flow. When the particle diameter is greater than the mean free path, a temperature

gradient is established within the particle, and both the gradients (bulk and in-particle) are

influenced by the conductivities of the gas and the particle.

Figure 2.8: Variation of thermophoretic coefficient with Knudsen number for

Brock-Talbot and MCMW correlations; Brock-Talbot correlation over predicts the

thermophoretic coefficient after Kn >2 [74]

(2.17)

28

2.8.2 Hydrocarbon Condensation

Condensation of species from the gas on the surface of the cooler occurs if the

temperature of the EGR cooler surface is less than the dew point temperature of the

species at the local partial pressure. Condensed hydrocarbons can create a sticky layer on

the surface promoting adhesion of particles and growth of the deposit layer [39].

Condensation creates a locally lower concentration region, setting up a concentration

gradient that drives diffusion of species from the gas, as shown in Equations (2.18) and

(2.19).

where,

ρg: Gas density ( kg/m3), Pr: Prandtl number

Sc: Schmidt number, cpg: Heat capacity (W/kgK)

yg: Mole fraction of vapor at i: interface, o: bulk mixture

2.8.3 Eddy Diffusion

The migration of particles from a high concentration to a low concentration region

is called diffusion. Submicron particles can diffuse due to eddies in the flow. The

deposition velocity due to diffusion in a turbulent flow is given by Equation (2.20) [75].

( )

( ) (2.18)

(

) (

) (2.19)

29

where,

Scp: Particle Schmidt number, Tm: averaged temperature (K)

u: friction velocity (m/s), Dp: particle diffusion coefficient (m2/s)

2.8.4 Turbulent Impaction

Particles from the exhaust stream can be deposited on the heat exchanger surface

through inertia. This occurs when the particle is large enough that it cannot easily follow

rapid changes in the gas flow direction. However, small particles have small relaxation

times, and thus follow the flow. The relaxation time should be compared to the smallest

time scale of the flow, known as the Kolmogorov scale KK and the largest time scale KL,

as shown in Equations (2.23) and (2.24). If the particle relaxation time is greater than KL,

the particle transport is inertially governed, else it is under the control of eddies.

√

(2.23)

(2.24)

(2.20)

(2.21)

(2.22)

30

2.8.5 Gravitational Forces

Gravitational drift velocity of the particles can be determined as shown in

Equation (2.25) [75]

where,

τ: particle relaxation time (s) , g: gravity (N/kg)

Abarham et al. [68] calculated the deposition flux (deposition velocity times

particle concentration) for the different mechanisms described above by assuming the

same mean concentration of particles in the exhaust gas. They found that at the

temperatures calculated, thermophoresis was at least two orders of magnitude larger than

the other mechanisms, as shown in Figure 2.9. The particle size distribution is overlaid on

the graph. Diffusion is important only when the particles are small (<50 nm) but still an

order of magnitude smaller than thermophoresis. Gravitational forces and turbulent

impaction are significant only when the size of the particle is larger than 1000 nm, which

is typically not observed in diesel exhaust. These results suggest that the dominant

mechanism for particle deposition in the EGR cooler is thermophoresis.

(

) (2.25)

(

) (2.26)

31

Figure 2.9: Comparison of deposition velocities for submicron particles [74]

Several investigators have attempted to model these processes to predict the

deposition velocity, soot layer thickness, and possible removal mechanisms [68,76–78].

These models, however, are usually 1-D/2-D and are based upon several assumptions,

and cannot exactly replicate experimental results. Nevertheless, these models serve as a

useful tool in isolating individual parameters affecting fouling of EGR coolers.

2.9 Deposit Removal Mechanisms

Although there is no agreement in the literature for self-cleaning mechanisms for

EGR coolers, several potential methods such as blow-out, flaking, mud-cracking,

oxidation, etc., have been suggested by Hoard et al. [39]. Different terminologies are used

by different researchers, but they all refer to the same form of deposit removal (for

32

example, mud-cracking versus fracturing). Blow-out occurs when the exhaust gas

removes the lose deposits from the surface due to shear force. Flaking occurs when the

deposits lose adhesion to the surface either due to a reduction in the adhesive forces

caused due to water vapor condensation, and/or hydrocarbon condensation. Oxidation

occurs when the temperature of the exhaust gas is high enough (~500°C) for the deposits

to burn off, however, most EGR coolers operate at temperatures much lower than the

oxidation temperature, and hence oxidation rarely takes place.

Epstein notes that the deposit removal process is not due to a lifting force as

commonly assumed, but a force trying to roll the particle downstream [69]. Charnay et al.

[79] hypothesize that a liquid layer, including water can loosen the deposit adhesion

during engine-off periods leading to a cooler recovery. However, the validity of these

mechanisms has to be confirmed experimentally. Abarham et al. [80] showed evidence of

water vapor condensate fracturing the deposit layer under cold coolant conditions.

Lance et al. [81] characterized deposits from 11 different on and off-road EGR

coolers supplied by various engine manufacturers. There was no consistent trend in

deposit mass distribution along the length of the cooler. Elemental analysis of the

deposits from these coolers showed that some deposits had large amounts of sulfur (15%

or more) and iron (10% or more) suggesting corrosion inside the EGR coolers. From the

microstructure analysis of the deposits, it was observed that the hydrocarbon

concentration was high close to the wall of the EGR cooler. Almost all coolers exhibited

some form of mud-cracking and “spallation” in the deposits suggesting probable recovery

mechanisms. Mud-cracking (or fracturing) refers to shrinkage of the deposit layer due to

33

loss of moisture. Spallation is a process by which lose deposits from the layer are

removed due to impact or stress caused due to difference in temperatures (thermal

cycling) across the deposit layer. Since the EGR coolers were exposed to different

conditions (on-road, off-road, different fuels, etc.), the exact reasons for mud-cracking or

spallation were not known, however, it appears that high temperature at the EGR inlet

can lead to cracks in the deposit layer.

2.9.1 EGR System Fouling Control

Zhan et al. [82] ran full scale EGR coolers with exhaust treatment systems

upstream. In the base case they used an uncoated catalyst and a flow-through-substrate.

The second system consisted of a diesel particulate filter (DPF) and an uncatalyzed wall-

flow filter with a diesel oxidation catalyst (DOC). The results from these experiments

demonstrated that the fouling is essentially due to the presence of hydrocarbons and soot

in the exhaust gas, since removing both allowed the cooler to maintain a nearly-clean

performance. One possible approach to minimize fouling is to draw the slipstream of

exhaust gas into the EGR cooler downstream of a DOC-DPF system, which is referred to

as the low-pressure EGR loop, as against the commonly used high-pressure EGR where

the slipstream is drawn upstream of the turbocharger before the DOC-DPF system.

However, low-pressure EGR results in an increase in fuel consumption, and hence

commonly avoided.

Lu et al. [83] investigated the effect of EGR filtration efficiency on EGR cooler

effectiveness using specially designed filters and found that the filters improved the

34

cooler effectiveness by about 20% over the baseline configuration. However, exhaust gas

treatment device upstream of the EGR cooler led to an increase in the pressure drop

across the EGR cooler. This means that the engine has to do more work to push the same

volume of exhaust gas through the EGR cooler, which results in a fuel economy penalty.

2.9.2 Diesel Oxidation Catalysts

Diesel oxidation catalysts (DOC) are aftertreatment devices for diesel engines,

used primarily to oxidize hydrocarbons and carbon monoxide in the exhaust stream. Most

DOCs come as a monolith honeycomb substrate coated with a metal catalyst and

packaged in small stainless steel containers. As the hot gases come in contact with the

catalyst, the HCs and CO gets oxidized to form CO2 and H2O. Typically, catalysts are

designed to oxidize the soluble organic fraction (SOF) of the particulate matter. Diesel

exhaust contains sufficient amounts of O2 for oxidation to take place, and O2

concentration ranges anywhere between 3-17% depending on the engine load. It is well

documented in the literature that the catalyst activity increases with temperature. Usually,

a minimum exhaust temperature of around 200°C is necessary for the catalyst to ‘light-

off’. At elevated temperatures, a catalyst can achieve up to 90% conversion efficiency

[84].

The role of EGR oxidation catalyst in reducing fouling in EGR coolers has not

been extensively studied. Only a few on-road diesel engines have EGR oxidation

catalysts upstream of the EGR cooler, and perhaps the Ford 6.4L Powerstroke diesel

engine introduced in 2006 was the first on-road vehicle with an ECAT [85]. The use of

35

oxidation catalysts is quite challenging in EGR applications for the following reasons.

Typically, at high engine loads, EGR% is reduced as it is known to reduce the peak

power output of the engine. However, high engine loads produce exhaust gas whose

temperature is high enough to produce 80-90% conversion efficiency of the catalyst [84].

On the contrary, at low loads, EGR% is higher; however, the gas temperature is low for

significant catalytic activity, as the catalyst would not have reached the light-off

temperature. Hence, ECAT’s are only effective in a certain range of engine loads.

Additionally, diesel engines produce exhaust gas with varying concentrations of HCs and

PM as a function of the engine condition, which limits the conversion efficiency of the

catalyst.

36

Chapter 3

Literature Review of High Efficiency Combustion

3.1 Introduction

In 2010, the DOE’s Vehicle Technologies Program initiated the SuperTruck

Program, whose goal was to design a heavy-duty Class 8 truck which demonstrated a

50% improvement in overall freight efficiency measured in ton-miles per gallon [86].

Along with the overall efficiency, each vehicle’s engine needed to show 50% brake

thermal efficiency moving towards 55%. In view of achieving such high efficiencies,

while still adhering to increasingly strict emissions regulations, alternate diesel

combustion strategies are being explored. The principle in achieving high efficiency

relies on optimization of the combustion systems, in terms of design, combustion

phasing, duration, etc. However, efficiency is governed not only by the combustion

system alone, but also by the nature of the fuel being burned. Advanced diesel

combustion is of great interest due to the promise of simultaneously reducing NOx and

PM emissions, while improving engine efficiency.

Some of the strategies commonly discussed today in the literature are the

homogenous charge compression ignition (HCCI), premixed charge compression ignition

(PCCI), partially premixed combustion (PPC), reactivity controlled compression ignition

(RCCI), etc. The following sections discuss in detail these combustion methodologies

with their merits and demerits.

37

3.2 Conventional Diesel Combustion

There are different stages of diesel combustion, which are well explained by the

heat release profile shown in Figure 3.1. The rate at which heat is released is described

chronologically by the crank angle at which the events occur during a compression

ignition combustion cycle.

Figure 3.1: Conventional diesel heat release profile [18]

Ignition Delay (a-b): This represents the time delay between start of injection and

actual start of combustion in the cylinder. During this process, the rate of heat release

drops below zero due to fuel absorbing heat during vaporization.

Phase of rapid combustion or the premixed phase (b-c): In this phase, the combustion

of fuel which has mixed with air occurs rapidly over a few crank angle degrees. This

phase is usually characterized by a high heat release rate, as seen by the sharp peak.

38

Phase of mixing controlled combustion (c-d): In this phase, the liquid fuel atomizes,

vaporizes, mixes with air, and finally burns with a diffusion flame. The rate of heat

release typically is not as high as the peak during the premixed phase, but it occurs

over a wider range of crank angle. Dec has shown that soot typically forms during the

diffusion burn phase [87].

Phase of late combustion (d-e): This can be termed as the last stage of heat release. It

is very low in its release rate, and might occur due to several reasons. It could be

because of some leftover fuel, or some energy stored in soot and fuel rich combustion

products. This happens over a few crank angle degrees.

Conventional diesel combustion process which has a stratified charge (zones of

air/fuel mixture) operates at high local temperatures which produce NOx emissions and at

equivalence ratios which produce PM emissions. The use of aftertreatment systems to

control both NOx and PM emissions has proven to be effective and necessary to comply

with the emissions regulations, however, it has been found that these aftertreatment

systems tend to have a negative impact on the engine efficiency and fuel consumption.

Hence there is a continuing need to identify potential methods of reducing these

emissions in-cylinder. The main idea in low temperature/advanced combustion is to have

the combustion process occur at temperatures below those at which NOx forms and at

equivalence ratios below those at which soot forms. Hence conventional diesel

combustion could be replaced with more premixed charge combustion. The next few

sections describe some of the advanced combustion strategies that employ premixed

charge combustion concept.

39

3.3 Homogeneous Charge Compression Ignition Combustion

The application of HCCI combustion is based on a combination of gasoline and

diesel engine operating characteristics. This is achieved by compressing the fuel air

mixture (or charge) to the point of autoignition. Essentially, the homogenous fuel air

mixture resembles a gasoline engines’ mixture preparation and ignition occurs not due to

a spark, but due to autoignition similar to the compression ignition process in a diesel

engine. Such a process results in the charge igniting at multiple locations simultaneously.

In contrast to conventional diesel combustion (diffusion controlled), HCCI reactions are

not limited by the mixing rate at the interface between the jet of fuel and surrounding

oxidizer. HCCI combustion has gained popularity as it can operate at diesel engine-like

compression ratios thus achieving greater efficiencies than gasoline engines [88].

Because the mixture is homogenous, the combustion process is cleaner and results in

lower NOx emissions due to lower combustion temperatures [89]. The main advantage of

such a process is that HCCI can be achieved with a wide variety of fuels [90]. Bessonette

et al. [91] suggested that the best fuels for HCCI operation may have autoignition

qualities between that of diesel fuel and gasoline.

Even though HCCI technology seems promising, the process is limited by a sharp

rise in peak cylinder pressure [92]. This can cause significant damage to the engine if the

engine is not designed to withstand these pressures. Additionally, since the combustion

process is not controlled by fuel injection or spark ignition, there is a difficulty in

controlling the combustion (and heat release) at the point of autoignition [93,94]. HCCI is

difficult to achieve with diesel fuel due to the challenge of vaporizing and mixing the fuel

40

thoroughly prior to combustion [12]. Even though HCCI combustion engines reduce NOx

emissions, the levels of HC and CO emissions increases [95]. This is because the early

injection causes over-leaning of the fuel-air mixture as well as the flame quenching at the

cold cylinder walls. The operation of the engine in HCCI mode is limited in its range of

operability over different speeds and loads as well as the cylinder pressure levels [96].

3.4 Premixed Charge Compression Ignition Combustion

PCCI combustion aims to achieve the same objectives of HCCI combustion i.e.

lean and homogeneous mixture to reduce NOx and soot emissions, with the difference

being in the way the mixture is prepared. In HCCI, the charge is homogeneous when

entering the cylinder after which it is compressed to the point of autoignition. In PCCI

combustion, the fuel in injected early into the cylinder and is combined with high EGR

rates to delay the start of combustion, thereby allowing all the fuel to be injected prior to

autoignition [97]. EGR offers a method of controlling the autoignition point for this type

of combustion. The overall reduced temperatures in the cylinder result in reduction of

NOx emissions.

Even though PCCI combustion has its advantages, an increase in HC and CO

emissions has been observed due to incomplete combustion of the fuel close to cylinder

wall [12,98]. These problems with HCCI and PCCI can be overcome by using different

types of fuel mixing and preparation, which has led to the development of partially

premixed combustion (PPC) and mixed mode or dual fuel combustion [99,100].

41

3.5 Partially Premixed Combustion

Partially premixed combustion lays in-between HCCI combustion and conventional

diesel combustion, in terms of fuel-air mixing. In PPC, a part of the fuel is injected early

during the compression stroke and then mixed with air to achieve premixed lean

combustion, and the remaining fuel is injected after TDC into the high-temperature

mixture. This eliminates locally rich regions, and the mixture is homogenous compared to

conventional diesel combustion. Moreover, combustion can be controlled by adjusting

the fuel injection timing. This results in improved engine efficiency (moving toward 50-

55% BTE) and reduced emissions [12,101].

One of the issues with using a high cetane number fuel such as diesel in PPC

combustion is that as the engine load increases, more fuel needs to be injected resulting in

longer injection durations. This results in some of the fuel being injected into the hot

products of combustion, which produces high levels of smoke [102]. This can be

minimized by extending the ignition delay period by using high levels of EGR, as well as

adjusting the injection timing [102]. Other ways to improve the combustion process

would be to use fuels with low cetane number and/or by optimizing the fuel reactivity

[103,104].

3.6 Mixed Mode Combustion

Mixed mode combustion or dual fuel combustion as applied to diesel engines

signifies the simultaneous combustion of gaseous and liquid fuel [105,106]. In mixed

42

mode combustion, a gaseous fuel is fumigated into the intake air and a conventional

diesel injection is used with the intention of igniting the premixed gaseous-fuel charge.

The phases of energy released are: combustion of the pilot fuel which was premixed with

air, combustion of the gaseous fuel in the vicinity of the pilot injection, and finally pre-

ignition reactions and turbulent flame propagation within the lean mixture [107], as a

result of combustion of the liquid fuel. Pawlak [106] showed that this combustion mode

might experience knocking or engine overheating.

3.7 Reactivity Controlled Compression Ignition Combustion

RCCI is a dual fuel engine combustion technology that was developed at the

University of Wisconsin-Madison Engine Research Center laboratories. RCCI is a variant

of HCCI that provides more control over the combustion process and has the potential to

dramatically lower fuel use and emissions [108]. RCCI uses in-cylinder fuel blending

with at least two fuels of different reactivity and multiple injections to control in-cylinder

fuel reactivity to optimize combustion phasing, duration and magnitude. The process

involves introduction of a low reactivity fuel into the cylinder to create a well-mixed

charge of low reactivity fuel, air and recirculated exhaust gases. The high reactivity fuel

is injected before ignition of the premixed fuel occurs, using single or multiple injections

directly into the combustion chamber. Addition of the second fuel allows significant

control over autoignition characteristics. Optimized stratification of fuel reactivity allows

control of combustion duration. Examples of fuel pairings for RCCI are gasoline and

diesel mixtures, ethanol and diesel, and gasoline and gasoline with small additions of a

43

cetane-number booster like di-tert-butyl peroxide (DTBP) [109]. RCCI combustion offers

practical low-cost pathway to more than 15% improvement in thermal efficiency

compared to conventional diesel combustion, with improved fuel economy (lower CO2

emissions). RCCI also offers great fuel flexibility and transient response [108].

RCCI allows optimization of HCCI and PCCI type combustion in diesel engines,

reducing emissions and the need for aftertreatment methods. By appropriately choosing

the reactivities of the fuel charges, their relative amounts, timing and combustion can be

tailored to achieve optimal power output (fuel efficiency), at controlled temperatures

(controlling NOx) with controlled equivalence ratios (controlling soot). Kokjohn et al.

have demonstrated an indicated thermal efficiency of almost 56% using the RCCI mode

of combustion [108]. Chapter 5 explores dual fuel combustion with fumigation of

liquefied gases into the engine air intake.

3.8 Fumigated Fuels for Dual Fuel Combustion

3.8.1 Dimethyl Ether (DME)

Dimethyl ether is the simplest ether, with a chemical formula CH3-O-CH3. DME

as a fuel in compression ignition engines has been considered since the l990s. Fleisch et

al. have shown that DME can be used in a diesel engine to obtain reductions in NOx

emissions meeting the California ULEV emissions requirement [110]. DME has a high

cetane number (higher than diesel) and it can be easily produced from a variety of

44

feedstock including biomass, coal and natural gas [111]. Table 3.1 compares some of the

properties of diesel and DME along with propane and methane.

Table 3.1: Physical properties of diesel, DME, propane and methane [112,113]

Property Diesel DME Propane Methane

Chemical Formula C10.8H18.7 C2H6O C3H8 CH4

Mole Weight (g/mol) 148.60 46.07 44.11 16.04

Boiling point (°C) 71-193 -24.9 -42.1 -162

Autoignition temperature (°C) 250 235 470 650

Stoichiometric Air/Fuel Ratio 14.6 9 15.6 16.9

Lower Heating Value (MJ/kg) 42.5 28.8 46.4 49.9

Cetane Number 40-55 55-60 - -

Octane Number - - 97 120

The cetane number describes the ignition quality of the fuel. The shorter the

ignition delay, the better the ignition quality of the fuel, and thus, the higher the cetane

number. As can be seen from Table 3.1, DME has a higher cetane number and a lower

autoignition temperature as compared to diesel. This means that DME when injected into

the cylinder can burn quicker than diesel with a shorter ignition delay. One of the reasons

attributed to the greater reactivity of DME in the combustion chamber is the lack of a

carbon-carbon bond [114]. Research on the oxidation of DME has demonstrated the

presence of OH, H and CH3 radicals during the propagation phase of the combustion

process [114]. The OH radical is then responsible for improving the ignition quality of

the fuel and shortening the ignition delay thus resulting in increased oxidation rates [115].

45

DME, on the contrary, has low lubricity (660 microns) which presents a challenge

for the fuel injection system [116]. This however can be overcome by using fatty acid

based lubricity improvers as demonstrated by Oguma et al. [117]. Literature has showed

that DME has a tendency to leak from the fuel injectors [118,119], which could be

potentially hazardous as DME can autoignite at much lower temperatures than diesel fuel

[120]. Even though DME is ideal to use during cold start conditions owing to its lower

boiling point, it has to be kept slightly pressurized as in the case of Liquefied Petroleum

Gas (LPG). The lower heating value of DME as compared to diesel also means that

greater amount of fuel has to be injected to provide the same brake power. The main

advantage of DME, however, is in its ability to reduce PM and NOx emissions [110], due

to a lack of carbon-carbon bond in the structure.

The numerous advantages of DME as listed above had led to many researchers

experimenting with DME and DME blends in both SI and CI engines

[92,110,112,115,120,121]. The fuel blend of DME and methane is one of the commonly

used mixture by researchers while experimenting with HCCI and mixed mode

combustion [112,115,122]. Methane has been popular for use along with DME as

increasing the methane content delays DME’s early ignition [112]. Work has been done

considering blends of DME with other fuels like propane and butane [123,124]. The

advantages and disadvantages of DME are summarized below [125].

Advantages

High oxygen content: The presence of fuel-bound oxygen and the absence of C–C

bond results in smokeless combustion.

46

Low boiling point: This leads to quick vaporization when a liquid-phase DME spray

is injected into the engine cylinder.

High cetane number: DME has better ignition quality compared to diesel fuel.

Disadvantages

Low calorific value: Since DME has lower energy content compared to diesel on a

mass and volume basis, a greater volume of DME has to be injected to obtain the

same power output as provided by diesel fuel.

Low viscosity: Lower than that of diesel fuel, causing leakage from the fuel supply

system which relies on small clearances for sealing. Its lower lubricity characteristics

can cause intensified surface wear of moving parts within the fuel-injection system.

3.8.2 Propane

Propane is produced as a by-product of two other processes, natural gas

processing and petroleum refining, and is one component of Liquefied Petroleum Gas

(LPG). The other components of LPG are propylene, butane, and butylene, and their

relative proportions vary according to the origin. Since propane has a high octane number

(97), it is a single stage fuel similar to gasoline. Propane is normally a gas, but is

compressed to a transportable liquid at a moderate pressure of 160 psi, and is stored in

pressure tanks at about 200 psi at 100°F [126]. When propane is drawn from a tank, it

changes to a gas before it is burned in an engine. In the United States, Autogas is the

common name for LPG when it is used as a fuel in internal combustion engines, and it

47

complies with the HD-5 specifications. HD-5 spec propane consists of a minimum of

90% propane, a maximum of 5% propylene, and other gases such as butane, butylene etc.

constitutes the reminder [126]. Autogas is a green, clean-burning alternative fuel and is

less expensive than gasoline or diesel [126].

Like methane, propane has also been used as a gaseous fuel for the premixed

charge in HCCI combustion [127,128]. Propane has also been used as a low cetane

number fuel along with DME to extend the advantages and complimenting the

deficiencies of using DME in CI engines [128]. Propane tends to combust all at once

compared to two heat release peaks observed for DME combustion [129], possibly due to

higher autoignition temperature of propane as compared to DME. For propane to be used

as a fuel in diesel engines, high compression ratios (>18) and inlet heating (~140°C) are

required [130]. In order to overcome this barrier, Yap et al. experimented with internal

trapping of the exhaust gases to raise in-cylinder temperatures [127]. They were able to

run the engine at a compression ratio (CR) of 15 without any intake air heating system

while observing reduced NOx emissions. Propane when used as a solitary gaseous fuel in

an HCCI combustion process faces the problem of poor combustion due to its high

autoignition temperature. But, propane when used in conjunction with another gaseous

fuel along with an injection of diesel can be used to control ignition and heat release.

3.8.3 Propane versus Methane

Both methane and propane have been used as fuels in internal combustion engines

as discussed in previous sections. Each of the fuels has its own pros and cons.

48

Propane has a higher boiling point than methane, so it's easier to liquefy, store and

transport.

Methane is lighter than air and thus tends to rise if released, while propane is heavier

and thus tends to sink to the floor and pool in enclosed spaces if it escapes.

Methane, when discharged into the environment is a greenhouse gas whereas propane

is not classified as such. Therefore, while propane will not contribute to pollution in

its unused state if released, methane will.

Propane, being more reactive than methane, gets oxidized before it reaches the

stratosphere and hence has a lesser impact on smog formation. Propane has zero

ozone depletion potential compared to methane [131].

Methane has a higher octane number (120) compared to propane (97). This implies

that propane has a higher cetane number than methane making it more suitable for

compression ignition than methane.

Propane when used alone can autoignite given high compression ratios and inlet air

heating, however, methane does not due to high autoignition temperature [112,130].

This gives an increased amount control over the ignition in terms of the range of

crank angles over which ignition can occur.

49

Chapter 4

Effect of Engine Operating Conditions, Coolant Temperature and Oxidation

Catalyst on Morphology and Composition of Deposits from a Fouled

Automotive Exhaust Gas Recirculation Cooler

4.1 Introduction

Diesel engines have higher efficiency and improved fuel economy compared to

gasoline engines. However, a key challenge with diesel engines lies with NOx and PM

emissions. NOx is a mixture of nitric oxide (NO) and nitrogen dioxide (NO2), of which

nitric oxide is by far the most abundant [18]. NOx is formed in regions where enough

energy is available for nitrogen to oxidize. Hence NOx formation is governed by higher

temperatures and the availability of oxygen [21]. PM is a complex mixture of organic and

inorganic compounds in the solid and liquid phases [132]. The aggregates formed of

primary spherical carbon particles are usually termed as soot, or the insoluble fraction

[133]. A layer of hydrocarbons is adsorbed but also condensed onto the particle

aggregates and are often referred to as the volatile organic fraction (VOF) or the soluble

organic fraction (SOF) [45,134]. Additionally, nitrates and water can be adsorbed on the

carbon-rich particles [135,136].

A common method of reducing NOx emissions in-cylinder, adopted by many

engine manufacturers, is by exhaust gas recirculation (EGR) [22,137–140]. The high CO2

content of the exhaust gas acts as a heat sink and reduces the adiabatic flame temperature,

which in turn reduces NOx emissions. Secondly, the circulation of exhaust gas into the

50

intake air dilutes the oxygen content of the intake air. This reduces the combustion

temperatures to further inhibit thermal NOx.

There are two types of EGR loops, the high-pressure and the low-pressure loop

[141]. In the high-pressure loop, EGR is drawn upstream of the turbocharger and is

mixed with fresh intake air downstream of the compressor. Hence the pressure in the loop

is at boost pressure. In the low-pressure loop, EGR is drawn downstream of an oxidation

catalyst – diesel particulate filter (DOC-DPF) assembly and is mixed with fresh intake air

upstream of the compressor. Hence the pressure in this loop is typically close to ambient.

Most engine manufacturers prefer to use the high pressure EGR as it is believed to offer

better fuel economy [39].

In most modern engines, heat exchangers (EGR coolers) are installed to cool the

exhaust gas before reintroduction into the engine intake. This reduces NOx emissions

further, however with a penalty. Soot and hydrocarbons from the exhaust gas migrate

toward the cold wall and begin to accumulate on the inside walls of the EGR cooler,

which results in a degraded performance of the cooler. These deposits form a layer which

is thermally less conductive than the stainless steel tubing resulting in lower thermal

effectiveness of the heat exchanger, often on the order of 20-30% and accordingly

clogging flow passages [79] as shown in Figure 4.1. As a consequence of this deposit

build-up, NOx emissions cannot be controlled to the desired levels.

51

Figure 4.1: Degradation of EGR cooler performance as a function of time [142]

Future emissions standards for all classes of diesel engines require increased EGR

flow rates and reduced intake charge temperatures, and hence particulate deposition

inside the EGR cooler has to be reduced or prevented so that engines can meet these

emissions regulations2. It is unlikely that fouling can be completely eliminated in the

presence of HCs and soot in the exhaust slipstream, unless the exhaust gas passes through

a series of catalysts and particulate filters as in the low-pressure EGR loop. Most current

on-road engines have a single EGR cooler primarily cooled by the engine coolant. So

typically, the temperature of the coolant is at 85-90°C. Some engines have twin-EGR

coolers (like the Ford 6.7L V8 turbodiesel [85]), driven by the need for additional

cooling. Engine manufacturers are discussing possibilities of using a standalone cooling

system to circulate the coolant at lower temperatures to help reduce NOx emissions even

further [58,143] .

2 EGR is just one component of the system of design and integration to achieve NOx targets.

52

Degradation of the EGR cooler performance in diesel engines has been studied

recently [35,36,53,64,67,144–147]. Although the fouling mechanism inside the EGR

coolers is not fully understood, most argue that thermophoresis and hydrocarbon

condensation are the dominant processes [78,142,148]. Several investigators have

attempted to model these processes to predict the deposition velocity, soot layer

thickness, and possible removal mechanisms [76–78,149]. However, the models are

usually either 1-D or 2-D, based on several assumptions, and do not replicate

experimental results.

Even with a rich literature on automotive heat exchanger fouling (presented in

Chapter 3), in particular diesel EGR cooler fouling, this phenomenon has not been clearly

understood, mostly due to the number of variables contributing to fouling and partly due

to engine calibration choices and operating conditions. There is very little published work

on the physical and chemical characteristics of the deposits from the EGR cooler,

especially knowing that engine operating parameters and boundary conditions in the EGR

cooler play a significant role in altering the nature of the deposit layer [39]. Most

published work in the literature examines fouling from an ‘engine and heat exchanger’

performance viewpoint and tends to ignore the analysis of EGR cooler deposits, for, the

latter is a complex process and requires access to specialty characterization equipment.

So there is a significant incentive to perform a ‘forensic analysis’ of the deposits to obtain

their properties and work backwards to understand the phenomenon of fouling. And

finally, understanding the properties of these deposits by performing an engine-based

parametric study will serve to develop fouling mitigation strategies in EGR coolers.

53

4.2 Objectives for Study of EGR Cooler Fouling

Perform a parametric study on the effect of engine operating conditions on the rate of

deposition, morphology and composition of deposits from a fouled EGR cooler.

Soot and hydrocarbon emissions from a diesel engine are dependent on the engine

condition. Low engine loads are known to produce more hydrocarbons due to low in-

cylinder temperatures while high load conditions are known to produce more soot

emissions and low hydrocarbon emissions [18,45–47]. However, the effect of varying

concentrations of these emissions on the properties of the fouled EGR cooler deposits

like microstructure, chemical composition, and elemental distribution has not been

studied yet. Therefore, it is of interest to examine the sensitivity of engine operating

conditions on EGR cooler fouling, in particular the rate of deposit accumulation and the

rate of effectiveness loss in the EGR cooler.

Identify probable routes to minimize fouling in EGR coolers.

To develop fouling mitigation strategies as a part of the engine duty cycle, it is

necessary to evaluate how different engine conditions affect the rate of fouling in EGR

coolers, due to inherent differences in the physical and chemical properties of the

deposits produced under these conditions. Based on the results obtained in the first

objective, different strategies such as use of EGR oxidation catalyst to oxidize

hydrocarbons, forced condensation of water vapor, etc., will be evaluated.

54

4.3 Experimental

4.3.1 Engine

An 8-cylinder Ford 6.4L “Powerstroke” direct injection turbodiesel engine

coupled to an eddy-current dynamometer was used as an exhaust generator. The engine is

equipped with two variable-geometry turbochargers, an EGR oxidation catalyst, and a

common rail fuel injection system. The engine has a brake power of 261 kW at 3000 rpm,

and a peak torque of 881 Nm at 2000 rpm. The original engine calibration complied with

EPA Tier 2 Bin 9 emissions standards. Engine operating parameters, especially the

engine control unit (ECU) parameters, were monitored via Inca v6.1 software. Additional

engine specifications are presented in Table 4.1.

Table 4.1: Engine specifications

Engine Type Ford Powerstroke 8 cylinder ‘V’ engine

Bore 9.82 cm

Stroke 10.5 cm

Compression ratio 17.2:1

Displacement volume 795 cc

Clearance volume 49 cc

Injection system Common rail direct injection

4.3.2 EGR Cooler Test Rig

Fouling studies on conventional EGR coolers are often difficult because of the

inability to control the feed-gas variables independently. Furthermore, most EGR coolers

55

come as welded assemblies, restricting the access to the deposits inside the tubes as

shown in Figure 4.2.

Figure 4.2: Conventional EGR cooler used in Ford “Powerstroke” engine

To overcome this situation, an in-house shell and tube heat exchanger with 6

surrogate tubes was designed such that the tubes could be removed easily from the

assembly as shown in Figure 4.3. The tubes were circular in cross-section, 0.25” in

diameter, 0.02” in wall thickness, and 11” in length, and were made of 304 stainless steel.

(The inlet manifold was re-machined and the manifold diameter was changed during

installation in the test rig to match the diameter of the exhaust pipe. The image shown is

from the design phase). A good seal between the gas and coolant was ensured by using

Swagelok ferrules at one end and removable graphite ferrules at the other. A slipstream

of the exhaust gas was drawn into the EGR cooler upstream of the turbocharger and

56

downstream of the EGR oxidation catalyst. This concept is similar to the high-pressure

EGR loop, the only difference being that exhaust gas was not recirculated into the engine

intake.

Figure 4.3: Model EGR cooler with 6 surrogate tubes

A high-temperature wedge flowmeter was installed to measure the volume and

mass flow rate of exhaust gas corrected to standard conditions. The calibrated flowmeter

has an uncertainty of 5% associated with it. A high-temperature valve was installed

downstream of all devices to control the flow rate of the slipstream exhaust gas. This

ensured that the pressure of the exhaust gas through the EGR cooler was at engine

exhaust pressure, which is much higher than ambient pressure. A recirculating chiller

(Neslab Thermoflex 1400) which could provide both heating and cooling was used to

control the temperature and flow rate of the coolant. This unit has a cooling and heating

capacity of 1.4kW and 1.0kW respectively, and comes with a 7.6L reservoir and a pump

capable of delivering 3.5gpm of coolant at 60 psig. The coolant temperature on this unit

57

can be varied between 20-90°C. Thermocouples and differential pressure transducers

were installed to measure the temperature and pressure drop across the EGR cooler. The

inlet manifold was designed to accommodate steam injection. The test rig was configured

such that all parameters were displayed in real-time via a custom-made LabView

program. An overview of the setup is shown in Figure 4.4 and Figure 4.5. Flow rate and

heat capacity calculations for the hot and cold side are provided in Appendix B. Since the

heat capacity of the cold (coolant) side was much larger than the hot side, the temperature

gain in the coolant was a minimum, given sufficient flow rate.

EGR Cooler

Flowmeter

HT Valve

Vent to atm

Diff. Pres. Gauge

T1 T2

Flowmeter

Valve

ReservoirPump

Neslab 1400 Closed loop Recirculating chiller

Exhaust inFord 6.4L engine

Figure 4.4: Test rig with in-house EGR cooler, flowmeter, high temperature valve

and recirculating chiller. Image is not to scale.

58

Figure 4.5: Engine test cell of EGR cooler test rig, showing the EGR cooler, high

temperature flowmeter and valve, pressure and thermocouple instrumentation.

Recirculating chiller is not shown in this image.

4.3.3 Modifications for Catalyst Study

The Ford Powerstroke engine is equipped with an EGR oxidation catalyst

(ECAT/EOC) as shown in Figure 4.6. The ECAT uses a honeycomb substrate with a

catalyst coating (Platinum/Palladium) with a metal loading of 50 mg/ft3, and is only

slightly larger in diameter than the pipes it is connected to. To facilitate a study without

the catalyst, the manifold consisting of the ECAT was disassembled and the honeycomb

structure containing the catalyst was removed from the assembly. The exhaust manifold

was then reassembled onto the engine.

EGR cooler

Flowmeter

Valve

59

Figure 4.6: Ford 6.4L engine drawing showing the EGR oxidation catalyst and dual

EGR coolers [85]

4.3.4 Fuel

The fuel for these experiments was an ultra-low sulfur diesel obtained

ChevronPhillips Chemical, details of which are presented in Table 4.2. Some relevant

fuel properties are provided in Appendix A.

Table 4.2: ChevronPhillips Chemical ULS 2007 diesel fuel properties

Properties Value

Cetane number 45

Oxygen content (%) 0

Specific gravity 0.8466

Heating value (MJ/kg) 42.8

Sulfur content (ppm) 9.7

T90 temperature (°C) 308

60

4.3.5 Emissions Measurement

An AVL Combustion Emissions Bench II was used to measure gaseous

emissions. Hot exhaust gases were sampled from the engine’s exhaust pipe by headline

filters through heated lines kept at 195ºC. NOx emissions were measured using an

EcoPhysics chemiluminescence analyzer. Total hydrocarbon emissions were measured

using ABB Flame Ionization detectors.

Particulate matter was sampled using a Sierra Instruments BG 3 particulate partial

flow sampling system. The total flow through the system was maintained at 75 lpm while

the dilution flow was maintained at 67.5 lpm, resulting in a dilution ratio of 10. The

particulate matter samples were collected on 47 mm Pallflex filters, which were placed in

an environmental chamber set at 25ºC and 45% relative humidity 48 hours prior to the

experiment to minimize errors due to different levels of moisture. These filters were

utilized for PM gravimetric measurements only, and no chemical extraction or thermal

analyses were performed. The average and standard deviation of three filters per engine

condition was used to represent the error bars.

4.3.6 Analytical Techniques

Deposit Microstructure: SEM

The microstructure of the deposits from the EGR cooler was analyzed using a

Hitachi S-3500N SEM. The deposits were placed on a vacuum safe double-sided carbon

adhesive tape and mounted on the sample holder. Since the conductivity of the deposits

61

was low, gold was sputtered on the sample to get good contrast on the images by

improving the thermal conductivity by charging. Both high (x 20K) and low

magnification (x 50) images were obtained from different locations on the sample. The

images were collected in the secondary electron mode (SE) to reveal the morphology and

topography of the sample. This is in contrast to the back scatter images (BSE), which are

useful for understanding differences in composition across the sample.

Chemical Signature: Pyrolysis GC-MS

Hydrocarbon-generated soot contains polyaromatic hydrocarbons (PAH) which

can be resolved by GC-MS. However, there is a considerable amount of material which is

derived from PAH compounds that cannot be detected by gas chromatography and are in

the 300-3000 Da range. Solvent extraction of soot followed by analysis by both GC and

HPLC can identify only the soluble fraction of the PAH’s. Pyrolysis is a method of

chemical analysis in which the sample is heated to decomposition in an inert atmosphere

or vacuum to produce smaller molecules that are separated by gas chromatography and

detected by a mass spectrometer. This technique for deposit analysis has proven to be

useful as it eliminates the process of solvent extraction and can be achieved with a

sample mass of a few micrograms. For chemical analysis of the deposits, a pyrolysis GC

(HP 5971 GC/MS and CDS Pyroprobe 1000) coupled to a mass spectrometer was used.

The pyroprobe interface temperature was set at 300°C. Pyrolysis was performed at 600°C

with a ramp rate of 5°C and a dwell time of 10s. The oven temperature was increased

from 40°C (1 minute hold) to 300°C at 4°C/min, and maintained at this final temperature

for 10 minutes. Interpretation of the mass spectra was carried out using an integrated

62

library. These conditions were selected based on a survey of pyrolysis GC experiments

performed on diesel soot [150,151].

CHN Elemental Analysis

Elemental analysis of the deposit was performed on a LECO CHN 600 analyzer

which measures total carbon, hydrogen, and nitrogen. Sulfur percentage was assumed to

be negligible and oxygen percentage was calculated from the difference. No prior

treatment was performed on the deposits and the analysis was performed on the as-

received sample.

Thermogravimetric Analysis

A thermogravimetric analyzer (TA instruments, SDT Q600) was used to

determine the volatile organic fraction in the deposits. This method enables analysis of

the continuous weight loss of the sample under inert and/or oxidant atmosphere. This

technique has proven to be very useful for analysis of diesel soot samples as it helps

avoiding time-consuming and potentially dangerous solvent extraction processes like

Soxhlet extraction in dichloromethane [152,153]. TGA repeatable experiments require

samples on the order of a few milligrams, while solvent extraction processes would

require samples on the order of a few hundreds of milligrams.

Siegl and Zinbo note that the weight loss up to ~200°C is due to a combination of

moisture and low boiling point HCs, weight loss up to ~ 350°C is from high boiling point

HCs which are sometimes engine oil related, and weight loss up to ~500°C is usually

63

from oxidized HCs and polymeric additives in the fuel [154]. Table 4.3 describes the

experimental method adopted for calculation of VOF as described by Yehliu [155].

Table 4.3: TGA procedure to determine VOF [155]

Step Procedure

0 Start with nitrogen

1 Ramp at 10°C/min to 30°C

2 Isothermal for 30 minutes to stabilize the sample

3 Ramp at 10°C/min to 500°C

4 Isothermal for 60 minutes to remove the VOF

5 Natural cooling of the sample

4.3.7 Engine Test Conditions

To achieve the objectives of this research plan, the study is divided into three

tasks, as outlined below.

Task 1: Evaluate the effect of engine operating conditions and coolant temperature

on EGR cooler fouling.

The objective of this test was to understand how engine conditions and coolant

temperature affected fouling in EGR coolers. For this study, two engine conditions,

representing low load and medium load were selected to produce different concentrations

of PM and HCs. The low load condition was close to an idle condition (81 Nm), and the

medium load (203 Nm) condition represented a cruising condition for the current class of

engine. Two coolant temperatures of 85°C and 40°C were selected for these experiments.

The model EGR cooler was exposed to exhaust gas for 9 hours. The experiments were

64

performed with the EGR oxidation catalyst in the exhaust manifold prior to the EGR

cooler. Detailed engine operating conditions are given in Table 4.4.

Table 4.4: Cruise and near-idle operating conditions

Engine Conditions Experiments Coolant Temperature

Experiments

Speed, Load 2150 rpm, 203 Nm 1400 rpm, 81 Nm 2150 rpm, 203 Nm

Mode Cruise Near-idle Cruise

EGR inlet

temp.

260-270°C 170-180°C 260-270°C

BSPM, g/kWh 2.27 1.64 2.27

BSHC, g/kWh 1.99 3.26 1.99

Coolant temp. 85°C 85°C 40°C

Vol. flow rate 180 ± 10 lpm corrected to standard conditions

Task 2: Examine the role of water vapor condensation inside EGR coolers for

monitoring thermal effectiveness recovery.

From the literature, it has been found that water can weaken the adhesive forces in

the deposits leading to a recovery of the thermal effectiveness in an EGR cooler. Water

vapor condensation inside the EGR cooler can be achieved without major modifications

to the EGR subsystems. In most engine/ECU calibrations, EGR is turned off during

engine start-up to prevent rough idling or cold start problems. During cold start, the

temperature of the engine coolant will be low enough to promote water vapor from the

exhaust gas to condense. The objective of this task was to assess the benefit of water

vapor condensation on the thermal effectiveness improvement and identify the critical

coolant temperature required for water vapor to condense inside the EGR cooler.

65

In three separate experiments, coolant was circulated at 25°C, 40°C, and 85°C on

tubes previously fouled for 4 hours such that the deposit bearing surface was exposed to

different temperatures, while maintaining the same exhaust gas flow rate through the

EGR cooler. Additionally, the effect of engine start and shut down periods on the EGR

cooler recovery was investigated at different coolant temperatures. Additional

experimental details are provided in Table 4.8.

Task 3: Examine the role of oxidation catalyst in minimizing EGR cooler fouling

Catalysts require a minimum temperature called the “light-off temperature” for

them to be at least 50% effective, and this temperature is usually around 200-220°C [84].

Certain engine conditions produce exhaust gas whose temperature is less than the light-

off temperature, rendering the catalyst ineffective. The objective of this task was to

evaluate the potential benefits of using oxidation catalysts to minimize hydrocarbon

concentration in the EGR cooler deposits, as a step toward minimizing fouling in EGR

coolers. Fouling studies were performed with and without an oxidation catalyst in the

exhaust manifold prior to the EGR cooler. The effectiveness of the EGR oxidation

catalyst was evaluated at the two engine conditions outlined in Table 4.4. Deposits from

the EGR tubes were then analyzed for their microstructure and chemical composition.

66

4.4 Results and Discussion

4.4.1 Effect of Engine Cruise and Near-Idle Conditions on EGR Cooler Fouling

The two engine conditions outlined in Table 4.4 resulted in EGR inlet

temperatures of 270°C and 170°C. The temperature profiles for each of these conditions

are plotted in Figure 4.7. The engine exhaust temperature was much higher than the EGR

inlet temperature upstream of the turbocharger. Additionally, the residence time of the

exhaust gas inside the EGR cooler is short, compared to the travel time through the

exhaust manifold before being vented into the atmosphere. These are important because

temperature and residence time play a significant role in changing the properties of the

deposits (longer residence time of the exhaust gas in the EGR loop promotes greater gas

side temperature drop, due to which, more HCs condense). From Figure 4.7, it can be

observed that the medium load condition which produces a higher EGR inlet temperature

experiences a greater temperature drop in the heat exchanger than the low load condition.

This observation is as expected due to a higher rate of heat transfer from the hot exhaust

to the cold coolant for the same mass flow rate and test duration.

67

100

150

200

250

300

350

400

0 100 200 300 400 500 600

Tem

pe

ratu

re,

°C

Time, min

Figure 4.7: Effect of engine operating condition on temperature profiles, 2150

rpm, 203 Nm, exhaust, 2150 rpm, 203 Nm, EGR inlet, 2150 rpm, 203 Nm,

EGR outlet, ●1400 rpm, 81 Nm, exhaust, 1400 rpm, 81 Nm, EGR inlet, ▲ 1400

rpm, 81 Nm, EGR outlet

Exhaust gas and EGR inlet temperatures remained fairly constant throughout the

test, however, EGR outlet temperature increased significantly for both conditions from

the start, indicating fouling of the tubes. The EGR outlet temperature increased from

149°C to about 183°C for the high load condition, and from 121°C to 139°C for the low

load condition over 9 hours. The drop in EGR outlet temperature every 1.5 hours was

mainly due to the replacement of a tube and partially due to recovery of the cooler during

the engine shut down period. Pressure drop (Δp) across the EGR cooler increased

marginally (~0.5 kPa increase) after the 9 hour test for both engine operating conditions.

The effectiveness change for the two conditions is shown in Figure 4.8.

68

-20

-15

-10

-5

0

5

0 100 200 300 400 500 600

Eff

ecti

ven

es

s C

ha

ng

e,

%

Time, min

Low Load

High Load


effectiveness change, 2150 rpm, 203 Nm, 1400 rpm, 81 Nm

It can be observed that the effectiveness for the medium load condition drops

more rapidly than for the low load condition, indicating a higher degree of fouling. This

can be explained on the basis of thermophoresis, which is a phenomenon driven by the

temperature gradient in the cooler. A high load condition results in a high temperature

exhaust gas, and increases the thermal gradient inside the EGR cooler. This increases

thermophoresis, and more particles from the exhaust gas migrate toward the cold wall.

This trend was also reflected in the amount of deposits collected for these conditions

shown in Figure 4.9, which confirms that the cruise condition has greater mass of

deposits collected inside the EGR cooler. (The deposits weight was measured by

weighing the tube before and after the experiment).

69

10

20

30

40

50

60

70

80

0 100 200 300 400 500 600

Av

era

ge D

ep

os

it M

as

s, m

g

Time, min

High Load

Low Load

Figure 4.9: Time varying effect of engine operating conditions on deposit mass,

2150 rpm, 203 Nm, 1400 rpm, 81 Nm

It is important to distinguish here that the constituents (PM/HC) of the exhaust

gas for the two conditions are different. High engine loads produce more soot mass

emissions and lower hydrocarbon emissions and vice versa (refer to Table 4.4). Diffusion

is important for transport of gaseous species, but not that important for soot particles in

the size of a few nanometers [20]. Hence thermophoresis becomes the dominant

mechanism of migration for the soot particles. EGR cooler effectiveness which drops

rapidly initially plateaus towards the end of the 9 hour test. The incremental mass gain

over time also reduced as shown in Figure 4.9. Similar observations in effectiveness have

been reported by several researchers [28-30].

To understand the microstructural change of the deposit as a function of time, five

tubes (each tube representing 1.5-7.5 hours exposure time) from the medium load

70

condition were milled down using a milling machine to 1-2 mils and cut open in the

center to produce tube and deposit samples which could be directly viewed under a

microscope. The pieces were taken from the center of the tube which minimized any

errors associated with fluctuations in flow or variations in the coolant temperature. The

deposits adhered firmly to the surface and milling did not dislodge any deposit, except at

the place where the final cut was made. Low magnification images of the deposit on the

tube surface are shown in Figure 4.10.

a)

b)

c)

d)

e)

Figure 4.10: Variation of deposit microstructure (low magnification) as a function of

time, a) 1.5 hours, b) 3.0 hours, c) 4.5 hours, d) 6.0 hours, and e) 7.5 hours

71

At 1.5 hours (Figure 4.10a), the deposits from the exhaust gas randomly

accumulated on the surface of the tube and the deposit layer appeared to align in the

direction of the exhaust gas flow. As time progressed, more deposits accumulated on the

surface of the tubes due to thermophoresis or due to deposits getting caught on the

existing layer (mechanical binding) to cover up the entire wall surface. By 7.5 hours, it

can be observed that the entire wall surface was covered and the deposit layer appeared to

be smooth and dense.

High magnification images (20,000X) of these samples are shown in Figure 4.11.

It can be observed from these images that the deposits had large pores in the beginning

and these pores got filled with more deposits (either due to accumulation of more soot

particles or due to hydrocarbons condensing) leading to a much denser deposit at 7.5

hours. Once the entire tube surface was covered with deposits, the deposit layer acted as

an insulator and the exhaust gas could no longer be cooled. It has been reported in the

literature that the porosity of the deposit layer is around 98% [66], which makes the layer

an insulator. This explains why the effectiveness change profile shown in Figure 4.8

reached an asymptote, and the EGR outlet temperature did not change much. In real EGR

coolers, it would take more time for the effectiveness to stabilize as the number of tubes

is much more than just six.

72

a)

b)

c)

d)

e)

f)

Figure 4.11: Variation of deposit microstructure (high magnification) as a function

of time, a) 1.5 hours, b) 3.0 hours, c) 4.5 hours, d) 6.0 hours, and e) 7.5 hours, and f)

9.0 hours

The deposits at different time intervals were analyzed using a Py-GC/MS

technique and the chromatographs are shown in Figure 4.12. From a direct visual

interpretation of the chromatographs, no significant change is observed in terms of the

peaks eluted at different time intervals. One possible explanation for this is that the

temperature inside the EGR cooler is too low for chemical reactions to take place. The

GC-MS data were analyzed and the relative percentages of aliphatics (C10-C17 alkanes,

C18-C25 alkanes) and aromatic hydrocarbons was obtained and plotted in Figure 4.13.

73

Some alcohols and oxygen containing species eluted, but their relative percentages were

low and were not plotted.

Figure 4.12: Py-GC chromatographs of species eluted as a function of time, a) 1.5

hours, b) 3.0 hours, c) 4.5 hours, d) 6.0 hours, e) 7.5 hours, and f) 9.0 hours

From Figure 4.13, it appears that the aromatic content marginally increased with

time, with the exception of the ‘3.0 hours’ case. This means that the net aliphatic content

(mainly from hydrocarbons condensing) reduces as time progresses. This can be

explained based on the temperature the exhaust gas is exposed to, as a function of time.

Initially, the surface of the cooler is clean and is at the temperature of the circulating

coolant, i.e. 85°C. But as the deposit layer builds up, the surface which now comprises of

the deposit layer is no longer at 85°C, but at a temperature higher than 85°C, and closer

to the exhaust gas temperature, which is much higher than the dew point temperature of

10 20 30 40 50

a.u

Retention time, min

(a)

(b)

(c)

(d)

(e)

(f)

74

most hydrocarbon species in the exhaust gas. This reduces hydrocarbon condensation.

Similar observations were reported by Sluder et al. [146], suspecting that hydrocarbons

are mostly located near the cold wall. The elemental composition for each of these

deposits is shown in Table 4.5. With the exception of 7.5 hours, hydrogen content

reduces with time indicating a decrease in aliphatic hydrocarbon percentage (aliphatics

have higher H/C ratio compared to aromatics for the same number of carbon atoms).

However, there exists a degree of uncertainty in these measurements as hydrogen can

attached to both aromatic rings and aliphatic chains. There is no obvious trend in nitrogen

and oxygen percentages, and these are rendered inert for comparisons in this study. These

variations would become clearer with longer duration tests. Thus, the EGR cooler

deposits undergo significant physical and minor chemical changes as time progresses.

0

20

40

60

80

100

Aromatics Aliphatics C10-C17 C18-C25

Perc

en

tag

e

Figure 4.13: Variation of aromatics and aliphatics percentage as a function of time

for 2150 rpm, 203 Nm engine condition, , 1.5 hours, 3.0 hours, 4.5 hours,

6.0 hours, 9.0 hours, and data unavailable for 7.5 hours condition

75

Table 4.5: Effect of time on elemental composition of deposits

Time, hours Carbon, % Hydrogen, % Nitrogen, % Oxygen, %

1.5 88.00 ± 0.20 1.15 ± 0.09 0.31 ± 0.15 10.54 ± 0.26

3.0 87.95 ± 0.05 0.99 ± 0.09 0.10 ± 0.10 10.95 ± 0.14

4.5 86.20 ± 0.20 0.94 ± 0.04 0.26 ± 0.05 12.59 ± 0.11

6.0 89.25 ± 1.15 0.90 ± 0.03 0.53 ± 0.21 9.32 ± 0.97

7.5 87.05 ± 0.75 1.21 ± 0.20 0.38 ± 0.06 11.36 ± 1.00

9.0 87.40 ± 0.80 0.83 ± 0.05 0.38 ± 0.07 11.39 ± 0.78

The previous section described the time varying effect deposit properties for one

operating condition, but it is also important to understand how different engine conditions

affect the deposit layer properties. Figure 4.14 shows the deposit microstructure at 203

Nm and 81 Nm. It can be observed that at 203 Nm condition (Figure 4.14a), the deposits

are coarse, formed mostly due to soot particles. These create smaller sized and more

numerous pores. Quantitative porosity measurements were not possible since such

measurements require deposit quantities of a few hundred milligrams, which the current

conditions did not generate in the time frame discussed. At 81 Nm condition (Figure

4.14b), the deposits appear larger, mainly due to greater hydrocarbon condensation,

leading to larger sized but fewer pores. The observations are consistent with the

observations from Marks and Boehman [156]. These results explain why the

effectiveness change shown in Figure 4.8 is different for the two conditions.

76

a)

b)

Figure 4.14: Effect of engine operating condition on EGR cooler deposit

microstructure, a) 2150 rpm, 203 Nm, b) 1400 rpm, 81 Nm

From the Py-GC chromatograms shown in Figure 4.15 and subsequent data

analysis in Figure 4.16, it appears that the net aromatic content did not change

significantly. However, it is interesting to observe that the low load condition has a

greater percentage of heavier alkanes compared to the medium load condition. It should

be mentioned here that soot particles are much denser and weigh more than the

condensed hydrocarbons. The variation in composition and density of PM and HC affects

the overall mass of the deposits in the EGR cooler.

77

10 20 30 40 50

a.u

Retention time, min

b)

a)

Figure 4.15: Py-GC chromatographs of species eluted as a function of engine

operating condition, a) 2150 rpm, 203 Nm, b) 1400 rpm, 81 Nm

0

20

40

60

80

100


Perc

en

tag

e


engine operating condition, 2150 rpm, 203 Nm, 1400 rpm, 81 Nm

78

From the elemental composition shown in Table 4.6, it can be observed that the

low load condition had a much higher content of hydrogen compared to the medium load

condition, with similar carbon content, indicating that a greater percentage of

hydrocarbons participated in deposition via condensation. The volatile organic fraction

(VOF) at the end of 9 hours from each load condition was obtained using a TGA and is

plotted in Figure 4.17. It can be observed that the deposit from low engine load condition

had a much higher percentage of volatiles compared to the deposits from high engine

load condition. This confirms the findings from Table 4.6, that the deposits from low

engine load condition were comprised of a greater percentage of hydrocarbons compared

to the deposits from high engine load condition.

Table 4.6: Effect of engine condition on elemental composition of deposits

Mode Carbon, % Hydrogen, % Nitrogen, % Oxygen, %

2150 rpm, 203 Nm 87.40 ± 0.80 0.83 ± 0.05 0.38 ± 0.07 11.39 ± 0.78

1400 rpm, 81 Nm 86.31 ± 0.88 2.30 ± 0.19 0.61 ± 0.04 10.78 ± 0.69

79

70

75

80

85

90

95

100

105

0 20 40 60 80 100

(VO

F m

ea

su

rem

en

t) W

eig

ht,

%

Time, min

Figure 4.17: Effect of engine operating condition on volatile organic fraction of

deposits after 9 hours test, 2150 rpm, 203 Nm, 1400 rpm, 81 Nm

These results conclude that engine conditions exert significant influence on the

EGR cooler deposit properties. Fouling becomes a severe issue especially when the

engine condition goes through an operating cycle of heavy load followed by idle, shut

down and cold start-up. It is necessary to prevent the EGR valve from getting fouled, as a

stuck EGR valve can cause rough idle and stalling of the engine [157]. This might lead to

elevated NOx emissions and defeats the purpose of using EGR.

4.4.2 Effect of Coolant Temperature on EGR Cooler Fouling

Coolant temperature plays a significant role in the fouling of heat exchangers.

EGR coolers on diesel engines are cooled by the engine coolant, whose temperature

80

varies between 85-90°C for fully warmed engines. Many engine and EGR cooler

manufacturers are discussing the possibility of having a standalone cooling system

circulating the coolant to the EGR cooler at much lower temperatures to reduce the

temperature of the charge going into the engine so that NOx emissions can be further

reduced [58,143]. This however is limited by packaging space in the design and the

additional energy necessary to operate the stand-alone system which can negatively

impact the fuel economy of the engine.

For this study, the coolant was circulated at 85°C and 40°C. The engine was

operated at 2150 rpm, 203 Nm and the volume flow rate of the exhaust gas was

maintained at 180 ± 10 lpm. The coolant flow rate was maintained sufficiently high to

minimize coolant side temperature gain. Additional details are presented in Table 4.4.

The effect of coolant temperature on the EGR cooler temperature profiles is

shown in Figure 4.18. The exhaust gas temperature and the EGR inlet temperature were

similar between the two runs. It is understood from the first principles of heat transfer

that the lower coolant temperature exchanges more heat and cools the exhaust gas further,

as observed in Figure 4.18. The EGR outlet temperature increased from 149°C to about

183°C (85°C coolant) and from 109°C to 142°C (40°C coolant).

81

100

150

200

250

300

350

400

0 100 200 300 400 500 600

Tem

pe

ratu

re,

°C

Time, min

Figure 4.18: Effect of engine operating condition on temperature profiles, 85°C,

exhaust, 85°C, EGR inlet, 85°C, EGR outlet, ● 40°C Nm, exhaust, 40°C,

EGR inlet, ▲ 40°C, EGR outlet

Thermal effectiveness change for the two operating conditions is plotted in Figure

4.19. It can be observed that the EGR cooler is more effective when the temperature of

the coolant is lower. Similar to the observations reported in the earlier sections, the

effectiveness stabilizes after a certain number of hours of operation. It is observed that

the recovery of the EGR cooler is higher when the coolant temperature is lower, even

though most of the effectiveness improvement is due to the removal of a tube every 1.5

hours. It is suspected that the cold coolant condition promotes greater water vapor

condensation inside the tubes assisting in a deposit removal process, similar to the

process described in the literature [80,158]. Additional details and results with cold

coolant effectiveness recovery are presented in section 4.4.3.

82

-20

-15

-10

-5

0

5

0 100 200 300 400 500 600

Eff

ecti

ven

ess

Ch

an

ge,

%

Time, min

Recovery associated with

shutdown and tube removal

Figure 4.19: Time varying effect of coolant temperature on EGR cooler effectiveness

change, 85°C, 40°C

The average deposit mass gain is plotted for each of these conditions in Figure

4.20. It can be seen that the 40°C coolant condition has more deposit mass than the 85°C

coolant condition. This result is consistent with the findings of Sluder et al. [12] and can

be due to two reasons. A cold coolant leads to an increase in thermophoresis due to a

higher temperature gradient in the flow. Secondly, a cold coolant increases the percentage

of hydrocarbons condensing from the exhaust stream, as the temperature is lower than the

dew point temperature of many hydrocarbons. It is known that the hydrocarbon dew point

temperature (analogous to water-vapor dew point temperature) increases with increase in

the chain length of aliphatics [159,160]. This means that heavier HCs have a tendency to

condense first as their dew point temperatures are higher compared to the lighter HCs.

83

20

30

40

50

60

70

80

90

100

0 100 200 300 400 500 600

Av

era

ge D

ep

os

it M

as

s,m

g

Time, min

Figure 4.20: Time varying effect of coolant temperature on the mass of deposits,

85°C, 40°C

The deposit microstructures for the two conditions are shown in Figure 4.21. It

can be observed that even though the engine operating condition is identical, the deposit

microstructures are different under different coolant conditions. The 85°C condition

deposits are coarse, comprised mostly of soot particles, however, the 40°C condition

deposits are larger, being comprised of more hydrocarbons, indicating that a greater

percentage of hydrocarbons condensed on the surface. However, the interplay between

thermophoresis and condensation is not clearly understood.

84

a)

b)

Figure 4.21: Effect of coolant temperature on deposit microstructure, a) 85°C

coolant, b) 40°C coolant

The Py-GC chromatographs are plotted in Figure 4.22. The aromatic

hydrocarbons eluted in the first 20 minutes, while the aliphatics eluted from 20 minutes

to 50 minutes. It is evident that the deposits from the cold coolant condition have more

aliphatics (heavier hydrocarbons) compared to the 85°C condition. Post processing

analysis of the chromatograph in Figure 4.23 shows that 40°C coolant condition has

about 20% more (relative proportions) heavier alkanes (C18-C25) than the 85°C coolant

condition. Even though the net aliphatic percentage remains the same between the two

conditions, it is important to note that on a mass basis, C18-C25 alkanes would weigh

more due to their higher molecular weight. The heavy hydrocarbons which condense

early on the surface of the EGR cooler will prevent the light hydrocarbons from

condensing due to an increase in the surface temperature of the deposit layer.

85

0 10 20 30 40 50

a.u

Retention Time, min

a)

b)

Higher aliphatics

Figure 4.22: Py-GC chromatographs of species eluted as a function of coolant

temperature, a) 85°C coolant, b) 40°C coolant

0

20

40

60

80

100


Perc

en

tag

e


coolant temperature, 85°C coolant, 40°C coolant

86

Elemental analyses of these deposits were performed and are shown in Table 4.7.

It can be observed that 40°C coolant condition had greater hydrogen content than for the

85°C coolant condition, which is an indication of higher percentage of aliphatic

hydrocarbons. Even carbon content increased for the cold coolant condition, which could

be a result of increased thermophoresis. Since these mechanisms are so complex, exact

contribution of each of them cannot be determined using the methods of this study.

Nonetheless these results make clear that coolant temperature plays a significant role in

altering the physical and chemical properties of the deposits from the EGR cooler.

Table 4.7: Effect of coolant temperature on elemental composition of deposits

Coolant Temp Carbon, % Hydrogen, % Nitrogen, % Oxygen, %

85°C 87.40 ± 0.80 0.83 ± 0.05 0.38 ± 0.07 11.39 ± 0.78

40°C 89.66 ± 0.54 1.37 ± 0.06 0.52 ± 0.05 8.45 ± 0.51

The VOF content for the deposits at the two coolant conditions is plotted in

Figure 4.24. It can be clearly seen that the deposits at 40°C coolant condition had a much

higher VOF compared to the deposits at 85°C, which is a direct consequence of

increasing hydrocarbon condensation, confirming the trends observed in the

chromatographs and the elemental composition profiles.

87

80

85

90

95

100

105

0 20 40 60 80 100

(VO

F m

ea

su

rem

en

ts),

We

igh

t, %

Time, min

Figure 4.24: Effect of coolant temperature on volatile organic fraction from

deposits, 85°C, 40°C

4.4.3 Role of Water Vapor Condensation on EGR Cooler Recovery

Results from the previous section suggested that water vapor condensation in the

EGR cooler tubes at cold coolant conditions provided a solution for EGR cooler

recovery. Water vapor condensation inside the cooler can be examined theoretically by

calculating the critical temperature required for the process. With an air-fuel ratio of 29.1

from experimental results, and by assuming an empirical formula for diesel fuel to be

CH1.8, the chemical reaction for diesel fuel combustion can be represented as shown in

Equation (4.1).

( )

(4.1)

88

The mole fraction of water in the products is 0.06. The pressure of the gas in the

EGR cooler is measured to be 172506.8 Pa which results in a partial vapor pressure of

water of 10819.8 Pa. The Antoine equation shown in Equation (4.2) is used to determine

the critical temperature for water condensation [161].

Antoine coefficients for water are: AA = 10.23, BB = 1750, CC = 235. Using

these values, we get an interface temperature of 47.4°C. This temperature represents the

critical temperature below which water vapor from the exhaust gas will condense in the

system. These calculations have been made for diesel combustion without EGR.

Introduction of burned gas into the cylinder would result in higher interface temperatures.

Under the cold coolant condition, the temperature of the surface was maintained at 40°C,

which was significantly lower than the critical interface temperature, which resulted in

water vapor condensing inside the EGR cooler leading to a greater thermal effectiveness

recovery by weakening the deposit-metal and deposit-deposit adhesive forces, as

observed in Figure 4.19.

Soot deposits are themselves hydrophobic because of a low H/C ratio [162].

However a surface impurity or a layer of organic hydrocarbon condensate on them can

result in a hydrophilic layer [163,164]. Water has the highest surface tension of all

liquids. If the degree of water saturation in the pores of the deposit layer is sufficiently

high, the capillarity in the pores is strong enough to break the soot clusters which are

connected by weak contact energy [165]. Soot hydration may lead to increasing soot

(4.2)

89

density and a reduction in the fouling factor which leads to refreshing of the EGR cooler

via deposit removal. Essentially, soot hydration can either dislodge the deposits or make

the deposit layer denser improving the thermal conductivity, but which of these processes

is dominant depends on the flow conditions in the EGR cooler.

To examine our results further, some findings from Abarham et al. [80] are

borrowed with permission from the authors. A test rig was designed at University of

Michigan which allowed direct optical access to the deposition process using a digital

video microscope. After fouling the surrogate tubes for 18 hours with 80°C coolant, the

temperature of the coolant was switched to 40°C, and they observed an evidence of water

condensate fracturing the deposit layer, as shown in Figure 4.25. When the temperature

of the coolant was switched to 20°C, there was a very clear evidence of water vapor

condensing, and the condensate appeared to form below the deposit layer, which was

caused the particles in the layer to be removed, as shown in Figure 4.26.

Figure 4.25: 5 minute time interval snapshots of deposit layer removal due to water

vapor condensation; shiny regions reveal the metal surface [80]

90

Figure 4.26: 5 minute time interval snapshots showing water vapor condensate

forming droplets below the deposit layer and the subsequent removal of the deposit

layer [80]

4.4.4 Effect of Engine Startup and Shutdown on EGR Cooler Recovery

Results from the previous section confirmed that low coolant temperatures

promoted EGR cooler recovery, especially during engine start-up. To investigate this, a

test procedure was designed to understand the influence of coolant temperature during

engine start-up under two conditions viz. daily engine start-up (starting the engine the

next morning) and start-up after a 2 hour engine shut down, as shown in Table 4.8. For

this experiment, the engine was operated at 2150 rpm and 203 Nm of load, and exhaust

gas was allowed to flow through the model EGR cooler after the engine had reached a

steady state condition and the exhaust gas temperature remained constant. 20°C coolant

temperature was selected to represent the actual temperature of the coolant during cold

start-up, 40°C coolant temperature was selected since it is close to the dew point

91

temperature of water vapor (section 4.4.3), and 85°C coolant temperature is the normal

temperature of the coolant when the engine is fully warmed.

Table 4.8: Test procedure for EGR cooler recovery monitoring

Step 1 Foul tubes for 4 hours. Engine shut down for the day.

Step 2 Engine start-up with coolant temperature at 20°C. Foul the tubes for 1.5 hours.

Engine shut down for 2 hours.


Engine shut down for 2 hours


Engine shut down for 2 hours


Engine shut down for the day.


Engine shut down for 2 hours.


Engine shut down for the day.

Step 8 Engine start-up with coolant temperature at 40°C.

The temperature profiles throughout the test are plotted in Figure 4.27. It can be

seen that the engine exhaust temperature and EGR inlet temperature remained constant

during the entire test, however, the EGR outlet temperature varied depending on the

coolant temperature. During each 1.5 hour test, the EGR outlet temperature increased as a

function of time, which suggested that the tubes were getting fouled, as observed earlier.

92

50

100

150

200

250

300

350

400

0 100 200 300 400 500 600 700

Tem

pe

ratu

re,

°C

Time, min

Temperature change associated

with restart and EGR cooler refreshment

Figure 4.27: Temperature profiles for EGR cooler recovery test, Engine exhaust

temperature, EGR inlet temperature, EGR outlet temperature

The change in effectiveness with time is plotted in Figure 4.28. During the first 4

hours of operation, the effectiveness of the EGR cooler dropped from 73.6% to 58.7%

and the engine was shut down for the day. When the engine was started the next morning

with a coolant temperature of 20°C, it was observed that the effectiveness improved by

22% and during a subsequent exposure to exhaust gas for 1.5 hours, the effectiveness

dropped to 74.4%. The engine was then shut down for 2 hours and restarted with the

same coolant temperature. The effectiveness improved by around 8% and subsequently

dropped to 77% after 1.5 hours. After 2 hours of engine shut down, the coolant

temperature was switched to 40°C. The effectiveness improved by 2% during this time,

while it dropped further down to 70% at the end of 1.5 hours. After 2 hours of engine

93

shut down, the effectiveness improved by only 1-2%, and subsequently dropped to 64%

at the end of the test. When the engine was restarted the next morning with 85°C coolant,

the effectiveness increased by about 8% and dropped to about 67% at the end of 1.5

hours. After a 2 hour shut down, the effectiveness did not change significantly. The next

morning, the coolant was circulated at 40°C, and it was observed that effectiveness

improved by about 12%. These results show that the EGR cooler experiences a greater

effectiveness recovery during the initial start-up (engine start-up after a long shut down

period of around 8-9 hours) and this recovery is the highest when the temperature of the

coolant is the lowest. The change in effectiveness during this entire experiment is

tabulated in Table 4.9. Similar observations were reported by Mulenga et al. [36].

0.55

0.6

0.65

0.7

0.75

0.8

0.85

0 100 200 300 400 500 600 700

Eff

ecti

ven

ess

, %

Time, min

Foul tubes

for 4 hours

da

ily s

tart

up

at

20

C

Coo

lan

t a

t 2

0C

Coo

lan

t a

t 4

0C

Coo

lan

t a

t 4

0C

da

ily s

tart

up

at

85

C

Coo

lan

t a

t 8

5C

da

ily s

tart

up

at

40

C

Figure 4.28: Effectiveness change versus coolant temperature at engine start-up

94

Table 4.9: Starting and ending EGR cooler effectiveness

Event Starting ℰ, % Ending ℰ, % Δℰ, ± %

Step 1 73.6 58.7 -14.9

Step 2 81.5 74.4 +22.8, -7.1

Step 3 82.4 77.8 +8.0, -5.0

Step 4 79.9 70.2 + 2.1, -9.7

Step 5 71.2 64.4 + 1.0, -6.8

Step 6 72.2 67.4 +7.8, -4.8

Step 7 67.9 63.6 +0.5, -4.3

Step 8 73.0 70.7 +9.4, -2.3

A possible explanation for these observations can be made on the basis of the

temperature inside the EGR cooler. EGR coolers are closed systems as the tubes are not

exposed to ambient conditions. During a long shut down period, the EGR coolers cool

down, and the system temperature remains fairly low. When exhaust gas flows over the

fouled tubes after a long shut down period, the temperature is still low enough for water

vapor to condense inside, which otherwise might not happen when the temperatures are

significantly higher after a 2 hour shut down event. Hence it is evident that thermal

effectiveness improvement is strongly affected by the coolant temperature. One potential

strategy to enhance the performance of the EGR coolers would be to circulate exhaust gas

through the EGR cooler during engine start-up, which is typically avoided in on-road

engine ECU calibrations. If such a calibration were to be employed, the aftertreatment

system needs to be capable of performing at high efficiency even at low temperatures, as

EGR during start-up tends to increase HC and PM emissions.

95

4.4.5 Effect of EGR Oxidation Catalyst on Temperature and Effectiveness Change

for Cruise and Near-idle Conditions

From the deposit analysis in the previous sections, it was determined that the EGR

cooler deposits are comprised mostly of heavy aliphatics (C17-C25). These aliphatics

typically arise from incomplete oxidation of diesel fuel and partly due to the lubricating

oil. It is hypothesized that the oxidation catalyst will reduce the deposition of these

aliphatic hydrocarbons in the EGR cooler. To evaluate the role of an oxidation catalyst,

experiments were performed at the two previously chosen conditions: cruise and near-

idle, with and without an EGR oxidation catalyst. In these experiments, the tubes were

not removed from the assembly and the experiment was completed in one single run of 9

hours. This would result in small changes in thermal effectiveness (no recovery every 90

minutes as reported earlier). An experiment was performed to determine the difference in

effectiveness with and without tube removal and it was found that the net change was

only about 5-6% (Appendix B).

Figure 4.29 and Figure 4.30 show the temperature profiles for the high and low

load conditions respectively. At the high load condition, the catalyst increases the EGR

inlet temperature by about 8-10°C due to heat release on oxidation, however, no

significant change is observed in the EGR inlet temperature at the low load condition

with or without the catalyst. The tubes appear to foul similarly irrespective of the catalyst

as seen from the temperature profiles.

96

100

150

200

250

300

350

400

0 100 200 300 400 500 600

Tem

pe

ratu

re,

°C

Time, min

Engine Exhaust

EGR Inlet

EGR Outlet


203 Nm, without catalyst: engine exhaust, EGR inlet, EGR outlet, with

catalyst: engine exhaust, EGR inlet, ▲ EGR outlet

100

150

200

250

300

0 100 200 300 400 500 600

Tem

pe

ratu

re,

°C

Time, min

Engine Exhaust

EGR Inlet

EGR Outlet


81 Nm load, without catalyst: engine exhaust, EGR inlet, , EGR outlet, with

catalyst: engine exhaust, EGR inlet, ▲ EGR outlet

97

The thermal effectiveness profiles for the high load and low load conditions are

shown in Figure 4.31 and Figure 4.32 respectively. The error bars represent the

confidence interval of the experimental measurements taking into account tube removal

in the ‘with catalyst’ experiment. The effectiveness change is similar with and without

the catalyst at the high load condition, however, the low load condition experiences lower

effectiveness drop in the presence of the catalyst, though the change is not very

significant. The shapes of the curves are also similar indicating that the rate at which

fouling occurs are similar. At high load, the curve drops rapidly initially but stabilizes

quickly, while at low load, there is a gradual drop in the effectiveness which stabilizes

eventually.

-20

-15

-10

-5

0

5

0 100 200 300 400 500 600

Eff

ecti

ven

ess

Ch

an

ge,

%

Time, min


effectiveness change at 2150 rpm, 203 Nm, without ECAT, with ECAT

98

-20

-15

-10

-5

0

5

0 100 200 300 400 500 600

Eff

ecti

ven

ess

Ch

an

ge,

%

Time, min


effectiveness change at 1400 rpm, 81 Nm, without ECAT, with ECAT

Hoard et al. [166] observed that the effectiveness change was about 15% lower in

the presence of an EGR oxidation catalyst, as shown in Figure 4.33. Explanations for

such a significant difference in results between the literature study and this work are

1. Differences in the space velocity and metal loading between the catalysts could have

resulted in differences in effectiveness change. Unfortunately, there is no technical

data about the catalyst from our study limiting comparisons.

2. Our experimental conditions produced a maximum inlet temperature of about 250-

270°C across the ECAT, which is not significant for the catalyst to perform at the

maximum conversion efficiency, while their engine condition resulted in an inlet

temperature much higher than 300°C resulting in an improved catalyst conversion

99

efficiency. It is understood that the EGR oxidation catalyst has a light-off temperature

close to 200°C.

3. The fuel used in our experiment was an ultra-low sulfur diesel with a sulfur

concentration <20 ppm. The fuel used by Hoard et al. [166] had a higher

concentration of sulfur (400 ppm) which could have resulted in the differences in the

condensation of organics on the deposit layer. If the fuel contains sulfur, PM may

contain SO3, sulfuric acid, or sulfates [132–134]. Zhao et al. showed that the number

concentration of nanoparticles for low sulfur fuel decreases compared to high sulfur

diesels, and the accumulation and coarse mode particles contribute to net increase in

the mass of the deposits [167].

4. Our experiments were performed for a total of 9 hours, however, Hoard et al. [166]

evaluated the performance of the catalyst over 200 hours. The duration of the test

could have resulted in differences in experimental findings.

Figure 4.33: Effect of ECAT on effectiveness change at high speed condition [166]

100

The average deposit mass at the end of 9 hours is plotted in Figure 4.34. Change

in the deposit mass due to tube removal has been incorporated in the error bar. It can be

seen that there is no significant change in deposit mass for the high load condition with or

without catalyst, however, at the low load condition, the catalyst seems to have reduced

the deposit accumulation, though the change is not too significant. Similar observations

were reported by Sluder et al. [148] who found that the oxidation catalyst had the least

effect on the deposit mass gain at the 85°C coolant condition, while its use was the most

beneficial at low coolant temperatures.

0

20

40

60

80

100

2150 rpm, 203 Nm 1400 rpm, 81 Nm

Av

era

ge D

ep

os

it M

as

s, m

g

Figure 4.34: Effect of engine operating condition on average deposit mass,

without ECAT, with ECAT

The deposit microstructure with and without the catalyst for the two operating

conditions are shown in Figure 4.35. From these images, it appears that there are no

obvious differences in the deposit microstructure with or without a catalyst.

101

Without Catalyst With Catalyst

a)

b)

c)

d)

Figure 4.35: Effect of engine operating condition on deposit microstructure, a) 2150

rpm, 203 Nm without ECAT, b) 2150 rpm, 203 Nm with ECAT, c) 1400 rpm, 81Nm

without ECAT, d) 1400 rpm, 81 Nm with ECAT

The chemical composition of the deposits from the high and low load conditions

are plotted in Figure 4.36 and Figure 4.37 respectively. At the high load condition, it can

be observed that the catalyst marginally increased the percentage of aromatics and

reduced the net aliphatic content. This suggests that the catalyst oxidized a noticeable

percentage of the light aliphatics (C10-C17) as expected, shifting the distribution toward

102

heavy aliphatics (C18-C25). This however did not result in a significant change in the

mass of the deposits collected in the EGR cooler as observed earlier. This is consistent

with the findings from Sluder et al. [148], who observed a similar distribution shift to the

heavier hydrocarbons as shown in Figure 4.38. At the low load condition, no significant

change was observed in the deposit layer composition. This could be because the exhaust

gas temperature was not high enough for catalytic activity.

0

20

40

60

80

100

Aromatics Total Paraffins C10-C17 C18-C25

Perc

en

tag

e



The effectiveness of the catalyst in oxidizing the HCs is plotted in Figure 4.39

[166]. It can be seen that the catalyst’s efficiency in oxidizing the HCs reduces as a

function of the chain length. In other words, the lighter HCs are more easily oxidized

compared to the heavier HCs. A thorough knowledge of the exhaust gas speciation in

103

terms of the hydrocarbons and their concentrations would have been useful, but was

unavailable due to limitations in the experimental setup.

0

20

40

60

80

100

AromaticsTotal Paraffins C10-C17 C18-C25

Perc

en

tag

e



Figure 4.38: Total ion count with and without ECAT [148]

104

Figure 4.39: Catalyst removal efficiency as a function of aliphatic chain length [166]

From these experiments, it can be concluded that under the conditions tested, the

catalyst was not very effective and did not reduce fouling in the EGR cooler. From the

deposits collected under high load, it was found that the catalyst oxidized some

percentage of the lighter aliphatic HCs shifting the py-GC distribution towards the

heavier aliphatic HCs. The heavy aliphatics after condensing on the EGR cooler surface

can prevent the condensation of the lighter HCs, which have much lower dew point

temperatures. Reducing the T90 distillation temperature of the diesel fuel by cutting the

heavy-end fractions (high molecular weight) of the diesel fuel would offer a potential

solution to minimize heavy HCs deposition in the EGR coolers. The remaining lighter

HCs can then be further oxidized with an oxidation catalyst. The performance of the

oxidation catalyst at other engine conditions needs to be evaluated, which is not

considered in the scope of this study.

105

4.5 Conclusions

The performance of a small shell and tube heat exchanger with surrogate tubes

was investigated under different engine conditions and coolant temperatures. Fouling had

a significant impact on the performance of the heat exchanger. It was observed that the

heat exchanger effectiveness dropped rapidly initially but stabilized after a few hours of

operation, irrespective of the engine condition. The deposits randomly accumulated on

the tube surfaces initially, but appeared to cover the entire surface within 9 hours of

operation. Thermophoresis and condensation of hydrocarbons played a major role in

fouling, depending on the engine condition and coolant temperature. High EGR inlet

temperatures from high engine loads led to greater thermophoretic soot deposition and

higher effectiveness loss. Different engine conditions produced different deposit

microstructures, resulting in different rates of effectiveness loss. Additionally, these

deposits varied in the net aromatic content and aliphatic content at different engine

conditions. From amongst all the deposits collected, it appeared that the aliphatic

contribution of the deposits is mostly due to long chain compounds (C17-C25), which are

from the heavy end of the diesel fuel.

Coolant temperature played a significant role in altering the nature of the deposits

in the EGR cooler, due to a greater extent of hydrocarbons condensing at low coolant

temperatures. This was evident from the high VOF content in the deposits at low coolant

temperatures. Low coolant temperatures promoted greater thermal effectiveness recovery

on start-up, due to water vapor condensing during engine shutdown periods. It was

106

evident that starting the engine by circulating the coolant to the EGR at low temperatures

provided a key solution to EGR cooler recovery, which could be adopted in the future.

From the work on EGR oxidation catalyst, it was envisioned that the catalyst will

reduce the deposition rate and deposit mass by oxidizing the hydrocarbons, but it was not

effective under the test conditions selected. Perhaps, the conditions at which the engine

was operated did not produce high enough temperatures for significant catalytic activity.

Or even if the catalyst was active, it was targeting the light hydrocarbons (as seen in

Figure 4.39) which the engine conditions did not produce much, leaving behind the

heavier hydrocarbons in the deposits. One possible solution to reduce the heavy

hydrocarbons in the deposits would be to reduce the T90 distillation temperature of the

diesel fuel to eliminate the heavy hydrocarbons in the fuel itself. This way, we can take

advantage of the catalyst in eliminating the lighter hydrocarbons in the EGR cooler

deposits and minimize fouling.

Nevertheless, the experiments performed provide a significant insight into the

fouling mechanisms leading to EGR cooler deposits, and provides novel information on

the evolution of microstructure and composition of the deposits. Based on the results

obtained, potential solutions for EGR cooler thermal effectiveness recovery were

suggested.

107

Chapter 5

Experimental Studies of High-Efficiency Combustion with Fumigation of

Liquefied Fuels into Diesel Engine Air Intake

5.1 Introduction

Development of high efficiency engines via advanced combustion strategies is

one of the most promising and cost-effective approaches to improving fuel economy of

the US vehicle fleet in the near- to mid-term, and has been a key focus of the DOE’s

Vehicle Technologies Program initiative [168]. Conventional diesel combustion

experiences regions of both rich and lean high-temperatures, forming soot and NOx,

respectively. Soot emissions can be effectively reduced through the use of a DPF;

however, DPFs require periodic regeneration which often increases fuel consumption

[16,17]. Additionally, engines relying on NOx traps or three-way catalysts (TWC) for

NOx emissions reduction must periodically operate rich to reduce the stored NOx; thus

reducing the fuel economy [14,15]. Hence it is clear that for overall engine efficiency,

dependence on aftertreatment systems for NOx and PM emissions needs to be minimized.

Advanced combustion or low temperature combustion is a relatively novel area of

research compared to conventional diesel combustion. Low temperature combustion

occurs at temperatures below those at which NOx forms and at equivalence ratios below

those at which soot forms, as seen in Figure 5.1. Both NOx and soot formation are

avoided if the combustion temperature remains below 1650K. Ensuring good mixing of

fuel and air to form a homogenous charge has been the key principle behind some of

108

these advanced combustion strategies, and thus avoid stratified combustion as with

conventional diesel fuel. A detailed explanation of these combustion strategies was

provided in Chapter 3. Currently, low ignition quality fuels are being studied in advanced

combustion modes to control combustion phasing and allow for higher brake thermal

efficiency. Two examples of such combustion are the partially-premixed combustion

(PPC) and the reactivity controlled compression ignition (RCCI). Hanson et al. [169]

demonstrated 53% net indicated thermal efficiency for a heavy duty engine with RCCI,

while adhering to EPA 2010 NOx regulations.

Figure 5.1: Φ – T map showing soot and NOx formation zones, with advanced

combustion modes, adapted from Dec [170]

Partially Premixed Combustion (PPC) provides the potential of simultaneous

reduction of NOx and soot for diesel engines. In PPC, a part of the fuel is injected early

109

during the compression stroke and then mixed with air to achieve premixed lean

combustion, and the remaining fuel is injected after TDC into the high-temperature

mixture. This eliminates locally rich regions, and the mixture is homogenous compared to

conventional diesel combustion. Moreover, combustion can be controlled by adjusting

the fuel injection timing. This results in improved engine efficiency and reduced

emissions [12,101,171,172]. Additional details can be found in Chapter 3.

RCCI combustion uses in-cylinder fuel blending with at least two fuels of

different reactivity and multiple injections to control in-cylinder fuel reactivity to

optimize combustion phasing, duration and magnitude [108,173]. The process involves

introduction of a low reactivity fuel into the cylinder to create a well-mixed charge of low

reactivity fuel, air and recirculated exhaust gases. The high reactivity fuel is injected

before ignition of the premixed fuel occurs, using single or multiple injections directly

into the combustion chamber. The mixing of the two fuels results in combustion taking

place at lower temperatures, which improves engine efficiency (due to lower heat transfer

losses) and lowers NOx emissions [174]. Exploring alternate blends of fuels under

different combustion strategies capable of improving engine efficiency and reducing

emissions is thus of paramount importance.

110

5.2 Hypothesis for High Efficiency Diesel Combustion

Propane can delay the autoignition of DME and improve the combustion

characteristics of diesel fuel in mixed mode combustion, without a detrimental effect

of high pressure rise and heat release rates in the engine.

5.3 Objectives for High Efficiency Diesel Combustion

Demonstrate high thermal efficiency using liquefied gases fumigated into a diesel

engine.

The objective of this task is to demonstrate that fumigating liquefied fuels into

diesel engine air intake to achieve dual fuel combustion can improve the brake thermal

efficiency of a diesel engine, compared to operating the engine on diesel fuel alone, while

holding the engine speed and load constant. To achieve this objective, different

proportions of DME and propane (10-40%) are fumigated into the air intake, each

combination representing the percentage energy equivalent substitution of diesel fuel in

combination with rest of the fuel being ultra-low sulfur diesel.

Examine the role of ignition quality of a fumigated fuel on combustion phasing and

thermal efficiency.

The objective of this task is to achieve dual combustion, with fuels of different

reactivity, in order to control the combustion and it’s phasing. In this experiment, propane

(poor ignition quality) is fumigated along with DME (good ignition quality) to suppress

111

the low temperature heat release event of DME, to shift the combustion phasing closer to

the top dead center. Engine performance is evaluated in terms of brake thermal

efficiency, brake specific energy consumption, pressure rise and heat release rates and

engine-out emissions.

5.4 Experimental

5.4.1 Engine

The test engine was a four cylinder turbo-charged, common rail diesel engine

whose specifications are shown in Table 5.1. The engine was coupled to a water cooled

250Hp Eaton eddy current dynamometer. Engine operating parameters, in particular the

injection parameters, were controlled using an unlocked electronic control unit (ECU)

which was connected to an ETAS MAC 2 interface via an ETK connection, which was

linked to a computer running INCA v5.0 software. Cylinder pressure signals were

measured using AVL GU12P pressure transducers. Needle lift data were obtained from a

Wolff Controls, Inc. Hall Effect needle lift sensor, which was placed on the injector of

cylinder 1. Diesel fuel consumption rate was calculated by measuring fuel tank weight

continuously, and mass flow rate of air into the engine was measured using a mass air

flowmeter. The temperature and humidity of the intake air was monitored through-out the

test and these values did not change significantly, and thus no corrections for emissions

were necessary.

112

Table 5.1: Engine specifications

Engine 2.5 L DDC/VM-Motori

No. of valves 4 valves/cylinder

Compression

ratio

17.5

Rated power 103 kW@ 4000 rpm

Peak torque 340 N-m@1800 rpm

Injection system Common rail, direct injection

5.4.2 Fuel

For these experiments, the baseline fuel was a certified ultra-low sulfur diesel

(ULSD) supplied by ChevronPhilips Chemical, details of which can be found in Table

5.2. As the experiments required two gases (DME and propane) to be mixed with the air

intake in specific proportions, it was important to set up the mixing process in a manner

that would be easy to control while at the same time ensuring a homogeneous mixture.

DME and propane were fumigated into the engine air intake through a custom designed

manifold, which consisted of four porous metal filters, customarily used as spargers, as

shown in Figure 5.2. The flowmeter used was a Matheson Gas FM7410 series flowmeter

capable of flowing 4 gases. Each flowmeter was calibrated for the specific gas flowing

through them at the required pressure and temperature. The DME tank was fitted with

tank heaters and pressure and temperature monitoring were added to maintain a constant

pressure and temperature of the fuel delivery.

113

Table 5.2: Specifications of ultra-low sulfur diesel fuel

Properties Diesel

Cetane number 45

Oxygen content (%) 0

Specific gravity 0.8466

Heating value (MJ/kg) 42.8

Sulfur content (ppm) 9.7

Figure 5.2: Custom intake air manifold system for DME and propane fumigation

5.4.3 Emissions

An AVL Combustion Emissions Bench II was used to measure gaseous

emissions. Hot exhaust gases were sampled from the engine’s exhaust pipe by headline

filters through heated lines kept at 195ºC. NOx emissions were measured using an

EcoPhysics chemiluminescence analyzer. CO and CO2 emissions were measured by two

separate Rosemount infrared analyzers, while the total hydrocarbons were measured

114

using ABB Flame Ionization detectors. Unfortunately, there were no PM mass

measurements for these tests.

5.4.4 Test Matrix

Different proportions of DME and propane were fumigated into the engine air

intake, and diesel fuel was injected through the common rail fuel injectors directly into

the cylinders. These proportions were selected after performing a preliminary study of the

maximum substitutions of DME and propane, to permit safe operation of the engine

without knocking. The changes in DME and propane are made in steps of 10%. A

preliminary study with 5% increments did not result in a significant change in the engine

performance, and was a result of the uncertainty in the measurement of the flow rates in

the flowmeter. The test matrix adopted for the test is shown in Table 5.3. Each operating

point represents a percentage of energy supplied through the fumigated fuels, with rest of

the energy coming from the diesel fuel. The case identified as 0% DME and 0% propane

represents baseline diesel engine condition. The engine was operated at 1800 rpm, under

25% load, and the injection strategy was limited to a single injection for ease of

combustion control. This condition represents medium speed and light load, and was

selected based on past experiments performed on the engine, which is very stable upon

reaching steady-state [113]. The injection timing was held steady at 7° before the top

dead center (TDC) by programming the engine’s ECU. Exhaust gas recirculation (EGR)

was disabled to eliminate its effect on engine performance and emissions. In this work,

the terms LPG and propane are used interchangeably, but represent the same fuel.

115

Table 5.3: Percentage of DME and propane energy fumigated into engine air intake

%Total

Substitution

Propane

0% 10% 20% 30%

DM

E

0% 0 10 20

30

10% 10 20 30

40

20% 20 30 40 50

30% 30 40 50 60


5.5.1 Effect of DME and Propane Fumigation on BTE and BSEC

BTE and BSEC values for each operating condition are shown in Figure 5.3 and

Figure 5.4 respectively. The general trend observed indicates an increase in the BTE

values with increasing DME substitution, though the change is not too significant. It can

be observed that BTE increases are greater with increasing propane than with increasing

DME. The maximum BTE was observed at 20% DME and 30% propane substitution.

The efficiency at this point was found to be 49.91% which is almost 12-13% points

greater than the BTE for the baseline diesel condition with no substitution. This appeared

to be a ‘sweet-spot’ among different combinations of DME and propane. Similarly, the

BSEC values appeared to decrease towards the right end of the graph with increasing

DME and propane substitutions. Overall, 24% less energy is required when the engine is

operated with 50% of the diesel energy substituted with 20% DME and 30% propane as

compared to baseline diesel condition.

116

To understand the improvements in engine efficiency (BTE), indicated thermal

efficiency (ITE) and engine frictional power values were calculated and plotted in Figure

5.5 and Figure 5.6 respectively. Gross indicated power (IP) is calculated by integrating

the area under p-v diagram during the compression and expansion strokes and friction

power (FP) is calculated by subtracting the brake power from the indicated power. ITE

increases more with propane substitution than with DME substitution, and follows the

same trend as BTE. DME and propane substitution results in slightly lower FP compared

to the baseline diesel condition perhaps due to shorter diesel injection durations, however,

increasing substitutions of DME and propane have no significant effect on FP. The

reason why 10% DME and 40% propane substitution has a high FP is unknown, and is

assumed to be a result of an erroneous pressure trace.

Pumping losses, which contribute to friction power in the engine, can reduce if

lower amounts of air are being drawn into the engine, assuming that the fuel was still

under the same pressure at the common rail as it were when diesel fuel alone was being

burned. However, this was not the case in our experiments as the mass of air flow at all

conditions remained identical, and the volumetric efficiency of the engine did not change

(less than 1%) much with the operating condition, as observed in Figure 5.7. (The

volumetric efficiency of naturally aspirated engines is about 70-80%. Since this engine is

turbocharged, the volumetric efficiency is higher. Volumetric efficiency can be higher

than 100%). Additionally, the fuel rail pressure remained constant and identical between

all the test conditions. These results confirm that the improvement in engine efficiency is

not due to reductions in friction power.

117

Kokjohn et al. [108] note that the increase in indicated power (thus brake power

and BTE) with RCCI combustion is due to a reduction in the heat transfer losses from

lower peak cylinder temperatures. Using numerical simulations, the authors found that

under identical engine speed and load condition, conventional diesel combustion and

RCCI resulted in ‘local’ peak cylinder temperatures of ~2800K and ~1700K respectively,

even though the ‘global’ or bulk-averaged cylinder temperature for RCCI was higher than

that for diesel. RCCI resulted in 43% lower heat transfer losses compared to conventional

diesel combustion (8.2% less of fuel energy is lost to heat transfer). From our

experiments, it is assumed that the improvement in engine efficiency is partly due to a

reduction in the local peak cylinder temperatures, though the actual values cannot be

determined.

20

25

30

35

40

45

50

55

0D

,0P

10D

,0P

10D

,10

P

10D

,20

P

10D

,30

P

10D

,40

P

20D

,0P

20D

,10

P

20D

,20

P

20D

,30

P

30D

,0P

30D

,10

P

30D

,20

P

30D

,30

P

Bra

ke

Th

erm

al

Eff

icie

ncy

, %

Figure 5.3: Brake thermal efficiency at varying DME and propane substitution

levels (D=DME, P=Propane)

118

4

5

6

7

8

9

10

11

0D

,0P

10D

,0P

10D

,10

P

10D

,20

P

10D

,30

P

10D

,40

P

20D

,0P

20D

,10

P

20D

,20

P

20D

,30

P

30D

,0P

30D

,10

P

30D

,20

P

30D

,30

P

BS

EC

, M

J/k

Wh

Figure 5.4: Brake specific energy consumption at varying DME and propane

substitution levels (D=DME, P=Propane)

0

10

20

30

40

50

60

0D

,0P

10D

,0P

10D

,10

P

10D

,20

P

10D

,30

P

10D

,40

P

20D

,0P

20D

,10

P

20D

,20

P

20D

,30

P

30D

,0P

30D

,10

P

30D

,20

P

30D

,30

P

Ind

icate

d T

herm

al E

ffic

ien

cy

, %

Figure 5.5: Indicated thermal efficiency at varying DME and propane substitution


119

0

1

2

3

4

5

0D

,0P

10D

,0P

10D

,10

P

10D

,20

P

10D

,30

P

10D

,40

P

20D

,0P

20D

,10P

20D

,20P

20D

,30P

30D

,0P

30D

,10P

30D

,20

P

30D

,30

P

Fri

cti

on

Po

we

r, h

p

Figure 5.6: Frictional power at varying DME and propane substitution levels

(D=DME, P=Propane)

50

60

70

80

90

100

110

0D

,0P

10D

,0P

10D

,10

P

10D

,20

P

10D

,30

P

10D

,40P

20D

,0P

20D

,10P

20D

,20

P

20D

,30

P

30D

,0P

30D

,10P

30D

,20

P

30D

,30

P

Vo

lum

etr

ic E

ffic

ien

cy

, %

Figure 5.7: Volumetric efficiency at varying DME and propane substitution levels

(D=DME, P=Propane)

120

Figure 5.8 through Figure 5.11 show plots of the cylinder pressure at different

substitutions of DME and propane. Under baseline diesel condition, the peak cylinder

pressure is about 55 bar. Addition of DME tends to increase the peak cylinder pressure

due to early autoignition of the fuel. This peak pressure, however, reduces with

increasing propane, which appears to introduce a second peak in the pressure curve which

is absent when only DME is present in the cylinder. The reduction in pressure is of the

order of 8-10 bar. Addition of propane delays the occurrence of the peak cylinder

pressure. For the operating condition at 20% DME and 30% propane, the increase in the

peak cylinder pressure compared to baseline diesel fuel is only about 6-7 bar, which is

quite nominal. Other conditions experienced a much higher peak cylinder pressure, with a

maximum for 30% DME substitution. For dual fuel combustion to be viable, the peak

cylinder pressure should be lower than the maximum allowable limit especially at high

engine loads, and the engine must be designed to withstand these pressures without

compromising engine integrity. Unfortunately, no experiments have been performed at

other engine conditions to monitor the rise in cylinder pressure.

121

10

20

30

40

50

60

70

80

-30 -20 -10 0 10 20 30

Baseline0D, 10P0D, 20P0D, 30P0D, 40P

Cy

lin

de

r P

ress

ure

, b

ar

Crank Angle, deg


40% propane substitution (D=DME, P=Propane)

0

10

20

30

40

50

60

70

80

-30 -20 -10 0 10 20 30

Baseline10D, 0P10D, 10P10D, 20P10D, 30P10D, 40P

Cy

lin

de

r P

ress

ure

, b

ar

Crank Angle, deg



122

0

10

20

30

40

50

60

70

80

-30 -20 -10 0 10 20 30

Baseline

20D, 0P

20D, 10P

20D, 20P

20D, 30P

Cy

lin

de

r P

ress

ure

, b

ar

Crank Angle, deg



0

10

20

30

40

50

60

70

80

-40 -30 -20 -10 0 10 20 30 40

Baseline

30D, 0P

30D, 10P

30D, 20P

30D, 30P

Cy

lin

de

r P

ress

ure

, b

ar

Crank Angle, deg



123

The heat release profiles are plotted in Figure 5.12 through Figure 5.15. From

these figures, it can be observed that DME exhibits the typical two stage heat release

process, a low temperature heat release (LTHR) and a high temperature heat release

(HTHR) event, which is a characteristic of the fuel and arises due to its combustion

chemistry. With increasing DME concentration, both low and high temperature heat

release peaks increase, while the combustion phasing is held constant. Addition of

propane to this mixture tends to delay the onset of ignition and shifts the DME

combustion process closer to TDC. Consider for instance, for the condition with 20%

DME and no propane, two peaks from DME combustion are observed (Figure 5.14), a

low temperature heat release peak at around 25° before TDC and a high temperature heat

release peak at 12° before TDC, with the rest being diffusion burn from diesel. But as

propane substitution is introduced in steps of 10%, it can be observed that the low

temperature peak is delayed and decreased in amplitude until at 30% each DME and

propane substitution there is a single heat release peak at 5° after TDC, essentially

shifting the entire DME combustion process closer to the TDC. Since all the heat from

DME and propane is being released closer to the TDC, the bulk cylinder temperature also

increases which results in diesel fuel combusting earlier.

Referring to Figure 5.14, under baseline diesel condition (no DME/propane)

diesel fuel ignition occurs at 8° after TDC with an ignition delay period of about 15°,

considering that the fuel was injected 7° before TDC. At 20% DME and 30% propane

substitution, diesel ignition occurs at around 1° after TDC, even though diesel fuel is still

being injected at 7° before TDC. This implies that diesel combustion advances due to an

124

increase in the cylinder temperature. At 30% DME substitution, diesel fuel burns much

before TDC, even before the piston is at the minimum volume, which results in lower

peak power output and decreased BTE as observed earlier, because this represents work

done against the piston. In theory, maximum engine efficiency is observed when the

entire fuel energy is released during maximum compression, or in other words, when the

piston is close to TDC. These results prove the hypothesis that propane can delay the

onset of DME combustion, altering the ignition quality of the fuel mixture and improving

the combustion phasing of the intake charge.

The bulk-averaged cylinder temperature plotted in Figure 5.16 through Figure

5.19 show unique trends. From Figure 5.16, it can be seen that increasing propane

percentage tends to reduce the mean cylinder temperature. With addition of 10% DME

(Figure 5.17), there is a slight increase in temperature before TDC compared to the

baseline diesel condition due to autoignition of DME, and addition of propane in

increments of 10% reduces this temperature in the cylinder due to poor autoignition

characteristics of propane.

With further addition of DME (say, 20%), a peculiar trend is observed. For one,

there are two small peaks before the top dead center, each representing the low

temperature and high temperature heat release events from DME, and this is suppressed

by the addition of propane. However, after the TDC, the bulk cylinder temperature is

greater than the baseline condition, and addition of propane actually increases this

temperature. This trend is mostly due to diesel fuel also burning earlier compared to the

baseline diesel condition. Similar trend is observed at 30% DME addition. This increase

125

in the temperature after the top dead center represents all the fuel being burned at once,

resulting in an improvement in engine efficiency, as observed previously.

-20

0

20

40

60

80

100

-20 -10 0 10 20 30 40

Baseline0D, 10P0D, 20P0D, 30P0D, 40P

Ap

pare

nt

Hea

t R

ele

as

e R

ate

, J

/de

g

Crank Angle, deg



126

-20

0

20

40

60

80

100

-40 -20 0 20 40

Baseline10D, 0P10D, 10P10D, 20P10D, 30P10D, 40P

Ap

pare

nt

Hea

t R

ele

as

e R

ate

, J

/de

g

Crank Angle, deg



-20

0

20

40

60

80

100

-40 -20 0 20 40

Baseline

20D, 0P

20D, 10P

20D, 20P

20D, 30P

Ap

pare

nt

Hea

t R

ele

as

e R

ate

, J

/de

g

Crank Angle, deg



127

-20

0

20

40

60

80

100

-40 -20 0 20 40

Baseline

30D, 0P

30D, 10P

30D, 20P

30D,30P

Ap

pare

nt

Hea

t R

ele

as

e R

ate

, J

/de

g

Crank Angle, deg



700

800

900

1000

1100

1200

1300

1400

1500

-20 -10 0 10 20 30 40 50

Baseline Diesel0D, 10P0D, 20P0D, 30P0D, 40P

Bu

lk-a

vera

ge

d T

em

pe

ratu

re, K

Crank Angle, deg


substitution and 0 – 40% propane substitution (D=DME, P=Propane)

128

700

800

900

1000

1100

1200

1300

1400

1500

-20 -10 0 10 20 30 40 50

Baseline Diesel10D, 0P10D, 10P10D, 20P10D, 30P10D, 40P

Bu

lk-a

vera

ge

d T

em

pe

ratu

re, K

Crank Angle, deg



700

800

900

1000

1100

1200

1300

1400

1500

-20 -10 0 10 20 30 40 50


Bu

lk-a

vera

ge

d T

em

pe

ratu

re, K

Crank Angle, deg



129

700

800

900

1000

1100

1200

1300

1400

1500

-20 -10 0 10 20 30 40 50


Bu

lk-a

vera

ge

d T

em

pe

ratu

re, K

Crank Angle, deg



5.5.2 Effect of DME and Propane Fumigation on Emissions

5.5.2.1 Total Hydrocarbon Emissions

Total hydrocarbon emissions (THC) at different operating conditions are plotted

in Figure 5.20. Compared to the baseline diesel condition, all substitutions have net

higher engine-out hydrocarbon emissions. THC emissions at 10% DME and 40%

propane are supposed to be much higher; however, the emissions bench was calibrated to

a span gas which could read up to 3250 ppm. It can be seen that THC emissions increased

with increasing propane substitution at constant DME levels. This could be due to two

reasons, 1) pockets of propane failing to ignite and passing through the cylinder

130

unreacted and 2) due to the lower reactivity and higher autoignition temperature of

propane which results in incomplete oxidation of propane. The increase in HC emissions

is a common observation when increasing the premixing of fuel, as observed in most

HCCI, RCCI, and PCCI combustion strategies [88,173,175].

0

500

1000

1500

2000

2500

3000

35000

D,0

P

10D

,0P

10D

,10

P

10D

,20

P

10D

,30

P

10D

,40

P

20D

,0P

20D

,10

P

20D

,20

P

20D

,30

P

30D

,0P

30D

,10

P

30D

,20

P

30D

,30

P

To

tal

HC

Em

iss

ion

s, p

pm

Figure 5.20: Total hydrocarbon emissions at varying DME and propane substitution


In comparison, THC emissions reduced with increasing DME substitutions at

fixed propane substitution, which is a direct consequence of shorter ignition delay period

for DME (DME is already in vapor phase on entering the cylinder) leading to smaller

over-rich and over-lean mixture regions. Aceves and Flowers have also documented

similar effects with DME and methane [130]. Ikeda et al. [176] observed that THC

131

emissions from fumigating DME with diesel fuel reduced in comparison to diesel

combustion alone.

5.5.2.2 NOx Emissions

Figure 5.21 shows the plot of NOx emissions at various DME and propane

substitutions. Since the brake power of the engine was fixed for all substitutions (brake

power is dependent on engine speed and load only), it was expected that the brake

specific values followed the same trend as that of the raw emissions. NOx emissions

appear to decrease with increasing energy substitution with DME and propane. The

lowest NOx value occurs at 30% DME and 30% propane substitution which represents a

41% decrease over the value for baseline diesel. Even though for 20% and 30% DME

substitutions, with different substitutions of propane show an increase in the bulk-

averaged cylinder temperature after the top dead center (Figure 5.18 and Figure 5.19),

NOx emissions seems to have reduced, which is counter-intuitive, since it is known that

NOx formation is governed by temperature and the timing of the high temperature in the

cylinder.

NOx emissions from CI engines burning DME vary depending upon the engine

conditions and the fuel supply system, as described in the literature [125]. Chapman and

Boehman [92], who observed similar reductions in NOx emissions with DME fumigation

note that the reduction is not purely due to thermal effects as commonly assumed, but

also due to a reduction in the heat release during the mixing controlled burning of the

diesel fuel. Kajitani et al. [177] reported that the NOx emissions level was higher with

132

DME than with diesel fuel in a CI engine that was tested at the same injection timing

recommended for diesel fueling. This observation was attributed to the longer duration of

the peak combustion temperature in the initial combustion period because of the shorter

ignition delay of DME. Cipolat [178] also observed an increase in NOx emissions with

DME fueling at low engine speeds in comparison to those of diesel fueling at the same

conditions of injection timing and injector opening pressure. On the other hand, when the

operating conditions of the engine were optimized for each fuel, NOx emissions from

DME was lower than that of diesel [179]. Longbao et al. [180] achieved low NOx

emissions from DME compared to those of diesel fuel by adjusting the injection timing.

These differences in NOx emissions arise due to variations in the way DME is introduced

into the combustion chamber i.e., via fumigation or injection.

As discussed above, it can be seen that NOx formation with DME is not entirely

governed by the bulk-averaged cylinder temperature. A ‘back-of-the-envelope’ adiabatic

temperature calculation for DME and diesel fuel at stoichiometric conditions cannot

explain entirely the trends in NOx emissions, as the flame in the cylinder will be

premixed around the region where the fuel is in gaseous state (perhaps from DME and

propane), and diffusion-type where the liquid diesel fuel is present.

From these experiments, it can be concluded that the reduction in NOx emissions

due to DME and propane substitution is due to an increase in the proportion of

homogenous oxidation (similar to the premixed burn in conventional diesel combustion)

which reduces the amount of heat release during the mixing controlled burning of diesel

fuel. This is confirmed in the heat release analysis presented earlier.

133

0

100

200

300

400

500

0D

,0P

10D

,0P

10D

,10

P

10D

,20

P

10D

,30

P

10D

,40

P

20D

,0P

20D

,10

P

20D

,20

P

20D

,30

P

30D

,0P

30D

,10

P

30D

,20

P

30D

,30

P

NO

x E

mis

sio

ns,

pp

m

Figure 5.21: NOx emissions at varying DME and propane substitution levels

(D=DME, P=Propane)

5.5.2.3 CO and CO2 Emissions

CO emissions plotted in Figure 5.22 tend to increase with increasing propane and

DME substitutions, and all substitutions have higher CO emissions than the baseline

diesel condition. This could be the result of incomplete combustion of propane, which

produces CO instead of CO2. When propane substitution is increased with 20%

substitution of DME, CO emissions remain around the same and actually decrease at one

point.

From the literature, CO emissions show some variations depending on the engine

system and operating conditions. Since DME injections are usually longer and at lower

pressures compared to diesel injection, CO emissions may increase [125]. Additionally,

134

DME fuel spray can impinge on the cooled combustion chamber wall which could

potentially raise HC and CO emissions by quenching the DME reaction process [181].

However, these reasons may not be responsible for increase in CO emissions in these

experiments as DME was fumigated into the engine air intake and not injected via the

fuel injection system. Ofner et al. noted that depending on the reaction process, a larger

amount of CO may be produced compared with diesel fuel, since there is production of

CHO and CH2O radicals involved in the combustion of DME [118]. CO may also be

produced in locations of over-lean conditions. Due to the faster vaporization of DME, the

local equivalence ratio can sometimes become too low to support combustion, which may

result in the increase of CO emissions [182,183]. Typically, CO emissions from internal

combustion engines are dominated by the air-fuel ratio. Since diesel combustion takes

place under over-lean mixtures, CO emissions are usually not that important.

Nevertheless, the A/F ratio for these conditions is shown in Figure 5.23. It can be

observed that A/F ratio increases with propane substitution and decreases with DME

substitution, which perhaps are a direct consequence of the fuel utilization due to

differences in fuel ignition behavior.

As seen earlier, CO emissions from engines operating with DME have

contradicting trends. From our experiments, the increase in CO emissions due to DME

and propane are mostly due to incomplete combustion of the gaseous reactants which

might result in locally rich mixtures. CO emissions reduce beyond 20% DME

substitution due to an increase in the bulk cylinder temperature which favors complete

oxidation of the fuel as seen in Figure 5.18 and Figure 5.19. This increase in bulk

135

cylinder temperature after 20% DME is due to diesel fuel also igniting earlier compared

to the baseline condition. Similar trends of CO emissions increasing initially and

decreasing later on have been reported by Salsing and Denbratt [184].

CO2 emissions plotted in Figure 5.24 appear to decrease mainly with increasing

propane substitution while they remain more or less constant with increasing DME. The

increased reactivity of DME leads to its oxidation into CO2 and H2O whereas propane

remains unburned to an extent and thus contributes in decreasing CO2 levels but

increasing HC and CO levels as observed earlier.

0

500

1000

1500

2000

2500

3000

3500

4000

0D

,0P

10D

,0P

10D

,10

P

10D

,20

P

10D

,30

P

10D

,40

P

20D

,0P

20D

,10

P

20D

,20

P

20D

,30

P

30D

,0P

30D

,10

P

30D

,20

P

30D

,30

P

CO

Em

iss

ion

s, p

pm

Figure 5.22: CO emissions at varying DME and propane substitution levels

136

25

30

35

40

45

50

0D

,0P

10D

,0P

10D

,10

P

10D

,20

P

10D

,30

P

10D

,40

P

20D

,0P

20D

,10

P

20D

,20

P

20D

,30

P

30D

,0P

30D

,10

P

30D

,20

P

30D

,30

P

A/F

Rati

o

Figure 5.23: Air-Fuel ratio at varying DME and propane substitution levels

3

3.5

4

4.5

5

5.5

0D

,0P

10D

,0P

10D

,10P

10D

,20P

10D

,30P

10D

,40P

20D

,0P

20D

,10P

20D

,20P

20D

,30P

30D

,0P

30D

,10

P

30D

,20

P

30D

,30

P

CO

2 E

mis

sio

ns,

%

Figure 5.24: CO2 emissions at varying DME and propane substitution levels

137

5.6 Conclusions

This work was intended to develop strategies for high efficiency combustion

(moving toward 55% BTE) and demonstrate that fumigating DME and propane into the

intake air to modify the cetane number of the fuel would lead to a better control over the

combustion process.

BTE values marginally increased with increasing DME substitution, while the

increase was higher with propane substitution. The increase in engine efficiency was

assumed to be a result of reduced peak cylinder temperatures locally and improved

combustion phasing with fumigation compared to conventional diesel combustion. A

maximum of ~49% BTE was observed at 20% DME and 30% propane substitution,

compared to a baseline efficiency of 37%, and this substitution resulted in a 24%

reduction in BSEC. This particular condition had a pressure rise of only 6-7 bar in the

cylinder, which was minimal and easily withstood by the engine. In general, it was

observed that the peak cylinder pressure was a maximum at high DME and low propane

substitution values. DME exhibited a two stage heat release process, while propane

substitution delayed DME’s early autoignition and shifted the combustion process closer

to TDC. The net effect of this was seen on diesel fuel igniting much earlier compared to

the baseline diesel condition (8° advanced compared to the baseline diesel condition at

the peak BTE).

Total hydrocarbon emissions increased with increased propane substitution and

decreased with DME substitution. The increase with propane was more substantial due to

138

a high autoignition temperature. NOx emissions reduced with increasing DME and

propane substitution due to an increase in the premixed autoignition combustion resulting

in less diffusion burn fraction from diesel fuel. CO emissions increased with increasing

propane and DME due to incomplete combustion. After 20% DME substitution, CO

emissions reduced due to an increase in the bulk-averaged cylinder temperature

promoting complete oxidation of the fuel. CO2 emissions appeared to decrease mainly

with increasing propane substitution while they remained more or less constant with

increasing DME substitution.

139

Chapter 6

Performance Evaluation of Ultra-Low Sulfur Diesels

6.1 Introduction

Fuel formulation and additives can have dramatic impacts on efficiency and

emissions performance of engines and vehicle systems. Removal of lead from gasoline in

the 1970’s, reformulation of gasoline’s in the 1980’s and 1990’s, and implementation of

ultra-low sulfur fuels in the last decade each enabled cleaner vehicles and reduced air

pollution as a consequence. The current interest in improving the fuel mileage of

passenger cars (in response to more strict CAFE standards) and trucks has sparked

significant efforts to increase the efficiency of engines and vehicle powertrains. Fuel

improvements have enabled cleaner vehicles, for example, ultra-low sulfur fuels have

enabled the introduction of clean diesel technology. But, one can ask the question of

whether fuel formulation can serve to improve thermal efficiency of engines directly and

determine if fuels should evolve to meet the future requirements of advanced combustion

strategies. The motivation for this work stems from the DOE’s Vehicle Technologies

SuperTruck Program, whose goal is to design a heavy-duty Class 8 truck which

demonstrates 55% BTE and a 50% improvement in overall freight efficiency measured in

ton-miles per gallon [86,185]. This chapter describes a study in which this question of

improving engine efficiency is addressed through a comprehensive fuel formulation

study.

140

An internal combustion engine can be analyzed as an open system which

exchanges heat and work with its surrounding environment (atmosphere). Reactants (fuel

and air) flow into the system and products of combustion (exhaust gases) flow out. A

fundamental measure of effectiveness of any practical IC engine is measured in terms of

exergy or availability, which is the maximum useful work transfer that can be obtained

from a system-atmosphere combination at a given state [18]. Availability is destroyed by

the irreversibilities that occur in any real process. Since estimating availability is difficult,

a more convenient definition called the fuel conversion efficiency is often used in

thermodynamic calculations relevant to IC engines.

Fuel conversion efficiency or thermal efficiency [18] is measured by how much of

the original chemical energy in the fuel is converted to mechanical work as shown in

Equation (6.1).

where,

: mass of fuel inducted per cycle

QHV: lower heating of the fuel

P: indicated or brake power

Maximum fuel efficiencies of IC engines are still well below the theoretical

potential. Today, typical BTE for CI engines are in the range of 35-40% depending on the

class and year of the engine [18]. BTEs higher than 60% are not possible due to inherent

(6.1)

141

irreversibility of unrestrained combustion and losses due to friction and heat transfer [18].

A significant portion of the available energy is lost due to irreversibilities, which

accounts for nearly 20-25% of the fuel’s exergy as shown in Figure 6.1. The destroyed

exergy appears as heat that cannot be transformed to useful work. Overcoming these

limits involves complex optimization of materials controls, thermodynamics and engine

architecture [186–189].

Figure 6.1: Thermodynamic comparisons of available fuel energy [186]

The composition of diesel fuel traditionally had modest influence on engine

efficiency compared to engine design, even though fuel composition had a significant

influence on emissions [190]. Recent research has shown that advanced combustion

engines, which hold promise for both improved fuel economy and low emissions with

142

modest or no aftertreatment systems, are sensitive to fuel properties [9–11]. Recognizing

this, the Fuels for Advanced Combustion Engines (FACE) working group was initiated,

whose objective was to recommend ‘designed fuels’ that will further the understanding of

fuel property impacts on advanced combustion processes, their efficiencies, and their

emissions [191]. The coordinating research council (CRC) fuels committee recognized

that the key fuel properties of immediate interest were the cetane number, aromatic

content, and T90 temperature, with the idea that fuels could be formulated in the future

which included other fuel properties. Fuels were carefully blended using non-traditional

blending streams and mixing pure components to maintain orthogonal (independent)

relationship between the selected fuel properties. Modifying any other fuel property was

not necessarily orthogonal and added to complexities in data analysis. In commercially

available fuels, these properties are not independent, as modifying any fuel property

invariably influences the other. A brief description of cetane number, aromatic content,

and volatility are summarized below.

Cetane Number

Cetane number of a fuel indicates how easily it can be ignited under the pressure

and temperature conditions in the combustion chamber of an engine. The higher the

cetane number, the shorter the ignition delay [192]. An increase in the paraffin content of

the fuel increases the cetane number and an increase in the aromatic content reduces the

cetane number [18]. Increasing the cetane number improves fuel combustion and

advances the start of combustion (phasing) [193]. The effect of cetane number on fuel

143

economy has been studied by many researchers in the past, and the general consensus is

that increasing the cetane number improves the fuel economy of the engine [194–196].

The effect of cetane number on NOx emissions seems to be dependent on the

combustion strategy. Under conventional diesel combustion, NOx emissions reduces with

increasing cetane number [8,197,198], while some report an increase in NOx emissions

under advanced combustion strategies [199,200]. Similarly, PM emissions are found to

be engine-dependent [197,201–203]. HC and CO emissions have been shown to decrease

with an increase in cetane number [197,198]. Increasing the cetane number produces the

greatest benefit when cold-starting with a relatively low cetane number fuel.

Aromatics

Building block for aromatic hydrocarbons is the benzene (C6H6) ring structure.

The ring structure is very stable and accommodates additional –CH2 groups in side chains

and not by ring expansion. More complex aromatic hydrocarbons incorporate ethyl,

propyl, and heavier alkyl side chains in a variety of structural arrangements [18]. For the

same number of carbon atoms, aromatics have higher boiling point, density, and

volumetric heating value compared to paraffins (CnH2n+2) and naphthenes (CnH2n).

Aromatics have cetane numbers ranging from 0-60. Aromatics with a single ring and a

long side chain have the maximum cetane number, while multiple ring aromatics have the

lowest cetane number. Hence altering the aromatic content would modify the ignition

quality of the fuel.

144

Fuel aromatic content does not seem to exert much influence on the fuel

economy and the overall combustion characteristics [9,204], however the reduction of

aromatic content has been found to reduce each of the regulated emissions [192,197,205–

207]. Studies on the influence of polynuclear aromatic hydrocarbons show that reducing

the di- and tri-aromatics in the fuel reduces emissions of HC, PM, and NOx [208]. Under

moderate to high loads, aromatic content plays a key role in determining the peak flame

temperature, which in turn affects NOx emissions

Volatility

Boiling point range is an important factor for controlling fuel quality. The

distillation range will affect other properties of the fuel such as density, viscosity, cetane

number, etc. T10 indicates the temperature at which 10% of the fuel will be evaporated

and it reflects the ease with which the fuel will start to vaporize. T90 indicates the

temperature at which 90% of the fuel vaporizes in the combustion chamber. In general,

for hydrocarbons of similar carbon number, volatility decreases (from higher to lower

volatility) in the order: branched alkanes > normal alkanes > cycloalkanes ≥ aromatics

[209]. Higher boiling point fuel components are often difficult to burn and contribute to

PM and HC emissions because of their low volatility. Studies of diesel fuel showed that a

reduction in final boiling point reduced PM emissions [210–213].

145

6.2 Objective

The objectives of this test were to examine the sensitivity of fuel conversion

efficiency to cetane number, aromatic content, and fuel’s volatility; and evaluate the

performance of a turbo-diesel engine while operating under a range of fuels referred to in

this study as the “Matrix 1” fuels. The fuels were comprised of a baseline ULSD, along

with six other diesel fuels varying in derived cetane number, total aromatic percentage

and distillation temperature, as shown in Table 6.1. These fuels are referred to here as

Fuel I through Fuel VI. The measured (and determined) variables include BTE, BSFC,

engine-out gaseous and PM emissions, exhaust particle size distribution, and heat release

profiles.

Figure 6.2: Layout of Matrix 1 fuels

146

Table 6.1: Fuel Matrix

Properties

Low target

High Target

VEV Program

High DCN High DCN High DCN High DCN Low DCN Low DCN

ExxonMobil Confidential Baseline

Diesel Low

Aromatic High

Aromatic High

Aromatic Low

Aromatic Low

Aromatic High

Aromatic

11-94923 Low T90 Low T90 High T90 High T90 Low T90 Low T90

Fuel Name Baseline

Diesel Fuel I Fuel II Fuel III Fuel IV Fuel V Fuel VI

D6890 DCN actual 30 55 45.1 55.4 49.2 52.9 56.9 36.4 31.3

Density (g/cm3) model 0.8318

0.7951 0.8356 0.831 0.7927

0.8036 0.8414

(actual) (actual) (actual)

D5186 Total Aromatics (wt%) actual

~11 ~44 31.54 12.7 46.6 40.4 11.3 15 48

(10 vol %) (40 vol %)

D5186 Poly Aromatics (wt%) actual

8.31 3.65 13.3 11.2 3.4 4.2 10.8

D86 T90 (oF) actual 554 626 593.8 569 575 616 619 558 560

Carbon content model 87.32%

86% 88% 88% 87% 87% 87% (actual)

Hydrogen content model 13.34%

14% 12% 12% 13% 13% 13% (actual)

Net heat of Combustion (BTU/lb) model

18,305

18,805 18,373 18,444 18,655 18,770 18,424 (actual)

147

6.3 Experimental

6.3.1 Engine

An 8 cylinder Ford 6.4L “Powerstroke” direct injection diesel engine coupled to

an eddy-current dynamometer was used for the experiments. The engine has a brake

power of 261 kW at 3000 rpm, and a peak torque of 881 Nm at 2000 rpm. Engine

operating parameters, especially the engine control unit (ECU) parameters, were

monitored using Inca v6.1 software and custom written Labview software. Engine

specifications are presented in Table 4.1

6.3.2 Particle Size and Distribution Measurements

Particle size distribution and concentration measurements were obtained with a

TSI 3936 Scanning Mobility Particle Sizer (SMPS). The SMPS instrument included a

TSI 3080 Electrostatic classifier with a Differential Mobility Analyzer (DMA) and a

3776 Condensation Particle Counter (CPC). A slipstream of the exhaust gas was drawn

into the SMPS using the BG3 sampling system at 1.5 lpm with a dilution of 10:1. The

sheath flow rate was maintained at 15 lpm resulting in a net dilution of 100:1 at the CPC.

Dilution ensured that the temperature and the net particle concentration of the exhaust gas

were in the measurable range of the SPMS. The particles were sampled in the range of

6.3 to 220 nm.

148

6.3.3 Test Conditions

The test was comprised of the following operating conditions.

Figure 6.3: Engine operating conditions

A baseline condition of 2000 rpm and 45% load was selected based on

recommendations for the class of this engine (provided by Jim Morris, retired, Volvo

Group Truck Technology), which was found to operate at 45% load a high percentage

of the time. This condition represented the engine operation under medium speed and

load, with an average BSFC value, typical of the driving conditions on a highway. A

NOx emissions target was set at 3.5 g/kWh, which was achieved by varying the

EGR% and modifying the injection timing on the ECU. The target value of 3.5

g/kWh of NOx emissions is in accordance with the emissions regulations for this

class of engine [214].

Three other conditions: 1400 rpm 20% load, 1400 rpm 60% load, and 2800 rpm 20%

load were selected to complete a map of the engine between low and high speeds and

149

low and high loads, respectively. The engine was unstable at 2800 rpm and 60% load

and the dynamometer could not be cooled during extended engine tests and hence that

condition was avoided. For these three engine conditions, the engine was operated

under default ECU calibration settings.

6.3.4 Notes on Experimental Conditions and Results

The fuels were not optimized under the baseline condition versus the other conditions

selected in this study.

The engine conditions and fuel runs were not randomized in the experimental outline.

No baseline run was performed at the end of each test to determine drift in the

experimental results gathered previously.

The error bars plotted in the figures represent statistical uncertainty of the data

collected over 5 minutes per operating condition. No statistical tools or optimization

subroutines have been used to represent the trends reported in this study.

The experimental procedures do introduce potential biases in the observed results.


6.4.1 Effect of Fuel Properties on BSFC and BTE

BSFC and BTE values at various engine conditions are plotted in Figure 6.4 and

Figure 6.5, respectively. The error bars for BTE have been plotted, but are very small and

150

cannot be seen on the plots. It can be observed that Fuel VI (low DCN, high aromatics,

low T90) resulted in the highest BSFC and lowest BTE, while Fuel IV (high DCN, low

aromatics, high T90) resulted in the lowest BSFC and highest BTE at all the engine

conditions studied. Fuel I (high DCN, low aromatics, low T90) also resulted in similar

BSFC/BTE compared to Fuel IV. Overall, it appears that a fuel with a high DCN and a

low aromatic content is suitable for high engine efficiency and low fuel consumption.

160

180

200

220

240

260

280

300

320

Baseline Fuel I Fuel II Fuel III Fuel IV Fuel V Fuel VI

BS

FC

, g

/kW

h

a)

150

160

170

180

190

200

210

220


BS

FC

, g

/kW

h

b)

150

160

170

180

190

200

210

220

230


BS

FC

, g

/kW

h

c)

200

220

240

260

280

300

320

340

360


BS

FC

, g

/kW

h

d)

Figure 6.4: Effect of fuel properties on BSFC at a) 1400 rpm, 20% load, b) 1400

rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20% load

151

25

26

27

28

29

30

31

32


BT

E, %

a)

35

36

37

38

39

40

41

42

43


BT

E, %

b)

30

32

34

36

38

40

42


BT

E, %

c)

20

21

22

23

24

25

26

27


BT

E, %

d)

Figure 6.5: Effect of fuel properties on BTE at a) 1400 rpm, 20% load, b) 1400 rpm,

60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20% load

A higher cetane number fuel (say Fuel I vs. Fuel V; Fuel II vs. Fuel VI) resulted

in improved BSFC and BTE, which is consistent with studies reported in the literature

[194,195]. This is because increasing the cetane number reduces the ignition delay period

and improves the combustion phasing. This is also confirmed from the apparent heat

release profiles plotted in Figure 6.12.

152

Increasing the fuel aromatic content (Fuel II vs. Fuel I; Fuel III vs. Fuel IV)

resulted in higher BSFC and lower BTE values, due to poor volatility of the heavier

compounds in the fuel which do not combust well. There aren’t too many studies in the

literature which report on the isolated effect of fuel aromatic content on specific fuel

consumption in a diesel engine. In one study, Bunting et al. [9] observed no significant

impact of aromatic content on fuel economy, however, the engine was operated under

HCCI conditions and not conventional diesel combustion. In another study, Kidoguchi et

al. [204] also observed no change in fuel economy with aromatic content, in a naturally

aspirated direct injection diesel engine, since fuels with different aromatic content

resulted in similar pressure rise and apparent heat release rates. However, in this work

(performed on a turbocharged engine), increasing the aromatic content increased the heat

release rate from both premixed and diffusion burn fractions, which could have resulted

in an increase in fuel economy. This shows that the engine efficiency is sensitive to fuel

formulation and combustion strategy.

The effect of T90 temperature on BSFC and BTE was insignificant. For example,

Fuel II and Fuel III varied in T90 temperature with other parameters almost identical, but

no significant changes were observed in BSFC and BTE. The apparent heat release

profiles were also similar with change in fuel volatility. The observed weak correlation

between T90 temperature and BTE/BSFC suggests that high T90 fuels may be used in

diesel engines without a significant compromise on the engine performance (later

paragraphs show that increasing T90 reduces HC/CO emissions with no change in NOx

emissions).

153

0.1

0.11

0.12

0.13

0.14

0.15

0.16


Vo

lum

e F

low

Ra

te,

L/m

in

a)

0.25

0.3

0.35


Vo

lum

e F

low

Ra

te, L

/min

b)

0.3

0.32

0.34

0.36

0.38

0.4


Vo

lum

e F

low

Ra

te,

L/m

in

c)

0.3

0.32

0.34

0.36

0.38

0.4


Vo

lum

e F

low

Ra

te,

L/m

in

d)

Figure 6.6: Effect of fuel properties on fuel consumption (L/min) at a) 1400 rpm,

20% load, b) 1400 rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20% load

From the fuel consumption rate (L/min) plots shown in Figure 6.6, it can be seen

that at 1400 rpm and 20% load, Fuel VI has the highest, while baseline ULSD and Fuel

IV have the lowest fuel consumption. At 1400 rpm and 60% load, Fuel V resulted in the

highest fuel consumption rate, while most other fuels had similar consumption rates. At

the baseline engine condition, Fuels I and V resulted in high fuel consumption, while

other fuels had similar rates. And at 2800 rpm and 20% load, Fuels I, II, and V had high

154

consumption rates, while other fuels had very similar values. The overall trends reported

in L/min are slightly different from BSFC (expressed in g/kWh) because of the variation

in fuel densities and heating values as seen in Table 6.1. Nevertheless, both the

comparisons are important for understanding the engine performance at various operating

conditions.

6.4.2 Effect of Fuel Properties on Emissions

6.4.2.1 NOx and PM Emissions

NOx emissions from each of the fuels are shown in Figure 6.7. NOx emissions at

2000 rpm and 45% load are maintained constant at 3.5 g/kWh by varying the ECU

parameters. At 1400 rpm and 20% load, Fuel VI resulted in the lowest NOx emissions,

while all other fuels had similar NOx emissions. At 1400 rpm and 60% load, baseline

fuel and Fuel VI resulted in the highest, while Fuel IV resulted in the lowest NOx

emissions. At 2800 rpm and 20% load, Fuel III had the highest NOx emissions while all

other fuels had similar NOx emissions within experimental uncertainties. In general, NOx

emissions are a maximum at 1) low speeds as there is more time for NOx formation and

2) high loads due to higher cylinder temperatures.

Increasing the cetane number had a mixed effect on NOx emissions. For most

engine conditions, changing the cetane number had no significant effect on NOx

emissions, but for some engine conditions, increasing the cetane number increased NOx

emissions. In general, increasing the cetane number reduces the premixed burn fraction

155

(Figure 6.12) which in turn reduces the peak pressure rise and localized gas temperature

leading to lower NOx emissions [215]. However, it has been observed that under low

loads, increasing the cetane number increases NOx emissions due to an increase in the

combustion duration [8]. The competition between the two processes of reduced ignition

delay and increased combustion duration leads to mixed NOx emissions trends as

observed earlier.

The effect of fuel aromatic content on NOx emissions was more pronounced at

low speeds. At low speed and low load, higher aromatic fuels resulted in slightly higher

NOx emissions than the lower aromatic fuels, due to higher flame temperatures resulting

from greater premixed burn fraction in the heat release as seen in Figure 6.12. This is

consistent with the findings in the literature [192,216]. However, the baseline fuel which

has an aromatic content in the mid-range of the high and low aromatic fuels results in

similar NOx emissions as that of the high aromatic fuels. At high speed and low load,

there is no clear distinction of the effect of fuel aromatic content on NOx emissions.

These differences may suggest that the effect of aromatic content may not present a

strong correlation on NOx emissions.

Modifying the T90 temperature did not have a significant impact on NOx

emissions at all the conditions tested in this study, as confirmed by the identical heat

release profiles in Figure 6.12.

156

0

0.2

0.4

0.6

0.8

1


BS

NO

x,

g/k

Wh

a)

0

1

2

3

4

5

6

7


BS

NO

x,

g/k

Wh

b)

0

0.5

1

1.5

2

2.5

3

3.5

4


BS

NO

x,

g/k

Wh

c)

0

0.5

1

1.5

2

2.5


BS

NO

x,

g/k

Wh

d)

Figure 6.7: Effect of fuel properties on BSNOx emissions at a) 1400 rpm, 20% load,

b) 1400 rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20% load

PM emissions are plotted in Figure 6.8. It can be observed that at 1400 rpm and

20% load, Fuel III resulted in the highest, while Fuel V resulted in the lowest engine-out

PM emissions. At 1400 rpm and 60% load, Fuel VI and Fuel IV resulted in highest and

lowest PM emissions respectively. At 2000 rpm and 20% load condition, Fuel VI resulted

in the highest, while Fuel I resulted in the lowest PM emissions. At 2800 rpm and 20%

load condition, Fuel I resulted in the lowest, and baseline fuel resulted in the highest PM

157

emissions. Fuel I (blend of high DCN, a low aromatic content and a low T90) was the

optimum combination for lowest PM emissions from this study.

Increasing the cetane number resulted in improved combustion and lowered

engine-out PM emissions, except at 1400 rpm and 20% load. From the literature, the

effect of increasing cetane number on PM emissions is not clear, with some reporting an

increase in PM emissions due to longer combustion duration [199,200], and others

reporting a decrease or no effect [197,201,217]. The mixed opinion is a result of

sensitivity of the fuel properties to different combustion strategies and classes of engines.

The effect of fuel aromatic content on PM emissions was not very clear. At low

speed and low load, fuels I and II resulted in similar engine-out PM emissions, while

fuels III and IV did show a significant difference. The effect of fuel aromatic content on

PM emissions was more pronounced at high engine loads, and a fuel with high aromatic

content resulting in high PM emissions. This is consistent with most findings in the

literature [192,197,205–207]. The increase in PM emissions is because the aromatics in

the fuel serve as precursors to the formation of soot and PAHs [218]. At the high speed

low load condition, there is a lot of variability in the observed data. The sensitivity of the

fuel at different engine conditions presents a challenge to isolate the effect of a single fuel

property on the measured variables.

PM emissions at different T90 temperatures seemed to be dependent on the

engine condition and the fuel aromatic content as well. For the engine conditions

operating on fuels with high aromatic content, the effect of T90 temperature was not very

significant, but when operating on fuels with a low aromatic content, a high T90 fuel

158

resulted in high PM emissions, due to poor volatility of the long-chain compounds, which

is consistent with the literature [210–212].

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6


BS

PM

, g

/kW

h

a)

0

0.05

0.1

0.15

0.2

0.25

0.3


BS

PM

, g

/kW

hb)

0

0.05

0.1

0.15

0.2


BS

PM

, g

/kW

h

c)

0

0.5

1

1.5


BS

PM

, g

/kW

h

d)

Figure 6.8: Effect of fuel properties on BSPM emissions at a) 1400 rpm, 20% load,


6.4.2.2 Hydrocarbon and CO Emissions

Brake specific hydrocarbon and CO emissions are plotted in Figure 6.9 and

Figure 6.10, respectively. It can be observed that Fuel VI had the highest HC and CO

159

emissions for all engine conditions. This could be a result of the low cetane number and

high aromatic content in the fuel. Fuel IV resulted in the least HC and CO emissions for

most engine conditions. HC and CO emissions are a direct consequence of incomplete

combustion of the fuel. Heavier boiling point compounds which have low volatility do

not completely burn leading to an increase in the HC emissions. In general, Fuels I and

IV resulted in the lowest HC and CO emissions. At some conditions, trends associated

with higher HC did not necessarily correlate to high CO emissions as observed from the

plots.

Increasing the cetane number (Fuel I vs. Fuel V; Fuel II vs. Fuel VI) resulted in

lower HC and CO emissions as expected and is consistent with the literature [197,198].

This trend is logical, given that highly ignitable fuels are more likely to burn completely.

Increasing the fuel aromatic content (Fuel II vs. Fuel I; Fuel III vs. Fuel IV)

resulted in higher HC and CO emissions. This could be due to a result of lowered ignition

quality of the fuel with increasing aromatic content and incomplete oxidation of the fuel

[11,219].

Increasing the T90 temperature (Fuel III vs. Fuel II; Fuel IV vs. Fuel I) resulted in

lower HC and CO emissions which is consistent with the results reported in the literature

[192]. This could have been due to higher flame temperatures resulting from increased

heat release from the heavy-end fraction of the diesel cut, leading to improved oxidation.

160

0

2

4

6

8

10

12

14


BS

HC

, g

/kW

h

a)

0

0.2

0.4

0.6

0.8

1

1.2

1.4


BS

HC

, g

/kW

h

b)

0

0.5

1

1.5

2


BS

HC

, g

/kW

h

c)

0

0.5

1

1.5

2


BS

HC

, g

/kW

h

d)

Figure 6.9: Effect of fuel properties on BSHC emissions at a) 1400 rpm, 20% load,


161

0

5

10

15

20


BS

CO

, g

/kW

h

a)

0

1

2

3

4

5

6

7

8


BS

CO

, g

/kW

h

b)

0

0.5

1

1.5


BS

CO

, g

/kW

h

c)

0

1

2

3

4

5

6


BS

CO

, g

/kW

h

d)

Figure 6.10: Effect of fuel properties on BSCO emissions at a) 1400 rpm, 20% load,


6.4.3 Effect of Fuel Properties on Particle Size and Distribution

Particles found in diluted diesel exhaust are found in three size modes, the

nucleation mode, Dp < 30 nm, the accumulation mode, 30 nm < Dp< ∼500 nm, and the

coarse mode, 500 nm < Dp < 10μm [220]. The precise boundaries between the three size

modes vary, but the nature of the particles in the three modes is quite different. The

nucleation mode is where freshly nucleated particles and droplets from condensation of

162

exhaust species are found. The accumulation mode is where most of the soot is found.

The coarse mode contains mainly soot particles that have been deposited on surfaces and

subsequently re-entrained. In this study, the particles were sampled in the range of 20-

220nm.

Particle size and distribution of the exhaust particles are shown in Figure 6.11. At

1400 rpm and 20% load, it was observed that the baseline condition resulted in the

highest concentration of particles while Fuel V and Fuel VI resulted in lower

concentrations with their distribution shifted to lower particle diameters. At 1400 rpm

and 60% load condition, Fuels IV, V, and VI resulted in a much higher particle

concentrations compared to all other fuels, while the rest of the fuels resulted in similar

distributions. At the baseline engine condition of 2000 rpm and 45% load, baseline fuel

and Fuel III had the highest and lowest particle concentrations, respectively. Fuel V and

Fuel VI exhibited a bimodal distribution showing the nucleation and accumulation phase

of the particles evidently, while all other fuels exhibited a unimodal distribution showing

only the accumulation mode particles. At 2800 rpm and 20% load, baseline fuel, Fuel IV

and Fuel V resulted in a similar highest particle concentration while Fuel III had the

lowest particle concentration. Fuel I which had the least engine-out PM emissions on a

mass basis appeared to be optimally centered in the particle distribution curve at all

engine conditions. The isolated effect of cetane number, aromatic content, and T90

temperature is not clear, as different trends are observed at different operating conditions.

However, it is clear that particle distribution is significantly influenced by the fuel

formulation and the engine operating conditions.

163

0

1 106

2 106

3 106

4 106

5 106

6 106

7 106

8 106

0 50 100 150 200 250

Baseline Fuel IFuel IIFuel IIIFuel IVFuel VFuel VI

Part

icle

Co

nce

ntr

ati

on

, #

/cm

3

Particle Diameter, nm

a)

0

2 105

4 105

6 105

8 105

1 106

0 50 100 150 200 250

BaselineFuel IFuel IIFuel IIIFuel IVFuel VFuel VI

Part

icle

Co

nce

ntr

ati

on

, #

/cm

3


b)

0

2 105

4 105

6 105

8 105

1 106

0 50 100 150 200 250


Part

icle

Co

nce

ntr

ati

on

, #

/cm

3


c)

0

1 106

2 106

3 106

4 106

5 106

6 106

0 50 100 150 200 250


Part

icle

Co

nce

ntr

ati

on

, #

/cm

3

Particle Diameter

d)

Figure 6.11: Effect of fuel properties on particle size and distribution at a) 1400

rpm, 20% load, b) 1400 rpm, 60% load, c) 2000 rpm, 45% load, d) 2800 rpm, 20%

load

6.4.4 Effect of Fuel Properties on Apparent Heat Release Rate

The apparent heat release profiles are plotted in Figure 6.12. For the engine

condition at 2800 rpm and 20% load, the heat release profile was not plotted as the data

obtained was very noisy. Because of the number of data points and the proximity of these

data points, it may be difficult to pick out trends from the graph. At 2000 rpm and 45%

164

load, the start of injection was fixed the same for all the fuels by programming the ECU.

At this condition, Fuel VI appeared to have a very high magnitude of premixed burn

fraction followed by Fuel V. Rest of the fuels resulted in similar burn rates. It is

important to mention that changing the fuel formulation impacts the fuel’s net heating

value and density, which influences the magnitude of the heat release.

0

50

100

150

350 360 370 380 390 400


AH

R, J/d

eg

Crank Angle, deg ATDC

a)

0

50

100

150

340 350 360 370 380 390 400


AH

R, J/d

eg


b)

0

50

100

150

200

250

300

350 360 370 380 390 400


AH

R, J/d

eg


c)

Figure 6.12: Effect of fuel properties on apparent heat release at a) 1400 rpm, 20%

load, b) 1400 rpm, 60% load, c) 2000 rpm, 45% load

165

In general, Fuels V and VI resulted in a delayed start of combustion compared to

all other fuels due to a low cetane number (higher cetane number results in shorter

ignition delay). A low cetane number fuel (Fuel VI vs. Fuel II; Fuel V vs. Fuel I) also

resulted in higher peak heat release rates under premixed burn. Reducing the fuel

aromatic content marginally reduced the heat release from both the premixed and

diffusion burn. Increasing the T90 temperature did not have any effect on the heat release

rates under the conditions studied.

6.5 Conclusions

It was observed that fuel conversion efficiency was sensitive to fuel formulation,

and this had a significant impact on engine’s performance and emissions. Three key fuel

properties: cetane number, aromatic content, and T90 distillation temperature were

studied in these experiments. Of the three properties, cetane number and aromatic content

influenced the engine performance more than the distillation temperature. Due to lack of

statistical data, it was difficult to capture the variation in some of the trends and isolate

conclusively the effect of fuel properties on the measured variable. Nevertheless, the

effect of each fuel property on the engine’s performance observed in this study is

summarized in Table 6.2. Based on the total number of plusses and minuses for each fuel

tested in this study, the following overall conclusions are drawn.

A fuel with a high DCN and low aromatic content results in high engine efficiency,

low fuel consumption, and low NOx and PM emissions.

166

A fuel with a low DCN, a high aromatic content, and a low T90 appears to be the

least desirable combination overall for engine performance.

The best fuel appeared to be Fuel IV, which was a combination of a high DCN, a low

aromatic content and a high T90.

Table 6.2: Summary of the effect of fuel properties on engine performance

Fuel Property BSFC BTE NOx PM HC CO

DCN (+) - + +, - - - -

Aromatic Content (+) + - + +, - + +

T90 (+) ~ ~ ~ + - -

+ indicates increase, - indicates decrease, ~ indicates no change

167

Chapter 7

Conclusions and Suggestions for Future Work

7.1 Summary

‘Clean Diesel’ technology is now the standard for all new U.S. vehicles from

passenger cars to highway commercial trucks. Since its introduction in 2007, many

technological advances have been made to address the system’s three key parts: clean

diesel fuel, advanced engine combustion strategies, and aftertreatment. Limiting on-road

fuel sulfur content has been advantageous to improve the performance of the

aftertreatment systems to lower engine-out emissions. Even though much progress had

been made to achieve high engine efficiency with low engine-out emissions, several

technical challenges exist that continue to need to be addressed. This dissertation aimed

to address the following challenges: 1) Concerns with EGR cooler fouling, since current

diesel engines depend on EGR to reduce NOx emissions in-cylinder, 2) achieving high

efficiency via dual fuel combustion by fumigating DME and propane into diesel engine

air intake, and 3) exploring optimization of fuel composition to maximize engine

efficiency and performance. Conclusions for each component of this study are provided

below.

7.1.1 Conclusions from EGR Cooler Fouling Study

Engine conditions played a significant role in dynamically changing the deposit

properties in the EGR cooler. High engine loads resulted in greater deposit

168

accumulation in the EGR cooler and the deposits were comprised of mostly soot

particles which had a coarse microstructure. On the contrary, low engine load resulted

in lower deposit accumulation, and the deposits were comprised of a high percentage

of hydrocarbons.

EGR cooler effectiveness dropped rapidly initially but stabilized (plateaued) after

several hours of exposure to exhaust gas.

EGR cooler deposits were mainly comprised of the heavy aliphatics (C18-C25),

which typically arise from the long chain compounds in diesel fuel and lubricating

oil, suggesting that lowering the T90 distillation temperature would reduce fouling in

EGR coolers.

Coolant temperature altered the nature of the deposits and forced a greater percentage

of hydrocarbons to condense on the surface of the EGR cooler.

Low coolant temperatures promoted higher EGR cooler recovery during engine shut

down and start-up conditions, confirming the role of condensed water vapor on

deposit removal.

Under the conditions studied, the oxidation catalyst did not seem to be very effective

in reducing EGR cooler fouling due to low gas temperatures across the catalyst

resulting in poor conversion efficiency.

7.1.2 Conclusions from High Efficiency Combustion Study

Dual fuel combustion was demonstrated with fumigation of DME and propane into

diesel engine air intake, where propane slowed the autoignition of DME due to

169

apparent chemical kinetic interaction between propane and DME molecules in the

fumigated intake charge.

DME combustion exhibited the typical low temperature and high temperature heat

release events with very early onset of reaction, due to low autoignition temperature

of DME. Introduction of propane delayed the onset of DME combustion by shifting

the combustion process closer to top dead center, coinciding better with the timing of

combustion of the direct injected diesel fuel.

A maximum of 49% BTE was observed for 20% DME and 30% propane substitution

(on an energy basis) compared to a baseline efficiency of 37%, and resulted in only

about 6-7 bar increase in the peak cylinder pressure. This condition had a 24% lower

brake specific energy consumption compared to the baseline diesel engine condition.

Total hydrocarbon emissions increased with increasing propane substitution and

marginally decreased with DME substitution. The increase with propane was more

substantial owing to a higher autoignition temperature.

NOx emissions reduced with increasing DME and propane substitution due to an

increase in the proportion of homogenous oxidation (similar to the premixed burn in

conventional diesel combustion) which reduces the amount of heat release during the

mixing controlled burning of diesel fuel.

CO emissions increased with increasing propane and DME until 20% DME. CO2

emissions appeared to decrease mainly with increasing propane while it remained

more or less constant with increasing DME substitution.

170

7.1.3 Conclusions from Fuel Impacts on Engine Performance

Fuel conversion efficiency was sensitive to fuel formulation and this impacted the

engine’s performance and emissions.

Increasing the fuel cetane number improved BSFC and BTE. There was no definite

trend in NOx emissions; however all other emissions reduced.

Increasing the fuel aromatic content reduced BTE and increased BSFC and engine-

out emissions for most conditions.

Increasing the T90 temperature did not have a significant effect on BTE, BSFC, and

NOx emissions. HC and CO emissions decreased while PM emissions increased.

A fuel with a high DCN, a low aromatic content and a high T90 temperature was

more suitable for high engine efficiency, low BSFC, NOx and PM emissions.

A fuel with a low DCN, a high aromatic content, and a low T90 temperature appeared

to be a less desirable combination overall for engine performance.

7.2 Suggestions for Future Work

7.2.1 Suggestions for Future Work on EGR Cooler Fouling

Evaluate the role of a low T90 fuel on EGR cooler deposit properties

The deposits from the EGR cooler were mostly comprised of heavy aliphatics

(C18-C25), which typically arise from incomplete combustion of the long chain (back-

end) compounds present in the fuel and the lubricating oil. It is hypothesized that

171

lowering the T90 distillation temperature of the diesel fuel will reduce the deposition of

the heavy aliphatic compounds in the EGR cooler. As a follow-up experiment, the lighter

hydrocarbons can be reduced using an oxidation catalyst to see the net effect on fouling.

Evaluate the role of steam injection into EGR coolers for effectiveness recovery

From the experimental results, it was evident that water vapor condensation

played a major role in the thermal effectiveness recovery of the EGR cooler during

engine shut down and start-up conditions. It is hypothesized that steam injection at

constant time intervals can keep the EGR cooler free of deposits.

Evaluate the role of thermal shocking of deposits to crack the deposit layer

Thermal stresses on the deposit layer can lead to cracks in the deposit layer.

However, the temperature under which this happens is still not very clear. Such an

experiment may be useful to understand a potential EGR cooler recovery mechanism.

7.2.2 Suggestions for Future Work on Dual Fuel High Efficiency Combustion

Explore dual fuel combustion experiments at other engine speeds and loads

The results reported in this work are performed at a single speed and load

condition, and may not be representative of the entire engine drive cycle. Hence, it is

necessary to perform similar experiments at other engine speeds and loads. Some other

key variables to experiment would be EGR and multiple injections, as against no EGR

and single injection investigated in this study.

172

Develop chemical kinetic mechanisms involved in DME and propane combustion

Some of the trends reported in this study (heat release rate, combustion phasing,

etc.) need further explanation, which can be provided by performing chemical kinetic

study of DME and LPG combustion using commercially available software like

ChemKin and using the chemical mechanisms provided by Livermore National Lab.

Particle mass and number concentration measurements

It is well known that soot is formed in fuel-rich regions under high temperature

conditions. The precursors of soot are unsaturated hydrocarbons such as acetylene

(C2H2), ethylene (C2H4), etc. Since DME has oxygen in the fuel, there will be an

influence on the soot precursors and the net soot emissions itself. There has been no

literature on the effect of DME and propane blends on particle mass, size, and number

distribution. The results presented in this study do not include particle mass and number

concentration (and distribution) measurements. Such a study will be useful in evaluating

the possibilities of commercializing this dual fuel, mixed combustion process, while

adhering to PM emissions regulations.

173

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193

Appendix A

Fuel Specifications

Table A.1: ChevronPhillips Ultra-low sulfur diesel fuel

Property Test Method Specification Value Unit

Specific Gravity ASTM D-4052 0.8400-0.8550 0.8466

API Gravity ASTM D-4052 34.0-37.0 35.6

Particulate Matter ASTM D-6217 <=15.0 1.1 mg/l

Cloud Point ASTM D-2500 2 FAH

Flash Point, PM ASTM D-93 >=130 155 FAH

Pour Point ASTM D-97 -5 FAH

Sulfur ASTM D-5453 7.0-15.0 9.7 ppm

Viscosity @40 C ASTM D-445 2.0-3.0 2.5 cSt

Hydrogen ASTM D-3343 13.2 WT%

Carbon Calculated 86.8 WT%

Poly Nuclear Aromatics ASTM D-5186 9.0 WT%

SFC Aromatics ASTM D-5186 30.0 WT%

Heat of Comb ASTM D-3338 18444 BTU/LB

Cetane Number ASTM D-613 43-47 45

HFFR Lubricity ASTM D-6079 <=0.4 0.3 mm

Distillation- IBP ASTM D-86 340-400 364 FAH

Distillation- 10 % ASTM D-86 400-460 413 FAH



Aromatics ASTM D-1319 28.0-32.0 28.8 LV %

Olefins ASTM D-1319 3.4 LV %

Saturates ASTM D-1319 67.8 LV %

194

Appendix B

Preliminary Calculations, Repeatability Studies, and Error Bars in Measurements

B.1 Coolant and Exhaust Gas Specific Heat Capacity

Coolant side calculations

Specific heat Cp,c = 3620 J/kg-K

Volume flow rate of the coolant = 6.8 lpm

Density of coolant = 971.9 kg/m3

Mass flow rate of coolant = volume flow rate * density = 6.6 kg/min

Heat capacity of the coolant CC = mass flow rate * Cp,c = 398.2 W/K

Exhaust side calculations

Specific heat Cp,e = 1067 J/kg-K

Volume flow rate of the exhaust gas = 180 lpm

Density of exhaust gas = 1.2 kg/m3 at standard conditions

Mass flow rate of exhaust gas = volume flow rate * density = 0.22 kg/min

Heat capacity of the coolant CH = mass flow rate * Cp,e = 3.91 W/K

Since CH << CC, coolant side temperature gain is a minimal

195

B.2 Repeatability Study

Experiments were performed to check the repeatability of the measurements

reported in this dissertation. Experimental conditions were maintained identical between

the two runs performed on different days. The tests are denoted as Test A and Test B.

Temperature, effectiveness and mass gain profiles are shown in Figures B.1-B.3

respectively. As observed, the data is very repeatable, within experimental uncertainties.

100

150

200

250

300

350

400

0 100 200 300 400 500 600

Data 1

Exhaust Temp, deg C (Test A)EGR Inlet Temp, deg C (Test A)EGR Outlet Temp, deg C (Test A)Exhaust Temp, deg C (Test B)EGR Inlet Temp, deg C (Test B)EGR Outlet Temp, deg C (Test B)

Tem

pe

ratu

re,

°C

Time, min

Figure B.1: Temperature profiles for Test A and Test B

196

0.4

0.45

0.5

0.55

0.6

0.65

0.7

0 100 200 300 400 500 600

Effectiveness (Test A)Effectiveness (Test B)

Eff

ecti

ven

es

s

Time, min

Figure B.2: Effectiveness change for Test A and Test B

20

30

40

50

60

70

80

90

0 2 4 6 8 10

Average, mg (Test A)Average, mg (Test B)

Av

era

ge D

ep

os

it M

as

s,

mg

Time, hours

Figure B.3: Average deposit mass gain for Test A and Test B

197

To understand what change tube removal does to temperature and effectiveness,

experiments were performed with and without tube removal at the same engine operating

condition. From Figure B.4, it can be observed that the temperature profiles remained

similar, except for the drop in temperature every 1.5 hours when the tubes were removed.

This is also reflected in the effectiveness change every 1.5 hours. From Figure B.5, it can

be observed that the net effectiveness change with and without tube removal is around 5-

6%. Hence, in our earlier experiments at high coolant temperature, it can be concluded

that removal of the tubes was the main reason for effectiveness improvement every 1.5

hours.

120

140

160

180

200

220

240

260

0 100 200 300 400 500 600

Exhaust Temp, Tubes RemovedEGR Inlet Temp, Tubes RemovedEGR Outlet Temp, Tubes RemovedExhaust Temp, No Tubes RemovedEGR Inlet Temp, No Tubes RemovedEGR Outlet Temp, No Tubes Removed

Tem

pe

ratu

re,

°C

Time, min

Figure B.4: Temperature change with and without tube removal

198

0.45

0.5

0.55

0.6

0.65

0 100 200 300 400 500 600

With tubes removalWithout tubes removal

Eff

ecti

ven

es

s, %

Time, min

Figure B.4: Effectiveness change with and without tube removal

B.2 Deposit mass variation across the 6 tubes

Special consideration was taken during the design phase of the EGR cooler to

ensure that the exhaust flow rate through each of the 6 tubes were identical. This was

achieved by carefully positioning the 5 peripheral tubes equidistant from each other and

the central tube. Additionally, the manifold cross-sectional area at the entry was greater

than twice the sum of the tubes’ cross-sectional areas, which is a common design

principle to ensure uniform flow distribution. Post design, a benchmarking study was

performed on the EGR cooler to ensure this. Even though the flow rate through each tube

was not measured, it was assumed that the uniformity of the flow was well represented by

the mass of the deposits collected in each of the tubes after a 9 hour testing interval as

shown in Figures B.5 and B.6. From these figures, it can be observed that the mass

199

distribution across the 6 tubes was uniform (within experimental uncertainties) and can

assume that the flow is also uniform.

0

10

20

30

40

50

60

70

80

Tube 1 Tube 2 Tube 3 Tube 4 Tube 5 Tube 6

De

po

sit

Ma

ss

Ga

in, m

g

Figure B.5: Benchmark study of the mass accumulation across 6 tubes of the EGR

cooler

0

10

20

30

40

50

60

70

80

Tube 1 Tube 2 Tube 3 Tube 4 Tube 5 Tube 6

De

po

sit

Ma

ss

Ga

in, m

g

Figure B.5: Mass accumulation across 6 tubes of the EGR cooler from a randomly-

chosen experiment

200

B.3 Error bars in measurements

Error bars are a graphical representation of the variability of data and are used on

graphs to indicate the error, or uncertainty in a reported measurement. They give a

general idea of how accurate a measurement is, or conversely, how far from the reported

value the true (error free) value might be. Error bars often represent one standard

deviation of uncertainty, one standard error, or a certain confidence interval (e.g., a 95%

interval). These quantities are not the same and so the measure selected should be stated

explicitly in the graph or supporting text. Another name for error bars is Confidence

Interval. The following error bars are adopted for the values reported in this dissertation.

The data reported in this dissertation follows partly the error bar reporting used by Moffat

[221].

Table B.1: Experimental uncertainties and error bars

Variable Sources of error Actual error bar

Deposit mass Weighing scale Standard deviation of 3

times mass measurements

Engine speed and

load

Load cell, speed sensor, digalog

calibrations

Speed and load errors are

not significant

Gaseous emissions

Span bottle concentrations, AVL

bench calibrations, fluctuations in

engine operating condition

Based on student t-test

with 50 sampling points

PM measurements

BG 3 calibration, fuel scale error,

fluctuations in engine operating

condition

Based on the standard

deviation of 3 filters

measured per condition

Py-GC MS Heterogeneous deposits, amount of

deposit placed in the quartz tube

5% standard error

reported in the literature,

201

actual error is even lower

CHN analysis Heterogeneous deposits, amount of

deposits

Standard deviation with

student t-test, 2 repeat

measurements

TGA Heterogeneous deposits, amount of

deposits <1% error

Gas flow rate Uncertainty in the reading due to

meniscus and calibration <5%

202

Appendix C

Calibration of High Temperature Flowmeter and Matheson Flowmeter

C.1 Wedge flowmeter calibration

A wedge flowmeter was used to measure the flow of the exhaust gas at varying pressures

and temperatures. The equations below summarize the calculations for wedge flow based

on Miller [222].

√

√(

)

Where,

Cd – coefficient of discharge

Hw – pressure drop, inches of water

ρf1 – upstream density at flowing conditions, lb/ft3, ρb – density at base conditions, lb/ft

3

d – flowing bore diameter, in

D – meter diameter, in

Y1 – Gas expansion factor based on upstream pressure

Nvρ – N factor for flowing volume with density determination

203

C.2 Matheson Flowmeter Calibration Charts:

Matheson flowmeter series 605, 604, 603 were calibrated to the flowing gas at 0

psig. Pressure and temperature correction factors were then added to calculate the flow

rate at actual flowing conditions. The calibrations are shown in Figures C.1 through C.4.

Figure C.1: Calibration for Flowmeter tube 605 at 0 psig for Propane

Figure C.2: Calibration for Flowmeter tube 603 at 0 psig for Propane

y = 0.0003x2 + 0.0817x + 0.0412 R² = 0.9998

0

5

10

15

20

25

0.0 50.0 100.0 150.0 200.0

Flo

w R

ate

(sl

pm

)

Scale Reading

y = -4E-05x2 + 0.0209x - 0.0398 R² = 0.9997

0

0.5

1

1.5

2

2.5

0 50 100 150 200

Flo

w R

ate

(sl

pm

)

Scale Reading

204

Figure C.3: Calibration for Flowmeter tube 604 at 0 psig for DME

Figure C.4: Calibration for Flowmeter tube 605 at 0 psig for DME

y = 1E-05x2 + 0.0506x + 0.0347 R² = 1

0.00

2.00

4.00

6.00

8.00

10.00

0.00 50.00 100.00 150.00 200.00

Flo

w R

ate

(sl

pm

)

Scale Reading

y = 0.0003x2 + 0.0958x - 0.0111 R² = 0.9999

0

5

10

15

20

25

0 50 100 150 200

Flo

w R

ate

(sl

pm

)

Scale Reading

205

Appendix D

Calculation of Heat Release Profiles from Pressure Traces

The apparent heat release rate was calculated based on the first law of thermodynamics

from measured cylinder pressure trace averaged over 200 cycles and over 2 such

measurements. The equation used to calculate is shown below.

(

)

(

)

Where,

Q = Net heat release, J

γ = Specific heat ratio

P = Pressure, kPa

V = volume, m3

θ = Degree of crank angle

VITA

BHASKAR PRABHAKAR

Ph.D., Energy and Mineral Engineering

The Pennsylvania State University, University Park, PA

Dissertation: Examination of EGR Cooler Fouling and Engine

Efficiency Improvement in Compression Ignition Engines

08/2009 – 05/2013

M.S, Mechanical Engineering

The Pennsylvania State University, University Park, PA

Thesis: Effect of Common Rail Pressure on BSFC versus BSPM

at NOx Parity

08/2007 – 08/2009

B.Tech, Mechanical Engineering

The National Institute of Technology, Tiruchirappalli, India

06/2003 – 05/2007

examination of egr cooler fouling and engine …

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