THE EFFECTS OF COMBUSTION CHAMBER DESIGN ON
TURBULENCE, CYCLIC VARIATION AND PERFORMANCE IN AN SI
ENGINE
By
Esther Claire Tippett
B.E.Mech (Hons) University of Canterbury, New Zealand. 1983
A THESIS SUBMITTED IN PARTIAL FULFILLMENT OF
THE REQUIREMENTS FOR THE DEGREE OF
MASTER OF APPLIED SCIENCE
in
THE FACULTY OF GRADUATE STUDIES
DEPARTMENT OF MECHANICAL ENGINEERING
We accept this thesis as conforming
to the required standard
THE UNIVERSITY OF BRITISH COLUMBIA
August, 1989
© Esther Claire Tippett, 1989 In presenting this thesis in partial fulfilment of the requirements for an advanced degree at the University of British Columbia, I agree that the Library shall make it freely available for reference and study. I further agree that permission for extensive copying of this thesis for scholarly purposes may be granted by the head of my department or by his or her representatives. It is understood that copying or publication of this thesis for financial gain shall not be allowed without my written permission.
Department of Mechanical Engineering
The University of British Columbia
Vancouver, Canada
Date: ABSTRACT
An experimental program of motored and fired tests has been undertaken on a single cylinder spark ignition engine to determine the influence of combustion chamber design on turbulence enhancement in the achievement of fast lean operation.
Flow field measurements were taken using hot wire anemometry in the cylinder during motored operation. On line performance tests and in-cylinder pressure data were recorded for the operation of the engine by natural gas at lean and stoichiometric conditions over a range of speed and loads.
Squish and squish jet action methods of turbulence enhancement were investigated for six configurations, using a standard bathtub cylinder head and new piston designs incorporating directed jets through a raised wall, a standard bowl-in-piston chamber and an original squish jet design piston. A non squish comparison was provided by a disc chamber.
Peak Pressure and Indicated Mean Effective Pressure (IMEP), two parameters char• acterizing performance and cyclic variability, showed that enhanced turbulence by com• bustion chamber geometry is effective in improving performance at lean operation. The single jet action directed towards the spark was most effective in improving the efficiency at high speed and lean mixtures. The addition of jets to the single jet, or jet chan• nels to the main squish action of the bowl- in-piston chamber, reduced performance and increased cyclic variability.
Mass fraction burn analysis of the cylinder pressure data showed that squish action was most effective in the main burn period. Configurations with large squish area and centrally located spark produced the greatest reduction in both the initial and main burn
ii periods.
The potential for the squish jet action to improve engine drivability and increase the knock limit was exhibited in reduced coefficient of variance of IMEP and reduced ignition advance requirements. Directions for further research to exploit this potential for engines operated by alternative fuels are identified.
iii Table of Contents
ABSTRACT ii
List of Tables viii
List of Figures xii
Nomenclature xxii
Acknowledgments xxvi
1 Introduction 1
1.1 Introduction and Background 1
1.2 Objective of this study 4
1.3 Turbulent Flow Field in an Engine 4
1.4 Discussion of Terminology in Turbulence Studies 6
1.5 Scope of Work 8
1.6 Structure of thesis 10
2 Literature Review 11
2.1 Introduction 11
2.2 Turbulence studies in Engines 12
2.3 Combustion Studies in Engines 16
2.4 Combustion Chamber Design 24
3 Experimental Apparatus and Method 27
iv 3.1 Introduction 27
3.2 Experimental Apparatus 28
3.2.1 Introduction . . . 28
3.2.2 Engine Bed . 29
3.2.3 Combustion Chambers 30
3.2.4 Modified bathtub pistons 32
3.2.5 Instrumentation 34
3.2.6 Data Acquisition 36
3.3 Motored Engine Tests 38
3.3.1 Introduction 38
3.3.2 Operational Procedures 39
3.3.3 Pressure Measurements 41
3.3.4 Hotwire Measurements 42
3.4 Fired Engine Tests 44
3.4.1 Introduction 44
3.4.2 Operational Procedures 45
3.4.3 Performance Measurements 46
3.4.4 Pressure Measurements 46
4 Data Analysis 48
4.1 Introduction 48
4.2 Motored Engine Tests 48
4.2.1 Analytical Procedure 48
4.2.2 Pressure Signal Processing 49
4.2.3 Anemometer Signal Processing 53
4.2.4 Flow Field Data Analysis 55
v 4.3 Fired Engine Tests 58
4.3.1 Analytical Procedure 58
4.3.2 Pressure Signal Processing 59
4.3.3 Performance Data Analysis 59
4.3.4 Combustion Analysis 62
5 Experimental Results and Discussion 66
5.1 Introduction 66
5.2 Motored Tests 66
5.2.1 Motored Pressure Results 66
5.2.2 Flow Field Results 68
5.3 Fired Tests 72
5.3.1 General Performance Parameters 72
5.3.2 Fired Pressure Results 75
5.3.3 Mass Fraction Burned Results 80
5.4 Experimental Uncertainties and Technique 86
5.4.1 Flow Measurement 86
5.4.2 Performance Measurements 88
5.5 Turbulence, Combustion and Performance 91
6 Conclusions and Recommendations 94
6.1 Introduction 94
6.2 Conclusions 95
6.2.1 Turbulence Studies 95
6.2.2 Performance and Combustion Studies 96
6.3 Recommendations 97
vi Bibliography 99
Appendices 190
A Instrument Specification and Calibration 190
B Hot Wire Anemometry Specification and Calibration 198
C BC Natural Gas Properties 202
D Pressure Filtering Methods 206
vii List of Tables
3.1 Ricardo Hydra Gasoline (or Gaseous fuel) Engine Specifications 104 3.2 Motored operating conditions for Pressure and Hotwire measurements at WOT 105 3.3 Fired operating conditions for Pressure measurements at MBT and Full Load (WOT) 105 3.4 Fired operating conditions for Pressure measurements at MBT and Part Load 105 4.1 Motored data Analysis program flow chart 106 4.2 Fired data Analysis program flow charts 107 5.1 Compression and Expansion coefficients for the motored condition. . . . 108 5.2 Engine performance as per SAEJ1349 for different piston geometries for stoichiometric and lean RAFR at MBT, WOT, and 20.0 rps 109 5.3 Engine performance as per SAEJ1349 for different piston geometries for stoichiometric and lean RAFR at MBT, WOT, and 33.3 rps 110 5.4 Engine performance as per SAEJ1349 for different piston geometries for stoichiometric and lean RAFR at MBT, WOT, and 50.0 rps 110 5.5 Engine performance as per SAEJ1349 for different piston geometries for stoichiometric and lean RAFR at MBT, 2.5 bmep, and 33.3 rps Ill 5.6 Engine performance as per SAEJ1349 for different piston geometries for stoichiometric and lean RAFR at MBT, 3.5 bmep, and 50.0 rps Ill
viii 5.7 Ignition Advance and Brake Thermal Efficiency for different piston geome•
tries for RARFwl.00-1.35 at MBT, WOT, and 20.0 rps 112
5.8 Ignition Advance and Brake Thermal Efficiency for different piston geome•
tries for RARF «1.00-1.35 at MBT, WOT, and 33.3 rps 113
5.9 Ignition Advance and Brake Thermal Efficiency for different piston geome•
tries for RARF «1.00-1.35 at MBT, WOT, and 50.0 rps 113
5.10 Imep, peak pressure and angle of occurance of peak pressure for different
piston geometries for stoichiometric and lean RAFR at, MBT, WOT, and
20.0 rps 114
5.11 Imep, peak pressure and angle of occurance of peak pressure for different
piston geometries for stoichiometric and lean RAFR at, MBT, WOT, and
33.3 rps 115
5.12 Imep, peak pressure and angle of occurance of peak pressure for different
piston geometries for stoichiometric and lean operation at, MBT, WOT,
and 50.0 rps 116
5.13 Imep, peak pressure and angle of occurance of peak pressure for different
piston geometries for stoichiometric and lean RAFR at, MBT, 2.5 bmep,
and 33.3 rps 117
5.14 Imep, peak pressure and angle of occurance of peak pressure for different
piston geometries for stoichiometric and lean RAFR at, MBT, 3.5 bmep,
and 50.0 rps 118
5.15 Initial (0-01% massburned) and Main (01-90% massburned) combustion
durations for different piston geometries for stoichiometric and lean RAFR
at MBT, WOT, and 20.0 rps 119
IX 5.16 Initial (0-05% massburned) and Main (05-90% massburned) combustion
duration for different piston geometries for stoichiometric and lean RAFR
at MBT, WOT, and 20.0 rps 120
5.17 Initial (0-01% massburned) and Main (01-90% massburned) combustion
duration for different piston geometries for stoichiometric and lean RAFR
at MBT, WOT, and 33.3 rps 121
5.18 Initial (0-05% massburned) and Main (05-90% massburned) combustion
duration for different piston geometries for stoichiometric and lean RAFR
at MBT, WOT, and 33.3 rps 122
5.19 Initial (0-01% massburned) and Main (01-90% massburned) combustion
duration for different piston geometries for stoichiometric and lean RAFR
at MBT, WOT, and 50.0 rps 123
5.20 Initial (0-05% massburned) and Main (05-90% massburned) combustion
duration for different piston geometries for stoichiometric and lean RAFR
at MBT, WOT, and 50.0 rps 124
5.21 Initial (0-01% massburned) and Main (01-90% massburned) combustion
duration for different piston geometries for stoichiometric and lean RAFR
at MBT, 2.5 bmep, and 33.3 rps 125
5.22 Initial (0-05% massburned) and Main (05-90% massburned) combustion
duration for different piston geometries for stoichiometric and lean at
RAFR MBT, 2.5 bmep, and 33.3 rps 126
5.23 Initial (0-01% massburned) and Main (01-90% massburned) combustion
duration for different piston geometries for stoichiometric and lean RAFR
at MBT, 3.5 bmep, 50.0 rps 127
x 5.24 Initial (0-05% massburned) and Main (05-90% massburned) combustion
duration for different piston geometries for stoichiometric and lean RAFR
at MBT, 3.5 bmep, and 50.0 rps 127
A.l Pressure transducer specifications for Kistler model 6121 191
A. 2 Pressure transducer calibration data for Kistler model 6121. No. 317205 . 192
B. l Hot wire anemometry equipment and specifications 199
B. 2 Hotwire calibration data for wire No. 1, yielding the calibration constants:
A=0.1575; £=0.4908; and n=0.360 200
C. l Composition of BC Natural Gas 202
C.2 Molecular weight of BC Natural gas 203
C.3 Higher and Lower Heating values of BC Natural gas 203
C.4 Viscosity calculations for BC Natural gas 204
xi List of Figures
3.1 Ricardo Hydra engine, dynamometer and control systems layout 128
3.2 Ricardo Hydra MKIII Gasoline (or gaseous fuel) Engine cross-sectional
and longitudinal views 129
3.3 Standard bathtub combustion chamber geometry 130
3.4 Single slot and castellated piston geometries 131
3.5 Bowl-in-piston and squish jet piston geometries 132
3.6 Instrumentation layout for the Ricardo engine test cell 133
3.7 Hot wire probe location through the spark plug entry for the bathtub and
flat cylinder heads 134
3.8 Acquisition hardware arrangement with fast pressure data acquisition hook
up 135
4.1 Comparison of the effect of window size on the turbulent intensity profile
for the bathtub chamber at WOT, 33.3 rps 136
4.2 Polytropic coefficent calculated from the ensembled pressure for single slot
chamber at MBT, WOT and 33.3 rps for RAFR=1.27 137
5.1 Motored pressure profiles for different chamber geometries at WOT, 20.0 rps. 138
5.2 Motored pressure profiles for different chamber geometries at WOT, 33.3 rps. 138
5.3 Motored pressure profiles for different chamber geometries at WOT, 50.0 rps. 139
5.4 Motored pressure profiles for different chamber geometries at WOT, 66.7 rps. 139
5.5 Motored pressure profiles for the single slot piston at WOT, for three
speeds; 20.0, 33.3, and 50.0 rps 140
xii 5.6 Motored temperature profiles for different chamber geometries at WOT,
33.3 rps 140
5.7 Window ensembled and cycle ensembled mean velocity and turbulent in•
tensity profiles for the single slot piston at WOT, 33.3 rps 141
5.8 Mean velocity profiles for different chamber geometries at WOT, 33.3 rps. 142
5.9 Turbulent intensity profiles for different chamber geometries at WOT,
33.3 rps • 142
5.10 Mean velocity profiles for the 'bathtub' group of chambers at WOT, 20.0 rps. 143
5.11 Turbulent intensity profiles for the 'bathtub' group of chambers at WOT,
20.0 rps 143
5.12 Mean velocity profiles for the 'bathtub' group of chambers at WOT, 33.3 rps. 144
5.13 Turbulent intensity profiles for the 'bathtub' group of chambers at WOT,
33.3 rps 144
5.14 Mean velocity profiles for the 'disc' group of chambers at WOT, 20.0 rps. 145
5.15 Turbulent intensity profiles for the 'disc' group of chambers at WOT,
20.0 rps 145
5.16 Mean velocity profiles for the 'disc' group of chambers at WOT, 33.3 rps. 146
5.17 Turbulent intensity profiles for the 'disc' group of chambers at WOT,
33.3 rps 146
5.18 Mean velocity profiles for the bathtub chamber at WOT, for three speeds;
20.0, 33.3 and 66.7 rps 147
5.19 Turbulent intensity profiles for the bathtub chamber at WOT, for three
speeds; 20.0, 33.3 and 66.7 rps 147
5.20 Mean velocity profiles for the castellated chamber at WOT, for three
speeds; 20.0, 33.3 and 50.0 rps 148
xiii 5.21 Turbulent intensity profiles for the castellated chamber at WOT, for three
speeds; 20.0, 33.3 and 50.0 rps 148
5.22 Mean velocity profiles for the squish jet chamber at WOT, for three speeds;
20.0, 33.3 and 50.0 rps 149
5.23 Turbulent intensity profiles for the squish jet chamber at WOT, for three
speeds; 20.0, 33.3 and 50.0 rps 149
5.24 Mean velocity profiles for the disc chamber at WOT, for three speeds; 20.0,
33.3 and 66.7 rps 150
5.25 Turbulent intensity profiles for the disc chamber at WOT, for three speeds;
20.0, 33.3 and 66.7 rps 150
5.26 Mean velocity profiles scaled with mean piston speed for the bathtub cham•
ber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps 151
5.27 Turbulent intensity profiles scaled with mean piston speed for the bathtub
chamber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps 151
5.28 Mean velocity profiles scaled with mean piston speed for the single slot
chamber at WOT, for two speeds; 20.0 and 33.3 rps 152
5.29 Turbulent intensity profiles scaled with mean piston speed for the single
slot chamber at WOT, for two speeds; 20.0 and 33.3 rps 152
5.30 Mean velocity profiles scaled with mean piston speed for the castellated
chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps 153
5.31 Turbulent intensity profiles scaled with mean piston speed for the castel•
lated chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps 153
5.32 Mean velocity profiles scaled with mean piston speed for the bowl-in-piston
chamber at WOT, for two speeds; 20.0 and 33.3 rps 154
5.33 Turbulent intensity profiles scaled with mean piston speed for the bowl-
in-piston chamber at WOT, for two speeds; 20.0 and 33.3 rps 154
xiv 5.34 Mean velocity profiles scaled with mean piston speed for the squish jet
chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps 155
5.35 Turbulent intensity profiles scaled with mean piston speed for the squish
jet chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps 155
5.36 Mean velocity profiles scaled with mean piston speed for the disc chamber
at WOT, for three speeds; 20.0, 33.3 and 66.7 rps 156
5.37 Turbulent intensity profiles scaled with mean piston speed for the disc
chamber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps 156
5.38 Brake thermal efficiencies for stoichiometric to lean operation for the 'bath•
tub' group of chambers at MBT, WOT and 20.0 rps 157
5.39 Brake thermal efficiencies for stoichiometric to lean operation for the 'bath•
tub' group of chambers at MBT, WOT and 33.3 rps 157
5.40 Brake thermal efficiencies for stoichiometric to lean operation for the 'bath•
tub' group of chambers at MBT, WOT and 50.0 rps 158
5.41 Brake thermal efficiencies for stoichiometric to lean operation for the 'disc'
group of chambers at MBT, WOT and 20.0 rps 158
5.42 Brake thermal efficiencies for stoichiometric to lean operation for the 'disc'
group of chambers at MBT, WOT and 33.3 rps 159
5.43 Brake thermal efficiencies for stoichiometric to lean operation for the 'disc'
group of chambers at MBT, WOT and 50.0 rps 159
5.44 Ignition advance for stoichiometric to lean operation for the 'bathtub'
group of chambers at MBT, WOT and 20.0 rps 160
5.45 Ignition advance for stoichiometric to lean operation for the 'bathtub'
group of chambers at MBT, WOT and 33.3 rps 160
5.46 Ignition advance for stoichiometric to lean operation for the 'bathtub'
group of chambers at MBT, WOT and 50.0 rps 161
xv 5.47 Ignition advance for stoichiometric to lean operation for the 'disc' group
of chambers at MBT, WOT and 20.0 rps 161
5.48 Ignition advance for stoichiometric to lean operation for the 'disc' group
of chambers at MBT, WOT and 33.3 rps 162
5.49 Ignition advance for stoichiometric to lean operation for the 'disc' group
of chambers at MBT, WOT and 50.0 rps 162
5.50 Ensembled fired and motored pressure profiles over four strokes for the
single slot chamber at WOT and 33.3 rps. Fired trace for MBT and
RAFR=1.00 163
5.51 Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT,
and 20.0 rps for RAFR=1.00 164
5.52 Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT,
and 20.0 rps for RAFR=1.27 164
5.53 Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT,
and 33.3 rps for RAFR=1.00 165
5.54 Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT,
and 33.3 rps for RAFR=1.27. 165
5.55 Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT,
and 50.0 rps for RAFR=1.00 166
5.56 Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT,
and 50.0 rps for RAFR=1.27 166
5.57 Fired pressure profiles for the 'disc' group of chambers at MBT, WOT,
and 20.0 rps for RAFR=1.00 167
5.58 Fired pressure profiles for the 'disc' group of chambers at MBT, WOT,
and 20.0 rps for RAFR=1.27 167
xvi 5.59 Fired pressure profiles for the 'disc' group of chambers at MBT, WOT,
and 33.3 rps for RAFR=1.00 168
5.60 Fired pressure profiles for the 'disc' group of chambers at MBT, WOT,
and 33.3 rps for RAFR=1.27 168
5.61 Fired pressure profiles for the 'disc' group of chambers at MBT, WOT,
and 50.0 rps for RAFR=1.00 169
5.62 Fired pressure profiles for the 'disc' group of chambers at MBT, WOT,
and 50.0 rps for RAFR=1.27 169
5.63 Fired pressure profiles for the 'bathtub' group of chambers at MBT, Bmep=2.5,
and 33.3 rps for RAFR=1.00 170
5.64 Fired pressure profiles for the 'bathtub' group of chambers at MBT, Bmep=2.5,
and 33.3 rps for RAFR=1.27 170
5.65 Fired pressure profiles for the 'disc' group of chambers at MBT, Bmep=2.5,
and 33.3 rps for RAFR=1.00 171
5.66 Fired pressure profiles for the 'disc' group of chambers at MBT, Bmep=2.5,
and 33.3 rps for RAFR=1.27 171
5.67 Fired pressure profiles for three different chamber geometries MBT, Bmep=3.5,
and 50.0 rps for RAFR=1.00 172
5.68 Fired pressure profiles for three different chamber geometries at MBT,
Bmep=3.5, and 50.0 rps for RAFR=1.27 172
5.69 Mass fraction burned curve for the bathtub chamber at MBT, WOT and
33.3 rps for RAFR=1.27 173
5.70 Mass fraction burned curves for different chamber geometries at MBT,
WOT and 33.3 rps for RAFR=1.00 174
5.71 Mass fraction burned curves for different chamber geometries at MBT,
WOT and 33.3 rps for RAFR=1.27 174
xvii 5.72 Mass fraction burned curves for the bathtub chamber at MBT, WOT and
RAFR=1.27 for five speeds; 20.0, 33.3, 40.0, 50.0 and 66.7 rps 175
5.73 Mass fraction burned curves for the 'bathtub' group of chambers at MBT,
WOT, and 20.0 rps for RAFR=1.00 176
5.74 Mass fraction burned curves for the 'bathtub' group of chambers at MBT,
WOT, and 20.0 rps for RAFR=1.27 176
5.75 Mass fraction burned curves for the 'bathtub' group of chambers at MBT,
WOT, and 33.3 rps for RAFR=1.00 177
5.76 Mass fraction burned curves for the 'bathtub' group of chambers at MBT,
WOT, and 33.3 rps for RAFR=1.27 177
5.77 Mass fraction burned curves for the 'bathtub' group of chambers at MBT,
WOT, and 50.0 rps for RAFR=1.00 178
5.78 Mass fraction burned curves for the 'bathtub' group of chambers at MBT,
WOT, and 50.0 rps for RAFR=1.27 178
5.79 Mass fraction burned curves for the 'disc' group of chambers at MBT,
WOT, and 20.0 rps for RAFR=1.00 179
5.80 Mass fraction burned curves for the 'disc' group of chambers at MBT,
WOT, and 20.0 rps for RAFR=1.27 179
5.81 Mass fraction burned curves for the 'disc' group of chambers at MBT,
WOT, and 33.3 rps for RAFR=1.00 180
5.82 Mass fraction burned curves for the 'disc' group of chambers at MBT,
WOT, and 33.3 rps for RAFR=1.27 180
5.83 Mass fraction burned curves for the 'disc' group of chambers at MBT,
WOT, and 50.0 rps for RAFR=1.00 181
5.84 Mass fraction burned curves for the 'disc' group of chambers at MBT,
WOT, and 50.0 rps for RAFR=1.27 181
xviii 5.85 Mass fraction burned curves for five different chamber geometries at MBT,
Bmep=2.5, and 33.3 rps for RAFR=1.00 182
5.86 Mass fraction burned curves for five different chamber geometries at MBT,
Bmep=2.5, and 33.3 rps for RAFR=1.27 182
5.87 Mass fraction burned curves for three different chamber geometries MBT,
Bmep=3.5, and 50.0 rps for RAFR=1.00 183
5.88 Mass fraction burned curves for three different chamber geometries at
MBT, Bmep=3.5, and 50.0 rps for RAFR=1.27 183
5.89 Mass fraction burned ratio bar graphs, relative to the bathtub chamber,
for MBT, WOT, and 20.0 rps for RAFR=1.00 184
5.90 Mass fraction burned ratio bar graphs, relative to the bathtub chamber,
for MBT, WOT and 20.0 rps for RAFR=1.27 184
5.91 Mass fraction burned ratio bar graphs, relative to the bathtub chamber,
for MBT, WOT, and 33.3 rps for RAFR=1.00 185
5.92 Mass fraction burned ratio bar graphs, relative to the bathtub chamber,
for MBT, WOT, and 33.3 rps for RAFR=1.27 185
5.93 Mass fraction burned ratio bar graphs, relative to the bathtub chamber,
MBT, WOT, and 50.0 rps for RAFR=1.00 186
5.94 Mass fraction burned ratio bar graphs, relative to the bathtub chamber,
for MBT, WOT, and 50.0 rps for RAFR=1.27 186
5.95 Mass fraction burned ratio bar graphs, relative to the disc chamber, for
MBT, Bmep=2.5, and 33.3 rps for RAFR=1.00 187
5.96 Mass fraction burned ratio bar graphs, relative to the disc chamber, for
MBT, Bmep=2.5, and 33.3 rps for RAFR=1.27 187
5.97 Mass fraction burned ratio bar graphs, relative to the castellated chamber,
MBT, Bmep=3.5, and 50.0 rps for RAFR=1.00 188
xix 5.98 Mass fraction burned ratio bar graphs, relative to the castellated chamber,
for MBT, Bmep=3.5, and 50.0 rps for RAFR=1.27 188
5.99 Indicated mean effective pressure per cycle for the single slot chamber at
MBT, WOT, and 33.3 rps for RAFR=1.27: 200 cycles 189
5.100Indicated mean effective pressure per cycle for the bathtub chamber at
MBT, WOT, and 33.3 rps for RAFR=1.27: 44 cycles 189
A.l Kistler and laboratory calibration curves for pressure transducer Model
6121 No. 317205, (used for single slot castellated, bowl-in-piston and fired
squish jet tests) 193
A.2 Kistler calibration curve for pressure transducer Model 6121 No. 282737,
(used for standard bathtub and disc chamber tests) 194
A.3 Kistler calibration curve for pressure transducer Model 6121 No. 317125,
(used for the motored squish jet test only) 195
A.4 Meriam calibration curve for Model 50MW20-1.5 No. S-4875-l,( used for
Natural gas flow rate). 196
A. 5 Meriam calibration curve for Model 50MC2-4F No. S-4875-2,( used for
Air flow rate) 197
B. l Hotwire calibration curve for Wire No. 1 Ramb = ll.MQ.Rop = 11.95£2,
(used for the bowl-in-piston chamber tests) 201
D.l Motored pressure trace at 50.0 rps low pass filtered at 6 kHz over the entire
range and over an initial region 208
D.2 Motored pressure trace at 50.0 rps with the region between 150 and 60
degrees BTDC replaced with a section under end tension 209
xx D.3 Motored pressure trace at 50.0 rps with the region between 180 and 60
degrees BTDC averaged and smoothed over 6 and 12 degree windows. . . 210
D.4 Expanded fired pressure traces at 33.3 rps, with and without averaging
and smoothing applied over 12 degree windows in the region between 150
and 55 degrees BTDC 211
xxi Nomenclature
A Calibration constant, hot wire calculation
AS Area of flame front
ABDC After bottom dead centre
ATDC After top dead centre
B Calibration constant, hot wire calculation
BDC Bottom dead centre
BBDC Before bottom dead center
BTDC Before top dead centre
BMEP Brake mean effective pressure (bar)
BP Brake Power (kW)
Bsfc Brake specific fuel consumption (g/kWhr)
Constant pressure specific heat (kJ/kg-K) cP
cv Constant volume specific heat (kJ/kg-K) ca Crank angle (degrees)
CL Con rod length
COV Coefficient of variance
D Piston bore d Bowl diameter
EVO Exhaust valve open
EVC Exhaust valve close h Convective coefficient (W/m2-K) i Number of record
xxii IMEP Indicated mean effective pressure (bar)
Integer Analogue interger
IVO Intake valve open
IVC Intake valve close
Lxg Lateral integral scale
Lxf Longitudinal integral scale
Lx Integral length scale
LT Integral time scale
LHV Lower Heating Value (kJ/kg)
MBT Minimum ignition advance for best torque (degrees rhb Mass engulfment rate
N Rotation speed (rps)
N Number of cycles n Calibration constant
Nu Nusselt Number
Pbar Pressure (bar)
PBDC Pressure at BDC (kPa)
Pamb Ambient pressure (kpa)
P{i) Pressure at i (kPa)
APpiiton Change in pressure due to piston motion (bar)
APcomb Change in pressure due to combustion (bar)
AP(i + 1) Change in pressure between point i and t + 1 (bar)
Pcomb{total) Total pressure change due to combustion (bar)
3 Qair Volumetric Air flow rate (m /sec)
Qtotal Total air and fuel flow supplied (m3/s)
xxni r Stroke
R(x) Spatial autocorrelation coefficient
RT Temporal autocorrelation coefficient
Ramb Ambient resistance of wire
Rop Operating resistance of wire
RAFR Relative air fuel ratio
Re Reynolds Number
RPS/rps engine speed revolutions per second
%SAREA Squish area
Sp Mean Piston speed (m/s)
Sg Position of piston from BDC position t Time
T Time period
T(i) Temperature at i (K)
Tamb Ambient Temperature (k)
Operating temperature (K)
Tamb Ambient temperature (K)
UT Turbulent flame speed (m/s)
U(t) Instantaneous velocity at time t (m/s) u(t) velocity fluctuation at time t (m/s) u Turbulence intensity (m/s)
U Mean velocity (m/s)
U(t,i) Instantaneous velocity at time t in record i (m/s)
U(tw,i) Average velocity at each window midpoint (m/s)
Uw{t,i) Interpolated window averaged velocity (m/s)
xxiv Uw(t) Window ensembled mean velocity (m/s) u\y{t,i) Fluctuating component of velocity (m/s)
u(tw,i) Fluctuating velocity at each window midpoint (m/s) u(tw) Averaged fluctuating velocity at each window (m/s) u\y{t) Representative rms intensity (m/s)
UE(t) Ensembled averaged mean velocity (m/s)
!££;(£, i) Fluctuating component of velocity (m/s)
UE(t) Rms intensity (m/s)
V Volume (m3)
VcBDC Volume at BDC before compression (m3)
VcBDC Volume at BDC after expansion (m3)
3 Vref Reference volume (m )
V, Swept volume (m3)
WOT Wide open throttle a Thermal coefficient of resistance (1/K) r)th Brake thermal efficiency (%)
7/„ Volumetric efficiency (%)
7C Polytropic coefficient of compression
7e Polytropic coefficient of expansion
A Ratio: crank length to conrod length
AT Taylor microtime scale
Az Taylor microlength scale
3 pu Density of unburned fluid (kg/m )
6 Crank angle from BDC (degrees) r Time (s)
xxv Acknowledgments
I am most grateful to Dr. R.L. Evans for his supervision and encouragement during this study. Further I would like to express my thanks for his concept of the 'squish jet' piston and for the opportunity to contribute to its evaluation. I also wish to express my appreciation to Professor P.G. Hill for his interest, enlightening discussions and guidance during Dr. Evans' research leave at Cambridge University.
I would like to thank the academic and technical staff of the Mechanical Engineering
Department especially:
A. Steeves for his help with the computing system;
S. Oshika and J. Richards for their assistance in the operation and 'fault
finding' of the test engine;
T. Besic for the machining of the modified pistons;
Dr. K.V. Bury for his clarity in statistical matters;
A. Kapil and other recent graduates of the AFL Group for their support and
legacy of information.
Also I record my appreciation of the interest and financial support provided by the
Canadian Gas Association for the Alternatives Fuels Laboratory.
Finally I would like to thank Professor J.D. Willms and R. Love for the use of their facimile system and Professor Helen Tippett, Victoria University of Wellington, for help in the production of this thesis.
xxvi Chapter 1
Introduction
1.1 Introduction and Background
Since the first days of engine design, studies aimed at improving the combustion process to optimize spark ignition engine operation have been much in evidence. Initially it was
sufficient to achieve smooth running at maximum power but the area of interest has expanded to meet specific requirements. The energy crisis of the 1970's promoted inves•
tigations into alternative fuels. In the increased environmental concerns and restrictions on exhaust emissions of the present day, improvements in the combustion process have
become a high priority.
To improve engine operation, the engineer is concerned with fuel economy, knock
propensity and emissions, which are all combustion issues.
Work output from the SI engine is obtained from a 'stepwise' cyclic production of
power, based on the successive burning of a fixed amount of charge in the combustion
chamber. The requirements of maximum pressure development and repeatability of the
torque output determines the acceptable range of operation of the system.
As the operating characteristics are pushed to the extremes of charge dilution for
economic and environmental reasons, the effect of cycle-to-cycle variations is increased.
Cycle-to-cycle variations in the combustion process from whatever cause are mani•
fested in cyclic variations in the output torque and consequently affect drivability. Where
these cyclic variations are large, the problems associated with operation at the extreme
1 Chapter 1. Introduction 2
limits, such as misfire and knock, can seriously affect the engine performance.
In balancing the requirements of modern engines, researchers and design engineers have aimed at the so called 'fast-lean' burn operation.
Lean burn operation
Lean operation, that is, at a relative air fuel ratio greater than stochiometric, has the advantage of increased fuel economy, higher thermal efficiency and lower levels of NOx emissions. In addition to efficiency from lower fuel consumption, the lower combustion temperatures developed and reduced pumping losses at part load associated with lean burn improve the efficiency.
The disadvantages of lean operation result from the slower burning leading to in• creased cyclic variations, and from higher engine operating temperatures due to the longer burn period. At the lean limit also, the level of unburned HC rises due to misfire and partial burn. At a normally stable operating condition cyclic variation, may cause the engine to exhibit knock tendencies or to misfire due to an extreme cycle being at the lean limit.
Fast burn operation
Fast burn operation has the advantage of increasing thermal efficiency as the combustion process approaches constant volume combustion as represented in the ideal Otto cycle, with higher peak pressure and hence greater indicated work. Fast burn also decreases the negative effects of lean operation, reducing cyclic variations and resulting in an extension of the lean limit and improved knock limits.
The disadvantages of fast burn are higher emissions of NOx from higher combustion temperatures, although more complete burning may reduce HC levels. Increased engine noise and roughness from the high rate of pressure rise can also be a limiting factor. Chapter 1. Introduction 3
Fast burn operation is particularly important with the use of slow burning fuels, such as natural gas.
Stable fast-lean burn operation
A balance of stable fast-lean burn operation, with minimal cyclic variation may be achieved by various means: optimization of mixture motion, increasing the flame speed through increased turbulence; optimization of the ignition system and geometry, selecting the number and position of spark points for shorter flame travel; and finally,throug h the chemical system by improving the reaction kinetics using additives for enhanced stability.
Extensive research into the nature of cyclic variations in engines has been carried out through turbulence studies in bombs, rapid compression machines and SI and CI engines.
While complicated by the interaction of the above variables, and the controversy of turbulence definitions in the engine environment, these studies have confirmed the early conclusions of Karim [1]:
"By far the greatest variable element in engine operation is that of charge
motion inside the cylinder, both in quality and direction."
A review by Young [2] confirmed that cyclic variations in both scale and intensity of the flow near the spark plug at the time of ignition, would have a profound effect on the cycle- to-cycle combustion characteristic. In his study of the causal relation between engine variables and cyclic variations in combustion, Young [4] deduced that while engine and operating variables may significantly affect combustion variation, they do not necessarily cause it. He concluded that in-cylinder velocity variations near the spark were the major cause.
With a predetermined engine system and operating fuel, the potential for optimization of the ignition and chemical systems is often limited. The influence of mixture motion on Chapter 1. Introduction 4
cyclic variations is therefore the area of greatest interest. In 1967, Karim [1] provided an early incentive for continued research in this area when he estimated economic increases through the reduction of gross variability as 10-20% at lean operation and 3-5% at optimum operating levels.
1.2 Objective of this study
The overall objective of this study was to investigate the influence of combustion chamber geometry on turbulence enhancement in the achievement of fast-lean operation of an SI engine.
Before discussing the scope of the work undertaken, the remainder of this chapter provides a brief background on the turbulence flow field in an engine and landmark studies in this field and discusses the terminology used.
1.3 Turbulent Flow Field in an Engine
The in-cylinder flow in an engine is determined by turbulence production influenced by geometric and dynamic considerations. Turbulence production occurs through the effect of piston motion on the fluid, shear flow past the intake valve and amplification of in- cylinder turbulence during the compression stroke due to the change in volume. The role of turbulence in affecting the flame speed was critically examined by Andrews et al [3] who showed that a higher level of turbulence increased the turbulent flame speed and burn rate. Various studies have shown that turbulent flame speed can be an order of magnitude greater than the laminar flame speed.
Mechanisms for increasing the burning rate by turbulence, postulated in several stud• ies, have been classified into both scale dependent and independent means.Damkohler's [5] flame wrinkling model is based on the effect of eddies, on a scale less than the thickness of Chapter 1. Introduction 5
laminar flame front, causing increased flame front area. Shchelkin's [6] model is based on the effect of eddies, on a scale larger than the flame thickness, higher gradients causing distortion of the flame front and hence increased area.
The models developed by Tabaczynski [7] and Blizard and Keck [8] are based on en• trapment of intermittent regions of activity in the flame front involving microconvection mixing of burned and unburned regions operating at all scales. The mass engulfment rate is determined by:
rhb = UT * Af * pu
where UT is the turbulent flame speed, Af is the area of the flame front and pu is the density of the unburned fluid. The turbulent flame speed being a function of laminar flame speed is also dependent on stoichiometry and gas properties.
The turbulent flow field may be characterized by a set of parameters using the theory of isotropic turbulence. Two distinct features of engine turbulence noted by Tabaczynski
[9] are that it was periodic and of the order of the mean. The study of turbulence in engines is further complicated by real engine flow parameters involving highly unsteady, inhomogeneous, non isotropic conditions and changing thermodynamic and geometric conditions.
The results of turbulence studies in engines from the pioneer work of Semenov [10] using Hot Wire Anemometry (HWA) techniques to more recent Laser Doppler Anemom- etry (LDA) work by Fraser and Bracco [11] are discussed in Chapter 2. The literature review also discusses the engine and operating variables studied by Young [2] and others, and their relative influence on cyclic variation. Chapter 1. Introduction 6
1.4 Discussion of Terminology in Turbulence Studies
As noted previously, there is some controversy in turbulence definitions in the engine environment.
Turbulence is denned as a quantity that shows random variation with respect to time and space from which statistically distinct averages can be discerned [12]. Separation of such averages with engine velocity measurements pose problems of separating cycle-to- cycle variations from turbulence velocity fluctuations.
Mean velocity is determined from the instantaneous velocity U(t) at any time t.
Turbulence is then characterized by the following means.
Mean velocity is determined:
where T is a sufficiently long time period to contain the turbulent fluctuations. The velocity fluctuation is then separated by:
u(t) = U(t) - U
Turbulence intensity is defined as the root mean square, rms, fluctuation u:
The periodic nature of engine turbulence requires non-stationery methods for distin• guishing the mean from the instantaneous velocity determined for each record t. The definition of turbulence used by Lancaster [13] was defined by the fluctuating component determined by:
u{t,i) = U(t,i) - U(t) - U(i) where U(t,i) is the instantaneous velocity at time t in record i, U(t) is the ensemble averaged mean at time t over record i's, and U(i) is a time averaged mean after U(t) has been subtracted from the instantaneous velocity for each record i. Chapter 1. Introduction 7
Using a similar approach Heywood [14] questioned whether the turbulent fluctuating components of the instantaneous velocity could be separated from the cycle-to-cycle variations.
The non-stationary method used in this study is discussed in detail in the analysis of test data, in Chapter 4. With the selection of an appropriate window, the non-stationary method uses a window averaging technique in place of the above equations.
Turbulence is distributed over a continuous range of eddy sizes. Scales of turbulence, a representative measure of the average values of quantities with respect to time and or space are calculated by correlation techniques. The spatial integral length scale is defined:
where R(x) is the spatial autocorrelation coefficient. In general practise, Taylor's hy- pothesis [15] relating spatial and time scales is applied. The time scales are determined from the temporal autocorrelation:
1 fT u{r)-u{t + T) dt where t is the correlation time and T is the period of measurement. Applying Taylor's hypothesis the spatial integral length scale is calculated:
U*LT
The Taylor microtime scale can also be calculated from this autocorrelation:
'\d2R{r)y K = 2 dr2 and similarly the microlength scale:
Xx = U * K Chapter 1. Introduction 8
For homogeneous isotropic turbulence the longitudinal integral scale is related to the lateral by:
Lxf = 2 * Lxg
Generally, the integral length scale in an engine is of the order of chamber height, 10mm and seen as a typical mixing length. The Taylor microscale is of the order of 0.1 mm and the Kolmogorov Scale at which dissipation occurs is of the order 0.01 mm.
The validity of the assumptions of homogeneous and isotropic turbulence used by
Taylor in engine operation is discussed in the literature review. Of the characteristic turbulence parameters described above, Andrews and Bradley [16] determined that the role of turbulence scale was important in the initial stages of combustion only, and that turbulence intensity was primarily responsible for increased flame speed.
1.5 Scope of Work
The main means of turbulence generation in an engine is by shear flow past the intake valve, which decays during the compression stroke. Turbulence enhancement is achieved by the break down of large scale organized motion either from 'swirl'- a tangential rotat• ing flow generated during the intake stroke through inlet port geometry or by shrouded valves, or from 'squish'- a radial injecting flow generated during the compression stroke before TDC: or a combination of these two flows.
As defined in section 1.2, the overall objective of this study was to achieve fast-lean burn operation of an SI engine through turbulence enhancement. The advantages of an increased burning rate to improve efficiency and reduce cyclic variations inherent in lean operation has already been discussed. This is of particular importance in slow burning alternative fuels. The specific focus of the work was to improve understanding of squish and enhanced squish effects in promoting such operation. Chapter 1. Introduction 9
A comparative study of chamber design, restricted to squish type combustion cham• ber configurations, was performed in two areas. First, a qualitative and quantitative comparison of turbulence concentrated on the intensity parameter was carried out, using hot wire measurements at the spark location. Second, the effect of chamber design on the engine performance was assessed, with a subsequent correlation of the turbulence and combustion measurements.
Analysis and evaluation of the experimental data was directed towards characterizing the cyclic variation through a statistical analysis of indicated mean effective pressure
(IMEP), and peak pressure variation. Investigation of the combustion process initial and main burn periods was aimed at determining the primary area of influence of each chamber design, and the link with the turbulence effects.
The area of particular interest in the study was the enhancement of the squish effect using squish jet pistons. However, consistency and accuracy of the test base for com• parison was also a specific interest. The engine operation variables therefore involved a range of speed, loads and air fuel ratios.
One of the difficulties associated with engine studies has been the multitude of inter• acting engine variables and dependence on the conditions of operation. Earlier engine studies, restricted to low operating speeds and cumbersome data acquisition systems were inclined to draw conclusions from single sets of data. More recent studies have qualified their results for specific engines or specific sets of conditions.
The range of test conditions and measurement parameters used in this study were selected with the aim of providing the broadest practical comparison with previous studies in this field, and to give a consistent basis for comparison with other current work.
Hence a standard bowl-in-piston was used as a comparison for the squish jet bowl-in- piston previously studied by Cameron [17]and Dymala-Dolesky [18]. A standard bathtub cylinder head chamber was used as a basis for comparison with studies using the new Chapter 1. Introduction 10
modified 'jet' bathtub pistons. Disc piston measurements were also taken as a non-squish case.
1.6 Structure of thesis
Chapter 1 has discussed the study objectives and provided a background to the continued interest in fast-lean burn operation as a means of improved engine operation, particularly in the use of natural gas and other alternative fuels.
Chapter 2 presents a review of the extensive literature on turbulence and combustion studies in engines, with a final section on combustion chamber design.
Chapter 3 describes the experimental apparatus and data acquisition systems used in this study, the analysis of which are described in Chapter 4.
Chapter 5 discusses the experimental results in detail, the constraints and limitations, and relates the present findings to previous studies of turbulence and combustion.
Chapter 6 presents the conclusions from this investigation and suggest directions for further research aimed at improving the combustion process to optimize spark ignition engine operation.
The list of references cited in this thesis precedes the tables, figures and graphed results.
Appendices A and B contain further details on instrumentation specifications and calibration procedures. Appendix C presents data on the properties of BC natural gas and lower heating value calculations. The final Appendix contains details of the filtering methods examined and used in the pressure data analysis. Chapter 2
Literature Review
2.1 Introduction
The review of the literature and background of engine studies has been divided into three sections. The first section follows the successive progression of engine investigations of turbulence from early bomb work in the 1970's to present day computer simulation models. The second section covers combustion studies in the same period. The last section involves both turbulence and combustion studies in combustion chamber design and includes an evaluation of previous research on squish jet pistons.
Studies of turbulence in engines, including the related studies on flow visualization bombs and rapid compression machines, have been concerned with characterizing turbu• lence, determining its source, achieving a reliable means of measurement and assessing its effects on cyclic variations in combustion. An outline of the major stages in turbulence studies is given with a summary of the findings together with a review of the techniques used in measurement.
The combustion studies review summarizes the engine variables investigated in assess• ing the causes and effects of cycle-to-cycle variations. Emphasis has been placed on those variables controlled in this study and those most likely to influence the test conditions.
Characterization of cyclic variations through pressure measurements is also reviewed.
More detailed reviews, on fluid motion within the cylinder of internal combustion
11 Chapter 2. Literature Review 12
engines, cyclic dispersion in spark ignition engine and the use of cylinder pressure mea• surement in combustion studies of engines, are contained in papers by Heywood [14],
Young, up to 1980 [2] and Amann [19] respectively.
2.2 Turbulence studies in Engines
Semenov's [10] studies of turbulent gas flow in piston engines in 1963 is regarded as a fundamental work. Hot wire anemometry (HWA) studies in a motored CFR engine with disc piston showed that shear flow through the intake valve was the primary source of turbulence generation followed by decay during the compression stroke and relaxation towards isotropic turbulence at top dead centre (TDC).
Tri-axial HWA measurements were carried out by Lancaster [13] in a motored CFR engine, with and without swirl, generated by a shrouded intake valve. His studies con• firmed isotropy and showed that the mean and rms velocity increased linearly with engine speed. Similar investigations by others [20, 21, 22, 23] also showed isotropy and scaling with engine speed. Bopp, Vafidis and Whitelaw [22] estimated scaling of turbulence at
TDC with mean piston speed in the range of 0.47 to 0.60 comparable with the findings of 0.5 for a disc chamber by Hall and Bracco [24].
Most recently, laser homodyne measurements in an engine with a flat topped piston and varying swirl conditions by intake geometry by Ikegani et al [25], confirmed that in no-swirl cases the turbulence field near the end of compression is almost uniform.
Hot wire anemometry
The major method of turbulence measurement in the 1970's was with hot wire anenom- etry, and extensive investigations were made into the accuracy of such methods [21]. Chapter 2. Literature Review 13
Dent and Salama [26] and Haghgooie et al [27] were among the first to measure tur•
bulent time scales using HWA techniques with high speed random signal analysis. Using
the above isotropic and homogeneous assumptions microlength scales were estimated to
be of the order of 1/10 mm.
Laser Doppler anemometry
Early laser Doppler anenometry (LDA) measurements such as those by Rask in 1979
[28] encountered problems with seeding, vapourization and visualization requirements of engine design. Advances in LDA techniques have been used by many researchers to verify HWA results and obtain more direct scale measurements. Fraser et al [29] measured integral length scales in a motored disc engine with and without swirl. They confirmed early measurements of Lancaster [13] and Hey wood [14], that is, integral length scale was of the order of the chamber clearance height at TDC, a l/5th ratio was proposed.
Fraser and Bracco [11] used a two-point, single probe-volume Laser Doppler Velocime-
try system (LDV) to measure length scales directly. Their results appear to indicate that for the specific chamber investigated, the lateral integral length scale from both
cycle-resolved and ensemble averaging did not scale with the clearance height, but was independent of it around TDC. A correlation between the larger spatial scales and the lower frequencies of fluid flow was also demonstrated.
These scale studies have overshadowed research based on Lancaster's [13] and Se-
menov's [10] conclusions that the turbulent intensity was the single parameter of impor•
tance. Later studies noted however, that the Taylor approximation for length scales still had value due to its ease of calculation.
Similarly, HWA methods are still in use due to their relative simplicity and inexpense.
Analyzing techniques suggested by Witze [30] from the comparison of hot wire anenome•
try and laser Doppler velocimetry used in IC engines, provide the basis for the continued Chapter 2. Literature Review 14
use of HWA as selected for this study.
The final area of turbulence studies reviewed concerns those linking the turbulent flow field and its effect on the flame speed, with combustion studies.
Matsuoka et al [31] compared flame arrival times by ion gap measurements in a fired engine, to mean velocities measured at the spark location while motored, using swirl and non-swirl conditions. They concluded that the mean velocity did not influence the flame speed, but that the small scale turbulence, produced by the breakdown of large scale motion, ie., swirl, did.
Schlieren photography techniques used by Gatowski et al [32] showed a wrinkled flame front developed after formation of the spherical flame kernel at spark, indicating laminar like burning immediately following the spark. Similarly, work by Keck, Heywood and
Noske [33] on early flame development using schlieren techniques, supported the wrinkled laminar flame model of turbulent structure.
The applicabilityof turbulence measurement in motored engines to fired engine oper• ation was originally supported by LDA measurements in a fired engine at 800 RPM by
Rask [28]. These showed that motoring and firing results were quite similar until the time of ignition and differed only when the flame was near the location of the velocity measurement. This work was also consistent with the rapid distortion theories of Witze and Martin et al [34, 35] of turbulent intensity increasing ahead of the flame from one dimensional of compression of the unburned charge.
Barton et al's [36] autocorrelation of peak pressure and cycle, showed no correla• tion, leading to the conclusion of 'no history' in the pressure fluctuations. This study found that cyclic variations depended on the intake, compression and combustion strokes of each individual cycle, allowing comparisons of fired cycles with motored intake and compression stroke analysis.
Advances in computer systems and analysis methods have shown promise of analytical Chapter 2. Literature Review
methods for prediction of flow behaviour. However Gosman [37] cautions that,
"The accurate prediction of the flow behaviour during individual cycles,
which is of interest in correlation with cycle-to-cycle variations in combustion
performance is believed to be outside the capabilities of the present method•
ology.".
Summary of turbulence study findings
The general findings of the turbulence studies since Semenov's pioneering
research in the early 1960's have been:
• Turbulence approaches a homogeneous and isotropic state near TDC of
compression.
• In the absence of turbulence production generated through chamber ge•
ometry, shear flow past the intake valve is the major source of turbulence
in the engine cylinder.
• Mean velocity and turbulent intensity are proportional to engine speed,
with a near linear relation.
• Cyclic variations increase with increased engine speed.
• In chambers without swirl or squish turbulent intensity at TDC scales
with mean piston speed, in the range 0.47 to 0.60.
• Turbulent length and time scales are inversely proportional to engine
speed.
• The integral length scale is of the order of chamber height at TDC.
• Turbulent microtime scales are of the order of 0.1 millisecs.
• Turbulent microlength scales are of the order of 0.1 mm Chapter 2. Literature Review
• Turbulent energy is concentrated in the low frequencies; below 1000 HZ.
• Flame travels as a laminar flamelet immediately after spark.
• Hot wire measurements in the motored engine may be approximated to
conditions in the fired engine prior to combustion.
The last point forms the basis of motored engine turbulent studies use in
combustion studies.
2.3 Combustion Studies in Engines
Combustion engine studies have been used to establish which engine variables
are responsible for cyclic variations, methods to measure this variation and
the means to control cyclic variations through the most influential variable.
Combustion experiments in engines were initially concerned with deter•
mining the source of the cycle-to-cycle variations and influencing factors.
Early investigations covered both CI diesel and SI engines with limited appli•
cations of information between valve and ported engines. Factors investigated
were, mixture type and preparation, the ignition system, engine speed, com•
pression ratio, combustion chamber geometry and mixture motion. At the
same time, the best means of characterizing the cycle-to-cycle variation mani•
fest in torque output and related to the pressure development, was evaluated.
From the recognition of the importance of in cylinder mixture motion,
investigations continued in the direction of increasing turbulence, through
combustion chamber and inlet geometry, and correlations of combustion phe•
nomena with turbulence. Barton et al's [36] statistical analysis of performance
data in a fired CFR engine confirmed Karim's [1] early theory that mixture Chapter 2. Literature Review
motion was the primary cause of combustion pressure variations. The impor•
tance of each engine operating variable was then determined by its affect on
flame speed, flame distance travelled, variations of the mixture motion or a
combination of these.
The difficulty encountered with much of the engine research is the inter•
action of engine variables and operating conditions. Isolation of the various
effects has often been impractical. Steps to overcome these problems in this
investigation are discussed in Chapter 3.
Chemical variables
Chemical factors of fuel air mixing, equivalence ratio, residuals and fuel type
were investigated by skip firing tests by Soltau [38] and similar methods by
others [1, 39]. Minimum variation was found at the equivalence ratio giving
the maximum power and shortest combustion duration. Hence, fuel air mix•
ture and residuals were found to be important to the extent they affected
the equivalence ratio; generally richer mixtures and faster burning fuels had
the least fluctuation. A semi-quantitative study by Hansel [40] verified that
leaner mixtures produced larger variations.
Ignition variables
Ignition variables, type of system, spark electrode geometry, and spark jit•
ter were shown to have negligible effects on cyclic variation in early studies
[2], aside from any change in the number or location of spark points which
decreased the maximum flame travel distance and hence decreased burning
duration. Recent investigators [41, 42] have found that the arc duration and
gap have significant effects on the combustion quality and stability in lean Chapter 2. Literature Review
mixtures, with the effects becoming more pronounced as ignition timing is ad•
vanced and load reduced. Kalghetti [43] argued that, regardless of the system,
once an ignition criterion and initiation criterion were met cyclic variations
could not set in once the flame kernel was larger than a critical size, with
the exception of some cyclic variation in the local mixture strength near the
spark. This does not necessarily contradict the early investigator's results
which were conducted principally at full load stochiometric conditions.
Engine speed
Winsor and Patterson [20] concluded that there was an overall weak relation
between engine speed and cyclic variations. An increase in engine speed, while
producing an increase in flame speed also showed an increase in variation of
peak pressure at stochiometric conditions, related to increased combustion
variations from the increased turbulence variations. Karim's [1] work with
iso-octane in the lean range showed a decrease in the coefficient of variance
with increased speed.
Similarly, compression ratio effects have generally been shown to be weak.
Barton et al [36] showed an increase in compression ratio slightly decreased
the cycle-to-cycle variation through increased flame speeds. The higher tem•
peratures and lower residuals inherent in higher compression ratios also affect
the reduction in cyclic variations [13].
Combustion chamber design
Optimization of combustion chamber design for the desired turbulence gen•
eration, is reviewed in Section 2.4. There is a general consensus that the best
combustion chamber to affect a reduction in cycle-to-cycle variations provides Chapter 2. Literature Review
the shortest combustion duration, ie., more open and more compact chambers
combined with a central spark location are preferred [4].
From his review of the results of early engine combustion research Pat•
terson [39] reiterated the importance of variations in mixture motion. More
recent studies have confirmed that temporal variation in velocity gradients
and turbulent mixture motion is very influential, causing variation in growth
rate, location of the flame kernel and flame speed at the spark location [33].
Pressure measurements
In the assessment of cyclic variations and combustion characteristics, piezo•
electric pressure transducer measurement of the in-cylinder pressure has been
the fundamental tool [19, 39, 44]. Results of early piezo-electric pressure
transducer selection and operation are discussed in the apparatus section.
Various characterizing parameters have been used:
• Maximum cylinder pressure within each cycle.
• Maximum rate of pressure rise.
• Crank angle at which peak occurs.
• IMEP covariance calculations.
• Burning times — combustion duration calculation.
• Flame arrival times.
Many researchers have used peak pressure [1, 20, 36, 38, 45, 46] as being
strongly related to the combustion rate however, Matekunas [45] cautioned
that peak pressure was not a good indicator of drivability. Similarily, crank
angle occurrence of peak pressure studied by Barton et al [36] was found to Chapter 2. Literature Review
be less useful than peak pressure. Nagayana's [47] studies on disc, squish
and swirl chambers showed a measurable relation between the cycle-to-cycle
fluctuations in IMEP and the vehicle surge limit. Amann's [19] review of
pressure characterization also gave a poor correlation between IMEP and peak
pressure. Kuroda's [48] fast burn studies showed that variation of the order
of 10% in IMEP coefficient of variance seriously affected vehicle drivability.
Combustion durations have been determined from flame arrival times
using ionization gap methods and heat release analysis methods using in-
cylinder pressure data. The combustion models have ranged from a simple
pressure ratio model developed by Rasswieler and Withrow [49] using com•
parison with flame pictures, to more complex heat release models detailed by
Gatowski [50] based on Krieger and Borman'6 [51] two zone model for spark
ignition engines.
Combustion process
The combustion process can be commonly divided into three or four phases:
1. ignition and kernel development.
2. flame development.
3. fully developed.
4. termination.
The first two phases are commonly termed 'the early combustion period'
and the third referred to as 'the main combustion period'. To date, most
interest has been taken in the early and main combustion periods, however,
recent remarks by Amann [52] suggest that the continued restrictions and Chapter 2. Literature Review
requirements of engine operation will require investigation and optimization
of the complex termination phase.
Effect of turbulence on combustion
Lancaster et al [44] conducted one of the first correlation studies of turbu•
lence and combustion data using HWA and pressure measurements, with heat
release model techniques. Comparing a disc chamber for a swirl and, a non-
swirl case, Lancaster et al found that the normalized flame speed was a linear
function of turbulence intensity with little influence of turbulent scale within
the limits of the measurements. Using LDA techniques Cole and Swords [53]
found a strong correlation between the mean velocity and peak pressure with
a weak relation with turbulence intensity.
Some other researchers, notably Nagayana et al [47] and Matekunas [45]
have examined the effect of turbulent generation by squish and swirl chambers
on combustion, using pressure combustion data only. Nagayama's comparison
of plain, swirl, squish, and swirl-squish chambers showed that squish was
active in the main phase of combustion and swirl in the early phase. From
the higher performance of the swirl-squish chamber Nagayana argued that the
squish generated before TDC was important in breaking up the swirl motion
generated during intake. Matekunas's comparison of three levels of swirl with
a disc chamber, suggested that in the absence of a large scale flow motion,
such as swirl, to provide energy to the smaller scales a lower mass burn rate
resulted. The effect of early flame development was also seen to be important
near the lean limit with negligible effect away from this limit.
Belmont et al's [54] extensive statistical analysis of cyclic variability in an
SI engine showed a strong memory element under certain conditions. Memory Chapter 2. Literature Review
was shown to be inversely proportional to the degree of cyclic variability.
One final point noted by Daily [55] on the inherent chaotic nature of
combustion, was that the longer the burn duration, the more control non•
linear combustion kinetics exerted over cyclic variation.
Operating variables
Although there is a plethora of engine performance data on various com•
bustion chambers and engine configurations, there has been little attempt
to standardize or isolate the variables examined. Early investigations of the
effect of swirl on SI and CI engines by Ma [56] showed negligible effects of
swirl in an SI engine. This was contradicted by Witze's studies [57, 58] which
included the effects of other parameters, ie., engine geometry and operat•
ing conditions. Witze also showed that cyclic variation was not necessarily
decreased by increasing the burn rate. Matekunas [45] found that very high
swirl caused pre-mature detachment of the flame kernel or possible quenching
of the flame by convection against the cylinder walls. The negative effects of
partial or complete quenching caused by excessive turbulence have also been
considered a limiting factor in improvement in lean burn operation in more
recent work by Sheppard and Bradley [59] and Saxena and Rask [60].
Summary of combustion study findings
The general findings from combustion studies on the influence of turbulence
on cyclic variation are:
• The greatest variable element in engine operation is the charge motion
within the cylinder on the micro-turbulence level. Chapter 2. Literature Review 23
• Cyclic variation presents a major obstacle to implementation of lean
burn.
• Cyclic pressure variation is a direct consequence of variation in the com•
bustion process and the rate of heat release.
• The strength of the relationship between peak pressure and indicated
work is dependent on the operating conditions.
• The maximum rate of change of peak pressure is less sensitive to indi•
cated work than peak pressure.
• Near the lean limit variation in the early flame development, and hence
fluctuation in the combustion duration, is the main cause of cyclic vari•
ation.
• Maximum engine stability with minimum IMEP variation is that with
the fastest and steadiest combustion.
• Reduction of the main combustion time mitigates the effects of early
variations and the converse.
• Longer main combustion duration allows greater influence of non-linear
combustion kinetics.
• Increase in cyclic variation reduces the degree of memory.
• Reduction of the coupling between IMEP and combustion through op•
timum phasing reduces the cyclic effects.
• Squish is most effective in the main part of combustion.
• Swirl is effective in the early part of combustion. Chapter 2. Literature Review
2.4 Combustion Chamber Design
The final section of this review deals with the studies of optimizing turbulence
by combustion design with an emphasis on squish.
Studies have shown [14, 61] that the optimum chamber design should
provide fast repeatable burn for high efficiency; good emission control through
large valve effective areas; and low wall surface areas to minimise heat loss and
avoid quenching and also must be geometrically practical for manufacture.
Gruden's [62] investigation of combustion chamber layout in modern pas•
senger cars showed that a combustion chamber located in the piston crown
was the simplest way to comply with these requirements. In line with Young's
[4] optimization criterion of an open and compact chamber, Gruden concluded
that the dimensions and the positions of the quench areas and the quench dis•
tance, ie., maximum flame travel, were important.
Overington and Thring's [63] work on a Ricardo hydra engine with variable
compression ratio and combustion chambers in the head and piston showed a
2-5% improvement in fuel economy for the chamber in the piston crown over
that of the head chamber, arguably from higher turbulence during combustion
induced by squish.
Squish effects
Early studies of the squish effect reviewed by Young in 1980, were predomi•
nantly in diesel configurations and gave conflicting or negligible results. While
more recent studies in diesel engines [64, 65] provide information on squish
and swirl interactions and the effects of combustion chambers shape on fluid
motion, only limited comparisons can be made with the SI engines. Chapter 2. Literature Review
Evans [66] proposed a variation of the standard bowl-in-piston squish de•
sign in 1985. He used channels in a bowl-in-piston chamber to enhance the
squish effect through developed jets. This design was evaluated analytically
and with HWA and combustion pressure measurements in a CFR engine by
Cameron [17, 67]. Initial results showed an increase in peak pressure and re•
duction in combustion duration when compared with standard bowl-in-piston
squish pistons.
Subsequent investigations were carried out by Dymala-Dolesky [18] to
evaluate the nature of the jets developed. The jets in general were shown to
diminish the effect of the main squish motion. IMEP covariance calculations
and mass fraction burned analysis indicated that the squish jet was most
effective in the latter half of combustion. The most promising chamber was
with eight jets angled towards the centre of the bowl, at spark location.
The effect of jet flow introduced in the cylinder was also investigated by
Nakamura et al [68]. A piston operated jet valve was used to direct air or
a super lean mixture towards the spark plug. A strong swirling flow was
observed by ionisation flame arrival time measurements, and improvements
in peak pressure were attributed to this. The optimum diameter and injected
fluid rate was 6 mms and 0.97 1/s equivalent to a jet of approximately 34 m/s.
The maximum jet velocities observed by Dymala-Dolesky at spark were 5-
20 m/s suggesting stronger jets formed by blocking the main squish effect
may have a more definite effect on the motion.
As defined in the objective, the principle aim of this project was to achieve
fast-lean burn operation of an SI engine through turbulence enhancement.
The findings from the turbulence and combustion studies reviewed in this Chapter 2. Literature Review
Chapter suggested several promising directions for further research. Chap•
ter 3 describes the apparatus and data acquisition selected in this study to
examine the effect of piston chamber geometry on turbulence performance. Chapter 3
Experimental Apparatus and Method
3.1 Introduction
The experimental investigation of the effects of combustion chamber design on turbulence, cyclic variation and performance of a spark ignition engine, was conducted in a Research Engine test cell facility of the Alternatives Fu• els Laboratory (AFL), in the Department of Mechanical Engineering, The
University of British Columbia.
The aim of the experimental measurement was to provide turbulence, combustion and performance data for a number of different combustion cham• bers. The ability of the Ricardo Hydra research engine to be run at speeds representative of modern IC engines (20-90 rps), and the ease of conversion to a number of different configurations using alternative cylinder heads and pistons, has led to its use in this work and at other research centers. Six combustion chamber configurations were studied using two types of cylinder head and six different pistons.
The tests for each chamber configuration were divided into two phases:
1. Motored engine tests using a DC dynamometer, to yield phasing and
flow measurement data.
2. Fired engine tests using BC natural gas as a fuel to yield combustion
and performance data.
27 Chapter 3. Experimental Apparatus and Method
The motored tests were run over a range of speeds at wide open throttle
(WOT). The fired test regime involved tests run at WOT full load, over
a range of speeds, and part load at two speeds for stoichiometric to lean
operation
In this chapter the experimental apparatus, instrumentation and the data
acquisition systems are first described, followed by a description of the two
test procedures. Calibration requirements, tests and data are also detailed.
3.2 Experimental Apparatus
3.2.1 Introduction
The general arrangement of the Ricardo Hydra Test Cell with an IBM PC link
setup in an adjacent control room is illustrated in Figure 3.1. The test cell con•
sists of a base mounted Single cylinder engine with electrical dynamometer,
oil and coolant modules, converter cabinet,transformer and control console.
The engine, fully instrumented with electronic transducers as supplied by
Ricardo Consulting Engineers, was equipped with a Cussons electronic con•
trol unit. This unit with instrumentation and data acquisition systems added
by the AFL Group enabled control and measurement of the main engine pa•
rameters, such as torque, speed, flow rates, throttle position, pressures and
temperatures through signals fed into an IBM personal computer. Measure•
ment of cylinder pressure and hotwire signals was also possible.
The basic engine system components relevant to this experimental work
are briefly described in the next sections. The instrumentation for monitoring
and obtaining performance data is described in section 3.2.4. More detailed
information is available in the AFL and Ricardo operation manuals [69, 70]. er 3. Experimented Apparatus and Method
3.2.2 Engine Bed
The Ricardo Hydra is a single cylinder, four stroke, water cooled spark igni• tion engine with vertical valves operated by an overhead camshaft. The 0.45 litre capacity engine has an 80.2 mm bore, 88.9 mm stroke and a compression ratio of approximately 8.93:1. Engine specifications are given in Table 3.1.
The engine can be operated to a maximum speed of 90 rps developing 15kW power. In this study, however limitations on the modified combustion cham• ber materials resulted in the operation in mid range only.
The standard bathtub cylinder head and flat piston configuration for gaso• line (or gaseous fuel) operation, cross section is shown in Figure 3.2.
The engine was coupled to an electrical DC trunnion mounted dynamome• ter which uses a regenerative load system and was operated through a KTK thyristor converter unit. The dynamometer's output, capable of absorbing
44 kw of engine power up to 5500 rpm was fed into the three phase 400 volt mains supply. The McClure dynamometer was also used to motor the engine where the power is drawn from the AC supply through the thyristors.
The engine ignition system used a conventional coil and spark plug ar• rangement with the primary coil circuit operated by a 'Lumenition' unit picking up speed and TDC reference from the flywheel. This system enabled manual control of the ignition timing from the control consol. A standard champion N6Y spark plug with 0.6mm gap was used.
The cooling and lubrication oil systems are mounted within the engine bed. These comprised a closed circuit pressurised coolant system of a water and antifreeze rust inhibitor mixture; oil and coolant circulation pumps; an Chapter 3. Experimental Apparatus and Method
integral oil filter and oil and coolant heat exchanger, oil and coolant tempera•
ture control valves, and mains water bypass valve. Temperature control, and
temperature maintenance during motored tests, was provided through sepa•
rate oil and coolant heaters with Spirac Sarco limit sensors set in the range
60-85 °C. Low level and low pressure sensors in the coolant and oil system
respectively, and thermocouples in each system, provide safety overheating
trips together with temperature indication at the control consol.
Fuel supply for operation with natural gas were provided by regulated flow
from BC Hydro mains, with needle valve control in the control room. Flow
measurement was provided by a Miriam laminar flow (50MW20-1.5) with
a differential pressure transducer and signal demodulator mounted between
the regulator and needle valve. The pressure drop across the laminar flow
element was also sensed by an inclined differential manometer in the control
room.
The air intake system consisted of an air filter, a 1 kW heater, throttle
body assembly with servo motor controlled throttle and inlet manifold. Sim•
ilar to fuel flow, a Miriam laminar flow element (50MC2-4F) mounted ahead
of the intake filter was used for flow measurement.
3.2.3 Combustion Chambers
The standard gasoline Ricardo Hydra engine configuration has a bathtub com•
bustion chamber located in the cylinder head and uses a flat piston with solid
skirt, two compression rings and one oil control ring. Six different combustion
chamber configurations were obtained by using the standard bathtub cylin•
der head with three piston shapes and a flat cylinder head with an additional
three piston shapes. Chapter 3. Experimental Apparatus and Method
These configurations were separated into two groups. The 'bathtub' group
consisted of the standard bathtub cylinder head with a flat piston configura•
tion and two modified pistons termed 'single slot' and 'castellated' pistons.
The 'disc' group comprised the flat cylinder head and non-squish disc piston
configuration and two types of bowl-in-piston pistons termed 'bowl-in-piston'
and 'squish jet' pistons. These configurations are shown in Figures 3.3 to 3.5.
Development and manufacture of the flat aluminium cylinder head with
extended pistons was carried out by R. Dymala-Dolesky and is detailed in
his MASc report [18]. To allow for a more central position of the spark plug,
the valve position was moved from the centre of the chamber. Exhaust port
and cooling manifolds were redirected and a smaller 12mm spark plug was
specified.
The engine rebuild for this series of configurations required several modi•
fications:
• Installation of a longer cylinder liner.
• Installation of a modified connecting rod.
• Modifications to the cylinder block and timing drive system.
• Addition of packing plates under the cylinder block.
Pistons
The extended pistons to accommodate the combustion chamber bowl were
cast from aluminium alloy A356, heat treated to T6, this being the same
material as that used in the flat cylinder head. The new bathtub type pistons
were manufactured from the pattern of the extended bowl type pistons. Sand
casting and heat treatment were carried out by a local foundry, then the piston Chapter 3. Experimental Apparatus and Method
blanks were machined in the machine shop of the Mechanical Engineering
Department. Consequently these pistons were heavier than the original flat
piston used in the bathtub configuration.
The limitations of the sand cast piston material and process compared
to a more standard forged or cast process with fast cooling in a permanent
mould, resulted in a reduction of material strength estimated to be 1/3 to 1/2
that of the original. This and the hybrid nature of the bathtub type piston
were considered in the choice of maximum operating speeds.
The modified bathtub pistons were designed to have the same compression
ratio as the bathtub with flat piston and the bowl in piston configuration.
Liquid displacement volumetric checks on the inverted cylinder head and
appropriate pistons with the valves in position and a pressure transducer
blank inserted were carried out using a thin oil. Within the accuracy of this
technique the combustion chambers were found to have the same clearance
volume.
3.2.4 Modified bathtub pistons
The desired effect of the modified pistons was to block off the plain squish
action occurring near TDC, and direct the flow through one or more narrow
slots to form jet action.
In his evaluation of the jet action produced by a new 'squish jet' piston
Dymala-Dolesky measured a weak jet effect in the piston bowl. He therefore
recommended the closure of the squish area to force flow through the jet
forming channels.
In this study it was proposed to create the forced jet action through slots
in a raised wall on the piston surface, with the intention that these slots would Chapter 3. Experimental Apparatus and Method
form channels as the wall entered the cylinder head cavity. The most practical
design configuration within the limits of the equipment available was the use
of the standard bathtub cylinder head and a modified piston.
The modified pistons were designed to have the same compression ratio
and be capable of operating over the same range of speed and loads as the
standard bathtub and bowl-in-piston chambers. A 5mm wall thickness was
used equal to the minimum thickness of the cylinder head. The combus•
tion chamber volume was maintained by the addition of a shallow bowl to
compensate for the raised wall.
To produce the maximum blockage effect at the earliest point before TDC
the maximum wall height possible without fouling of the intake or exhaust
valves was desired. The piston position per crank position was determined
by:
2 Se = r/2 * (1 + cosO) + CL*(1- - (\sin6) )
where Sg is the distance from BDC, X = r/2*CL,rie the stroke and CL
is the con rod length. The position of the valves were determined from their
cam profiles and a maximum wall height of 12 mm chosen to give a minimum
of 1 mm clearance. This clearance was verified by measurements in the actual
configuration. The 12 mm wall was calculated to produced a blocking effect
after 36 degrees BTDC.
Squish area
A final note on the combustion chamber designs deals with the difference
in squish area between the bathtub group of pistons and the bowl-in-piston
types. Squish area is determined from the percentage of top squish surface Chapter 3. Experimental Apparatus and Method
area of the piston bore: 7r * £)2/4. For the simple bowl-in- piston this is
determined by: D2 - d2 %SAREA = D2 * 100
where D is the piston bore and d is the bowl diameter. The bowl-in-piston
pistons used in this study had a squish area of 70%. From similar calculations
the bathtub type piston squish area was approximately 27%-29% allowing for
the spark entry point. The effect of this difference in squish area is discussed
in Chapter 5.
3.2.5 Instrumentation
A schematic of the instrumentation layout is shown in Figure 3.6.
Pressure measurements in the combustion chamber were made with a
Kistler 6121 piezo-electric pressure transducer mounted within an extended
sleeve in the cylinder head of the engine. The charge output signal was fed
to a Kistler model 5004 charge amplifier to yield voltage data proportional to
the cylinder pressure.
The pressure data were digitalized at a rate of 1 sample per 0.2 crank angle
degree for the motored tests and at 1 sample per crank angle degree for the
majority of the fired engine tests. The pressure signal was also displayed on
a Textronic oscilloscope during operation. The transducer was recessed from
the combustion chamber by half its diameter to protect it from the effects of
thermal shock, as recommended by Brown, Benson and Pick [71, 72]. The
selection, mounting and operation of the pressure transducer and amplifier
system used in the Ricardo test cell were consistent with the 'ideal transducer
specifications and recommended operation' of earlier investigators reviewed Chapter 3. Experimental Apparatus and Method
in Chapter 2. The transducer specification and calibration curve are given in
Appendix A.
Hotwire voltage measurements were taken with a TSI 1226 high tempera•
ture probe, suitable for high temperature environments, in conjunction with
a DISA M-10 anemometer bridge circuit in the constant temperature mode.
The signal was low pass filtered at 20 KHz with a DISA 55D26 signal condi•
tioner before being digitalized at a rate of 1 sample per crank angle degree.
Hot wire sensor
The sensor used a Platinum-Iridium wire, 6.3 micrometers in diameter and 1.5
mm in length. The probe was inserted in the cylinder head through the spark
plug hole using a two part specially machined swaged fitting sealing around
the probe shaft. Two adapter fittings were required for the two different
spark plug sizes used in the cylinder head. The sensor wire was positioned
approximately 2 mm below the spark point as shown in Figure 3.7. The
anemometer system specifications are given in Appendix B.
An AVL model 360c/600 optical pickup, mounted on the crank shaft was
used to generate clock pulses every 0.2 degrees of crank angle and a trigger
signal at BDC for the data acquisition circuitry.
Volumetric air and fuel measurements were obtained from the differential
pressure sensors across their respective laminar flow elements, displayed on
inclined differential manometers in inches of water gauge. The manufacturers
calibration curves and constants at standard normal conditions are given in
Appendix A. The airflow meter was calibrated and compensated for pulsating
flows.
Engine speed was measured by a tachometer attached to the dynamometer Chapter 3. Experimental Apparatus and Method
shaft and controlled by a set-speed potentionmeter on the control consol. A
difference in the input setting and true engine speed sensed in the converter
cabinet resulting in appropriate correcting control signals sent to the thyristor
triggering circuits. The engine speed was therefore maintained at a constant,
± 0.2%, irrespective of the operator adjustments to throttle settings, fuel/air
ratio or spark advance.
Torque measurements were made by strain gauge load cells mounted on
the torque arm. The output of these cells was provided in continuous display
at the control consol. Static calibration of the torque measuring system was
carried out using a 20 Nm calibration weight provided in the dynamometer
pedestal. These calibration tests are discussed in the method Chapter 4.
Ignition control and measurement were provided by a multiturn dial with
a range of 70 degrees BTDC to 20 degrees ATDC, sending a set point signal
to the Lumenition electronic ignition unit.
Additional pressure transducers and thermocouples mounted on the en•
gine rig provided operating condition information and safety trips for the air
intake and exhaust system, and the engine coolant and oil modules.
Control over the engine motoring and fired conditions was carried out by
the Cussons electronic control system. Control over the acquisition process
was carried out by the IBM PC, using signals from a majority of the above
instruments fed through to the control room via a milspec connector.
3.2.6 Data Acquisition
Two data acquisition systems were employed for this study. A slow data
translation system was used primarily for engine monitoring and performance Chapter 3. Experimental Apparatus and Method 37
information, while a fast, ISAAC 2000 acquisition system was used for cylin•
der pressure, BDC signal and Hotwire voltage data collection.
The 'slow' data translation system operated at 27.5 KHz and collected
signals from the engine instrument analogue transducers, eg; airflow, fuel flow,
speed, torque, and ignition advance. These signals were optically isolated
and low-pass filtered at 60 Hz before passing the AD converter inside the
PC. The hardware arrangement detailed in Figure 3.8 shows the circuit box
containing the screen terminal boards DT752 and DT709 which connect to
the data translation DT2801 analogue to digital converter board installed in
the PC.
The 'fast' acquisition system consisted of an ISAAC 2000 unit with 64 K
of buffer memory, off-line block transfer of data and a flexible control system.
Sampling rates of 200 kHz were available on four channels, using labsoft II
software and the IBM PC.
Data acquisition programs
The data acquisition programs used in conjunction with the test cell instru•
mentation system were initially developed for the Alternatives Fuel Labora•
tory by A.Jones. Further details of the system circuitry and programs may
be found in the operations manual [69].
The slow data acquisition system was run using the 'DATAQ' program
which scans the channels of the DT2801 board for sensor information. One
hundred values were read in from each channel and averaged. The signal volt•
ages were then converted to the appropriate units and calculations made for
engine performance parameters; Brake Power, Brake Mean Effective Pressure
(BMEP), Thermal efficiency, Brake Specific Fuel Consumption and relative Chapter 3. Experimental Apparatus and Method
air fuel ratio. These data were displayed to the PC screen and updated every
5 seconds providing operating control information. The original programs
were modified to provide ignition advance display also. Collected data for a
desired set point were stored on disc drive B at a rate of 400 samples per
channel. Similar conversion calculations and performance calculations are
made on this data using the 'CRUNCH' program.
Data from the fast acquisition system were obtained using an interac•
tive program, (a modified version of 'HOTWIRE' [69]), allowing a variable
number of channels, acquisition speeds and number of cycles to be obtained.
Consecutive or nonconsecutive acquisition was also controlled integrally.
Data trigger signal
On initiation of the system at the desired test conditions, the phasing of the
acquisition was controlled by trigger BDC and clock pulses from the DT2801A
board. The ISAAC unit was triggered to start acquisition from the 'next'
BDC signal after an expansion stroke. The data stored by the ISAAC were
transferred to the IBM PC prior to storage on floppy discs.
The data collected on disc were then transferred to the Department's mini•
computer, a VAX 11/750 unit using an 'Ethernet' communication package.
3.3 Motored Engine Tests
3.3.1 Introduction
Flow measurement tests were conducted in the engine, while motored by
the DC dynamometer for each chamber configuration at the conditions given
in Table 3.2. Although most of the flow measurements were taken at three Chapter 3. Experimental Apparatus and Method
different speeds for the wide open throttle condition (WOT), gaps in the data
set were the result of unstable operating conditions and associated difficulties
with the measuring equipment.
Hot wire anemometry techniques were used to yield velocity and tur•
bulence information for the flow near the spark plug location. Digitalized
hotwire voltages and cylinder pressure acquired at the rate of one sample per
0.2 crank angle degree, were recorded for 44 nonconsecutive cycles using the
fast acquisition system and 'HOTWIRE' program. During the data acquisi•
tion procedure, engine operating conditions such as airflow rate and ambient
pressure and temperature conditions were noted.
The repeatable nature of the motored pressure trace made it possible to
record single channel data on each floppy disc, thus the hotwire and pressure
measurements were taken from two separate sets of 44 cycles with the engine
operating at the same conditions. The motored pressure data required for
the turbulence analysis were also used as a powerful check on the phasing of
the trigger signal and acquisition and the operation of the engine.
Operational procedures used in running the engine, specific to the mea•
surements made are herewith described followed by details of the pressure
and hotwire measurement methods.
3.3.2 Operational Procedures
The extent of the engine changes required for each combustion chamber
configuration, and the subsequent extended period over which the experi•
ments were made, emphasized the need for standard operation procedures
and checks to minimize or eliminate extraneous engine variables. The impor•
tance of maintaining accurate calibration of the engine instrumentation, used Chapter 3. Experimental Apparatus and Method
in establishing the test condition, to obtain a reliable set of data was also
recognized.
Engine rebuilds were carried out in accordance with the manufacturers'
recommendations [70]. These changes included fitting the appropriate piston,
conrod, cylinder liner, spaces and cylinder head assembly using piston rings
and head gasket reserved for each, checking of the belt drives, valve timing,
and oil and flushing of the coolant system.
The newly machined pistons, modified bathtub and replacement bowl-in-
piston, were fitted with new piston rings and run in as detailed in the method
section 3.4.
All motored tests were run at wide open throttle indicated by a leveling off
of the maximum air intake flow. This was confirmed by a visual check of the
throttle butterfly valve position. It was particularly important to reset the
manual control of the throttle position for maximum gain when the engine
was rebuilt from the diesel configuration with injector rack control.
Prior to any measurements, the system was powered for a minimum of one
hour to warm up the electronic circuits and then zeroed. The water heater
and Senco limit valve 6ensor were used to maintain the coolant temperature in
the range 65-80 °C during the motored tests. Initial problems encountered in
reaching this temperature resulted in a rectified valve position and resetting
of the Senco limits. Manual override was also used to ensure correct operating
temperature. When using previously fired pistons, operation at low load and
speed could also be used to heat the system. The standard start up procedure
was then used with preheat of the lubrication oil to 40 °C, partly open throttle
and low speed settings. Chapter 3. Experimental Apparatus and Method
Static calibration checks performed on the dynamometer torque arm sys•
tem, using the 20 Nm calibration weight, rezeroing and resetting the gain
as required, gave an acceptable error of 1%. Examination of the dynamic
behaviour however, revealed a possible 'sticktion' problem in the trunnion
bearings. Care was taken to repeatedly check the torque readings during
testing for cold and hot operation, allowing settling time, or jiggling, prior to
zeroing. This resulted in an non elimitable error of ± 0.6 Nm.
3.3.3 Pressure Measurements
Motored pressure measurements were taken primarily to provide operating
pressure and temperature data for calculation of the gas properties used in
the analysis of the hotwire signal. The sampling rate of 0.2 crank angle degrees
was therefore chosen to be compatible with the hotwire rate of acquisition.
For each test speed 44 nonconsecutive cycles of data were taken. Each cycle
was triggered from BDC covering the 720 degrees of the exhaust, intake,
compression and expansion strokes. The fast acquisition system was used
with the program 'HOTWIRE' and the data stored on floppy discs.
Static calibration of the pressure transducer and amplifier system, using
a dead weight tester over a limited range, was carried out prior to the engine
tests, confirming the manufacturers' calibration constants. Calibration curves
for the pressure transducers and testing details are given in Appendix A.
The amplifier was operated with the sensitivity, mechanical units to volt•
age dial, set to the manufactures calibration constant, gain set to 5.0 and on
medium response. The pressure trace was also observed on an oscilloscope
and monitored for correct transducer operation, eg., overload error evidenced
by a faulty trace. During the course of the engine tests damage from engine Chapter 3. Experimental Apparatus and Method
seizures and faulty pressure transducers led to extensive checking of the BDC
signal, phasing and transducer units.
3.3.4 Hotwire Measurements
The hotwire measurement process involved three stages;
1. Preparation and welding of the probe sensor wire.
2. Calibration of the probe and anemometer bridge system.
3. Insertion in the engine and fast data acquisition.
The harsh operating environment in the motored engine combustion cham•
ber resulted in frequent breakage of the probe sensor wire. The probe was
subjected to rapidly changing pressures and high temperatures, stray oil and
residue particles, engine vibration and a turbulent flow field. Unfortunately
no wire was able to survive the vibration on shutdown or subsequent removal
of the probe, frequently wire breakage also occurred during speed changes
from test point to point or during the test.
Preparation
The sensor wire was prepared by spot welding the 6.3 micrometer Platinum
Iridium wire to the probe support needles using a DISA 55A12 welding unit
and microscope system. After welding and visual inspection for a clean weld,
the sensor was annealed, through the anemometer bridge circuit at the oper•
ating temperature of 600 degrees Celsius for 6-8 hours until a stable resistance
was recorded. The 600 degree overheat temperature was compatible with rec•
ommendations of Witze [30] and the limitations of the platinum iridium wire
reviewed by Vines [73]. Chapter 3. Experimental Apparatus and Method
The resistance of the wire and compensation for the probe and cable was
determined using the bridge circuit and the operating resistance of the system
obtained from:
Rop — -Ramb(l + "(Top — Tamb))
where a was the thermal coefficient of resistance is specified by the manufac•
turer.
Calibration
Calibration of each wire sensor, with the anemometer bridge and low pass
filter, was performed against a pitot tube in a small wind tunnel, range 0.5-
16 m/s, at ambient temperature and pressure. The hotwire system registers
change in the heat transfer from the sensor due to the cooling effect of the
flow. Being a function of the temperature differential between the sensor
and the fluid flow, this heat transfer process is sensitive to the operating
temperature and pressure and gas properties of the flow. Ideally calibration
at all operating pressures and temperatures should be made, this was however
impractical.
An analytical model involving a Nusselt-Reynolds relation was therefore
used to obtain empirical calibration coefficients of the form:
Nu = A + B * Ren
This approach enabled extension of the calibration data to the operating
conditions. The techniques and theory behind this method are described in
Chapter 4. Detailed information on the welding and calibration procedures,
and a representative calibration data set and curve are given in Appendix B. Chapter 3. Experimental Apparatus and Method 44
To minimize the probe time in the engine and reduce possible breakage
occurrences, the engine was first warmed up and stabilized at the test con•
dition. On shut down the probe was inserted and the engine returned to
the set condition as smoothly as possible. Calibration and operation of the
anemometer was carried out with a 20 kHz low pass frequency filter to pre•
vent aliasing of the signal. The acquisition rate was dependent on the engine
speed ranged from 36 kHz to 90 kHz.
3.4 Fired Engine Tests
3.4.1 Introduction
Performance and combustion pressure measurements were conducted in the
fired engine fueled by natural gas, at the conditions and configurations given
in Tables 3.3 and 3.4. The highest operating speed for the pressure measure•
ments was reduced from 66.7 rps to 50.0 rps, after siezure of the bowl-in-piston
piston. An examination of the engine indicated this was probably due to a
combination of the weaker materials in the non-standard piston and failure
of the water cooling system. Gaps in the data set for part load conditions
were due to the difficulty of maintaining stable engine operation at those
conditions.
Engine performance parameters were also taken using the slow data ac•
quisition system over these ranges. All tests were run at optimum ignition,
minimum spark advance for best torque, MBT, determined by the operator.
The fired tests were performed directly after the motored engine tests
without altering the configuration. New pistons were the exception to this Chapter 3. Experimental Apparatus and Method
procedure, where the cylinder head was removed and the piston surface in• spected for damage after the run in tests.
3.4.2 Operational Procedures
In this section the fired operational procedures are described followed by details of the performance and pressure measurements. System checks and motored operation were carried out in similar manner to the motored tests and the recommended start up procedure at low load, partial open throttle employed. The ignition was initially set to 25 degrees BTDC and the gas flow controlled via the needle valve in the control room. The desired operating condition was then reached by slowly increasing the speed and coordinating fuel and throttle openings. Part load conditions were achieved by reducing the throttle and fuel in tandem. Stable engine conditions at the desired load and MBT, with coolant and lubricating oil held between 70-90 °C were maintained prior to any measurements. The new pistons, ie; replacement bowl-in-piston and single slot modified bathtub, were run in under a combined part and full load schedule to a max• imum speed of 70 rps over 20 hours. The procedure followed Ricardo man• ufacturer recommendations and previous piston experience, to ensure proper seating and sealing of the newrings [70] . The slow data acquisition system previously detailed, was used to moni• tor engine operation with intermittent checks on the torque calibration and updating of the ambient conditions made during the test process. Chapter 3. Experimental Apparatus and Method 46
3.4.3 Performance Measurements
At each test point, after stable operation for a minimum of ten minutes, en•
gine operating data, processed via the data translation board from instrument
analogue signals, were taken using the slow acquisition system. Similar perfor•
mance parameters to those displayed on the PC monitor, were then available
using the 'CRUNCH' program on the stored converted signals which had been
averaged over 400 samples. The performance parameters stored were, speed,
Brake power, BMEP, Bsfc, thermal efficiency, torque, ignition advance, air•
flow rate, natural gas flow rate, relative air fuel ratio and ambient pressure
and temperature data.
3.4.4 Pressure Measurements
The fired engine cylinder pressure data obtained at each test point was taken
for use in combustion measurement and cycle-to-cycle analysis. For this rea•
son consecutive cycle acquisition was desired.
The pressure transducer and charge amplifier system of the motored tests
were used with the amplifier gain set to 10.0 to prevent overloading of the
acquisition system. The analogue signal from the amplifier was digitalized
at a rate of one sample per crank angle degree. During consecutive cycle
acquisition the Bottom Dead Centre (BDC) signal was used to trigger the
fast acquisition at the start of the first data set only. Subsequently the BDC
signal was also recorded on a second channel for use as a phasing check on
the pressure signal.
The pressure signal processing method provided a check on the position
of BDC, relative to the pressure signal, and rejected out of phase cycles. For Chapter 3. Experimental Apparatus and Method
data acquired per crank angle degree this resulted in a maximum of 3 cycles
being rejected. However it was found that data acquired per 0.2 crank angle
degree were liable to major phasing errors, therefore only nonconsecutive data
were taken at this rate.
The original 'HOTWIRE' program was modified for the acquisition of
consecutive data and expanded for 200 cycles. These were stored on two
floppy discs per test condition.
Ambient pressure and temperature and the air and fuel flow rates from the
differential pressure manometers were also recorded for use in the analysis.
Extensive data were obtained from each phase of the tests for the six
chamber configurations over a period of 20 months. The schedule for this
stage of the investigation was greatly extended due in part to the difficulties
encounted with probe breakage but also to the range of operating conditions
studied for each configuration. This however, allows for comparison with
studies over a broad range. The techniques used to analyse the data are
discussed in Chapter 4. Chapter 4
Data Analysis
4.1 Introduction
The techniques used in the analysis of the collected data have been divided into two areas, first those used in the processing of the raw analogue signals to render raw velocity and pressure information, and second those used in the interpretation of the raw data through the selection of characterizing parameters. The method used on the motored engine test data to provide data acquisition phasing checks and flow field information is first detailed followed by the fired engine test data analysis methods used on both the
'fast' and 'slow' aquired performance data. While the selection of the various techniques and choice of input parameters of previous research studies, have been primarily covered in the introduction, some further general comments are included in this chapter.
4.2 Motored Engine Tests
4.2.1 Analytical Procedure
The sequence of events required in the analysis of the motored engine test data is given in Table 4.1. The programs and methods used for this por• tion of the study were based on similar studies of turbulence measurement in
48 Chapter 4. Data Analysis
spark ignition engines and a rapid compression machine carried out by grad•
uate students for the Alternatives Fuels Laboratory Group [17, 18, 74, 75].
Modifications to the base programs have been made for successive experi•
ments, specific to the equipment dimensions and experimental variables, in
this investigation for example, in the acquisition rate and number of cycles
acquired.
The calculations were carried out on a VAX 11/750 following transfer of
the motored pressure and hotwire data from the PC stored floppy discs. Pro•
cessing of the pressure analogue signal was carried out in the PRESS-ANAL
program, after a conversion program, ISAAC2VAX, was used to condense
the individual cycle data into one file and strip the header information. A
smoothing-filtering program was then used on the second half of the data
set prior to calculation of the cylinder temperature, the area of interest being
the compression and expansion strokes. The ensembled cylinder pressure and
temperature trace data were then used in HW-ANAL with the raw hotwire
data, also reduced, and the calibration constants from HW-Cal, to yield
raw velocities. Flow field analysis parameters were then calculated in the
TURBULENCE program.
An additional program was used for logarithmic plots of pressure versus
volume to check phasing of the pressure signal data with the cylinder volume
assignment, using the crank angle position from the optical pickup.
4.2.2 Pressure Signal Processing
The digital data analogue signal was converted to a relative pressure by a
scaling factor, a function of the gain and charge amplifier setting, ie., me•
chanical unifs/volts. Following recommendations by Lancaster et al [76] the Chapter 4. Data Analysis
cylinder pressure at BDC after the intake stroke was then equated to a refer•
ence condition, a simple equivalent to the pressure in the intake manifold, to
obtain the absolute cylinder pressure. The average manifold intake pressure
was assumed equal to the ambient pressure times the volumetric efficiency.
PBDC = Pamb * f]v
The pressure in bars was determined by:
*- = <^-io)...o
where the analogue signal is digitalized 0-4096 for -10 to +10 volts with the
gain set to 5.0 for the motored tests.
The volumetric efficiency for a four stroke engine is defined as the ratio of
the actual mass of air supplied per cycle to the theoretical mass of air required
at standard conditions ( 15 °C, 101 kPa). Volumetric efficiency varied with
combustion chamber and speed, and was calculated using measured flow rates
by: 2 * Qair Vv = V. * RPS
3 where QaiT is the total amount of air m /sec supplied and Vt is the volume
swept.
The scaled pressure data were ensemble averaged over the 44 nonconsec-
utive cycles prior to use in the temperature calculations. Information on
individual pressure traces was used in a separate program for phase checking.
The importance of accurate total volume calculations with pressure phasing
is highlighted in its use in the assessment of the work inventory of an cy•
cle, ie: IMEP calculation. Logarithmic plots of pressure and volume provide
information on the phasing accuracy. Chapter 4. Data Analysis
The main inaccuracies in pressure assignment and their evidence may be
summarized
• Improper reference pressure, exhibited as curvature in the compression
stroke, which is a scaling problem only.
• Improper clearance volume, exhibited as curvature in the end of the
compression stroke.
• Improper phasing, exhibited by the crossing of the compression and ex•
pansion slopes, this may also indicate a faulty transducer.
Extensive checks were carried out on the pressure transducer system used
in this work by the author and concurrent users of the equipment [77], re•
sulting in the removal and replacement of suspect transducers and data. The
occurrence of peak pressure in the motored traces were found to be within
2 degrees of TDC, the discrepancy being due to irreversibilities due to heat
transfer, confirming correct phasing.
Maintenance of the same clearance volume, and subsequent compression
ratio, in the different combustion chambers, checked by liquid displacement
measurements of the cavities, was confirmed by the similarity of the motored
pressure traces. It was noted, however, other effects such as varying heat
loss to the cylinder wall and head, can cause differences in the traces. A
comparison of the motored pressure data was also useful in assessing the
reliability and repeatability of tests carried out over an extended period.
A partial smoothing-filtering program was used on the reduced ensem•
bled pressure data, in the initial region, BDC to 60 degrees BTDC, prior
to temperature calculation. Details of the filtering method used, and the op•
tions examined, are given in Appendix D. The smoothing was done to remove Chapter 4. Data Analysis 52
transient effects in the region of inlet valve closing, possibly due to vibration,
without affecting the critical region near TDC.
The gas cylinder temperature was determined indirectly through its poly-
tropic relation with pressure. Previous investigations [74] have examined
various methods assuming:
• Adiabatic compression and expansion.
• Polytropic compression and expansion using calculated values of fc and
• Adiabatic compression to IVC, application of the perfect gas law prior
to EVO and adiabatic expansion after EVO.
Witze [30] showed that adiabatic computation was appropriate during
the compression stroke and increasingly poor after 50 degrees after TDC,
through comparison with actual measurements. The last assumption, using
the ambient temperature as the reference temperature at intake BDC, was
taken in this study, and the temperature determined by:
P(i) ^ intake BDC to IVC T{i) = Tamb * PB DC
P(t) V IVC to EVO 7/(0 = T{i - 1) * p. * ^ P(i-1) V(i-1)
P(i) ^ EVO to expansion BDC T(i) = T(i - 1) * v The polytropic compression and expansion coefficients differ from isen-
tropic for air for a variety of reasons. The compression coefficient (7,.) is less
than isentropic due to residual gas and increased temperature. The expan•
sion coefficient (7e) is less than isentropic due to residual reactions and higher
temperature. Chapter 4. Data Analysis
In the motored engine residual gas reactions are not in question, how•
ever the polytropic coefficient does vary with engine speed and combustion
chamber. The isentropic coefficient for air at standard conditions was 1.365.
In trial runs 7C and 7e were both set to 1.35 consistent with previous work
by Dymala-Dolesky [18] and current bathtub chamber work by Kapil [77] who
used values of; 7C: 1.35 to 1.352 and 7e: 1.38. In the final analysis the motored
pressure compression coefficient was calculated from the logarithmic slope of
the pressure from the fired cases, averaged over speed and stoichiometric ratio
for each configuration. The expansion coefficient of 1.35 was used for all cases.
The compression values ranged from 1.27 to 1.37.
4.2.3 Anemometer Signal Processing
The analogue hotwire signal registered change in the electrical resistance due
to the heat transfer from the wire sensor due to the cooling effect of the flow.
This heat transfer is a function of the temperature differential between the
sensor and the fluid flow and is sensitive to the gas properties, themselves de•
pendent on the pressure and temperature in the measurement regime. Hence
it was possible to determine the gas velocity at the point of measurement by
applying an analytical correction to the anemometer output for the separately
determined gas properties at the particular operating conditions.
An empirical Nusselt-Reynolds relationship formed from the calibration
data has been used. An outline of the techniques and theory involved in this
analysis follows.
Heat transfer from the wire occurs through radiation, bouyant and forced
convection and conduction along the wire to the supports. The equations of
Lancaster [13] have been used in this analysis. These were based on work Chapter 4. Data Analysis
by Collis and Williams [78] and Davies and Fisher [79] on heat transfer from
electrically heated cylinders applying modifications to Kings Law, which re•
lates the Nusselt number to a Reynolds number based on the wire diameter
and gas properties:
Nu — a + bRei
These studies made the assumptions that, conduction to the supports may
be neglected for sufficiently long wires, ie., with length to diameter ratios in
excess of 200, radiation may be neglected and that the convective bouyancy
was important in low speed flows only.
The early methods solved the one dimensional heat balance equation to
obtain an expression for the mean wire temperature with the convective heat
transfer coefficient, h, solved iteratively. In a different approach Lancaster
used a ID energy equation involving the gas temperature, resistance of the
wire at zero condition and the current, for the heat distribution along the
wire. This equation was then integrated to find an expression for the mean
temperature with h unknown. The coefficient h, found by an iterative method
was used to evaluate the Nusselt number. Applying the calibration coefficients
to the analytical model the Reynolds number was obtained and hence raw
velocity data were calculated.
The selection and calculation of the free stream gas temperature for the gas
property determination, described in the pressure section 4.2.2, was suggested
by Witze [30] as giving the best agreement with LDA studies. Chapter 4. Data Analysis
4.2.4 Flow Field Data Analysis
Returning to the definition of turbulence as a "fluctuating velocity component
superimposed on the bulk velocity of a flow" [24], the means of defining the
turbulent intensity, a characterizing parameter of the turbulent flow field, is
subject to debate. Many different definitions have been used in both HWA
and LDA analysis [13, 22, 35, 80, 81]. Generally these methods may be divided
into ensemble averaging methods and cycle resolved methods.
The limitations of ensemble averaging a nonstationary process, where the
ergodic definition of the ensemble average being equal to the time average
does not hold, is evident in the analysis of engine turbulence. The periodic
nature of the mean flow causes cyclic variations in the mean to be included
in the estimation of turbulence. In contrast the nonstationary description of
Lancaster [13] divides the instantaneous velocity into a mean, time averaged
mean velocity and a turbulent component.
A cycle-by-cycle nonstationary time averaging method developed by Cata•
nia and Mittica [82] was used to extract turbulence data without including
the cyclic fluctuations of the mean flow. This method involved the ensemble
time averaging of the raw velocity in a defined period, or window size, and
the fitting of a smoothed velocity curve similar to the method of Rask [80].
The following outline describes the procedure employed in the turbulence
analysis and the velocity component definition.
The separation of a mean velocity per cycle was obtained by evaluating
the mean from the instantaneous velocity in a given window by trapezoidal
integration. A cubic spline curve fitting routine was applied to the mean value
at the center of each window and then interpolated. These traces were then Chapter 4. Data Analysis
ensemble averaged over the number of cycles to give a representative mean.
The average at each window midpoint was defined by:
1
U(tw,i)=- U(t,i)dt
Cubic spline curve fitting and interpolation yielded: Uw{t,i) and the window
ensembled mean calculated from:
iy t=i
The fluctuating component of the velocity was defined by:
uw(t,i) = U(t,i) - Uw(t,i)
The squared average fluctuation at each window midpoint was then evaluated
by:
2 [«(*-,«)]' = 4 r? \U{t,i) - Uw(t,i)] dt
These were ensemble averaged over N cycles and the square root taken:
1 N u(tw)
Cubic spline curve fitting and interpolation was again applied to yield: uw{t)
a representative rms intensity.
Ensemble averaging was also carried out for comparison, where the en•
semble mean is defined by:
uE(t) = -Ni:u(t,i)
The fluctuating component was separated by:
uE{t,i) = U(t,i)-UE{t) Chapter 4. Data Analysis
and the rms intensity defined by:
Mt) = ^^Yt\u(t,i)-uE(t)i)
In addition the relative turbulent intensities, referenced to the mean flow
and representative kinetic energy terms were calculated.
Scaling of point data with the mean piston speed defined by: Sp = 2rN
has been used as a quantifying characteristic in some studies. However Lan•
caster cautions that "quantitative interpretation of these single point mea•
surements is unwarranted" [13]. This calculation has been performed for
comparative purposes only.
In selection of the window size to separate the 'turbulent' fluctuating
velocity component from the mean and cyclic variations, a balance must be
maintained. Too small a window size may result in the loss of low frequency
turbulence, ie., having been interpreted as mean flow, too large a window
size and the mean flow .motion may be attributed to turbulence. Catania
and Mittica [82] showed that a window size in the range 4 to 12 crank angle
degrees had little effect on the mean, while the 8 to 20 range was insensitive
for the intensity, giving consistent values.
A 12 degree window was used for this work consistent with previous work
by Dymala-Dolesky [18] and supported by a sensitivity check. Figure 4.1
shows the effect of a range of window sizes on the turbulence intensity. Win•
dow sizes between 4 and 18 degrees had a negligible effect on the mean while
the window sizes between 8 and 15 degrees gave the most consistent value of
turbulent intensity. Chapter 4. Data Analysis
4.3 Fired Engine Tests
4.3.1 Analytical Procedure
Data from the two acquisition systems were analysed to yield engine per•
formance information. The primary object of this work was to obtain in•
formation of the effect of the various combustion chamber shapes on cyclic
variations. This is one of the main limiting factors to the operation of an en•
gine at the extended limits of lean mixtures, particularly with slow burning
fuels such as natural gas.
Analysis of the fired pressure data was made to provide a characteristic
measure of this variability, ie; COV of IMEP, and to provide a comparison
with engine performance, ie; fuel consumption and efficiency. The majority of
the analysis was performed on the fired cylinder pressure data; a fundamental
engine variable.
The sequence of events followed in the analysis are shown in Table 4.2.
After processing of the analogue pressure signal in a similar manner to that
used in the motored engine tests, characterizing parameters were calculated,
eg; IMEP and peak pressure, these were ensembled averaged and statistical
analysis carried out to quantify the cyclic variations. Combustion, an inter•
mediate phenomenon, was also investigated through a mass fraction burned
analysis. Details of the techniques used, and the choice of the characterizing
parameters are described in the following sections.
The analytical procedure has been illustrated for the data acquired over
200 consecutive cycles at a rate of one sample per crank angle degree. The
data obtained at the rate of one sample per 0.2 crank angle degree for 44
cycles were reduced to the lesser rate and minor modifications made to the Chapter 4. Data Analysis
program.
4.3.2 Pressure Signal Processing
The method for converting the digital analogue signal to absolute pressure
data was similar to the motored pressure process using a gain factor of 10.0.
The reference pressure was again taken as the pressure in the intake manifold
at BDC calculated by:
T PBDC = Pamb *Vv* ~^ amb
_ 2 * Qtotal Vv ~ V. * RPS
3 where Qtotal was the total amount of air and fuel in m /sec supplied and Vt
was the volume swept.
A smoothing routine similar to that for the motored pressure data was
applied to the data in the inlet valve closing region up to 30 degrees be•
fore spark, transient spikes in this region causing errors in the mass fraction
burned analysis. This smoothing routine was performed within the FIRERUN
program.
The individual cycle pressure data with volume assignment was then used
to calculate performance parameters. Ensemble averaging over the total num•
ber of cycles also provided a representative fired pressure trace.
4.3.3 Performance Data Analysis
Two measurements of performance data taken: on line monitoring through
the slow acquisition system, and cylinder pressure data via the fast acquisition
system. The on-line monitoring was used primarily as an indication of the Chapter 4. Data Analysis
engine operating efficiency and stability over a wide range of speeds and
fuel ratios, while the fast acquisition system was used in characterizing that
performance at a given speed and fuel ratio.
The analogue signals corresponding to engine speed, torque, natural gas
and air flow rates and ignition advance taken at 27.5 kHz and averaged over
400 samples by DATAQ were scaled using instrument calibration factors and
the flows corrected for standard pressure and temperature given ambient con•
ditions. The ambient temperature used was an average of the manifold intake
temperature measured by thermocouple and the test cell room temperature.
The CRUNCH program was then used to evaluate the performance pa•
rameters. Brake Power, BMEP, Bsfc, relative air fuel ratio (RAFR) and brake
thermal efficiency were calculated:
Brake Power =Torque * Speed *2TT/1000
BMEP =Torque *4TT/(L/2)
Bsfc =mg * 3600/Brake Power
T]tn =Brake Power/rrifl * LHV
A brake power (BP) correction for standard conditions assuming 85 %
mechanical efficiency in accord with SAE J1349 requirements was then per•
formed, being defined by:
9 B BPcorrected = BP * (1.18 * (^ -)(^)° - 0-18)
-Tomb -^yo
The composition of the fuel, BC Natural Gas, and the calculation of the
Lower Heating Value (LHV) used in this analysis is given in Appendix C.
The Indicated Mean Effective Pressure (IMEP) was calculated from both
the individual pressure traces and the ensembled trace generated from the Chapter 4. Data Analysis
analogue processing. Statistical analysis was performed to calculate the mean,
standard deviation and the coefficient of variance (COV) of IMEP from the
ensemble average.
where VCBDC was the volume at BDC before compression and VCBDC was
the volume at BDC after expansion. The IMEP coefficient of variance was
determined by:
IMEP
This statistical analysis was also applied to the peak pressure data and the
angle of occurrence of the peak pressure. IMEP and its COV was selected as
the characterizing parameter for cyclic variation from those reviewed, due to
the direct relation between engine drivability and fluctuations in the output
torque ie. IMEP COV [19].
Peak pressure and position were also used for comparison with previous
squish jet work by Dymala-Dolesky [18] and with other researchers [1, 36,
45, 46]. The conditions at which the engine operates have been shown to
influence the governing characteristics. Statistical studies by Belmont et al
[54] using autocorrelation techniques over a wide range of conditions show,
that in the lean burn operating regime all of the primary cycle features are
strongly related.
The preferred sample size used in the fired analysis was 200, in line
with Amann's recommendations of a minimum of 100 and recent research
by Matekunas [45] 120, Nakamura [68] 200, and Lancaster [76] 100-300 and
within the capabilities of the acquisition and computation systems. Very large
sample sizes have generally been restricted to frequency information [76] or Chapter 4. Data Analysis
with on line point data acquisition, ie., peak pressure measurements by Pat•
terson [39] and Belmont et al [54] with 2048 cycles taken. Increased sampling
error was evident in the smaller data set of 44 cycles. However theses data
were useful in comparison with tests taken on the same pistons, using this
number. A cautionary note is supplied by Brown [71] "Consistency is no in•
dication of accuracy". Lancaster also maintained that the number of cycles
of data required was dependent on the use of the data and the stability of the
running condition.
4.3.4 Combustion Analysis
The final phase of analysis of the individual pressure traces was used to inves•
tigate the combustion process. The mass fraction burned was evaluated based
on a simple procedure developed by Rassweiler and Withrow [49] from the
correlation of cylinder pressures with flame front photographs. After extract•
ing the pressure rise due to piston motion from the pressure records Rass•
weiler and Withrow found that the fractional pressure rise due to combustion
was approximately equal to the fractional mass burned at the corresponding
instant in the combustion period.
Using this technique in the current work, the effect of piston motion and
combustion were evaluated as though taking place sequentially not simulta•
neously.
p(i) piston O APCOMB Chapter 4. Data Analysis
The pressure increase due to piston motion was calculated assuming poly•
tropic compression and expansion by:
which was extracted from the pressure record. The increase of pressure due
to combustion alone was then evaluated as occurring at a constant volume at
each stepped volume. The change from variable volume to constant volume
was made by scaling to a reference volume, ie. Volume at TDC. Rassweiler
and Withrow justified this from earlier work by G. Brown in 1925 showing
"Pressure increase produced in a const ant-volume bomb by liberating a given
amount of energy in a given mass of charge is inversely proportional to the
total volume of the bomb".
= v({ + *) + y(0 v * _L_
2 VTDC
APcomb = {AP(i + 1) - AP^) * Vref
The mass fraction burned was then calculated by dividing the summed
pressure increase due to combustion at each point by the total pressure in•
crease due to combustion:
Mass fraction burned(i) = ^ eomb(')
W Y,APcomb{total)
By definition, this simplified procedure scales the mass burn fraction be•
tween 0 and 1. Studies by Amann [19] show good agreement of this method
with more advanced computer techniques based on multiple flame zone heat
release models.
Start of combustion was taken at the crank angle of spark occurrence. End
of combustion was determined where the pressure increase due to combustion
was zero or negative. Chapter 4. Data Analysis 64
Several methods of polytropic exponent calculation were investigated:
• Experimental determination of 7C and 7e and an average used as per
Rassweiler and Withrow; 7 of 1.3 [57].
• Experimental determination of 7C prior to spark and 7e from end of
combustion with a linear relation used in the intermediate region, as per
Dymala-Dolesky [18].
• Arbitrary choice of an appropriate 7, as per Boisvert; 7C of 1.3 [74].
Rasswieler and Withrow found that changes between 1.25 and 1.35 had
little effect on their analysis, and noted that for the equations for constant vol•
ume bomb development to hold, the same polytropic exponent was required
in the burned and unburned region. In this study the polytropic coefficient
for compression was calculated from the average slope of the logarithmic
pressure-volume curve in the range ten degrees before spark. This value was
used throughout. A sensitivity analysis showed that individual values of -yc
calculated from each pressure trace were required in the massburn calculation
with an average value of 7C only suitable for the ensembled pressure calcula•
tion. Figure 4.2 shows a typical plot of the polytropic coefficient versus crank
angle.
Characterizing combustion parameters of the initial burn period and burn
duration were calculated using the ranges 0 to 1% and 0 to 5% and 1% to 90%
and 5% to 90% respectively. The two ranges were calculated to provide a
wider basis for comparison with others in the arbitrary selection of initial
and main burn period. Statistical analysis of the mean, standard deviation
and COV of these parameters was carried out similarly to the pressure data
calculations. Chapter 4. Data Analysis
The findings generated in the motored engine and fired engine tests of
the six chamber configurations are discussed in the next chapter. Tables 5.2
to 5.24 and Figures 5.1 to 5.100 present the test results and performance
measurements obtained ove a period of 20 months using two data acquisi•
tion systems. As noted in Chapter 1, the range of tests and measurement
parameters selected for this investigation was designed to allow the broadest
comparison with conclusions drawn from previous combustion and turbulence
studies. Chapter 5
Experimental Results and Discussion
5.1 Introduction
In this chapter the results obtained from the experiments conducted on the
Ricardo research engine are presented and discussed. In the first two sec• tions the results from the motored and fired tests are presented. The third section evaluates the uncertainties of the measuring techniques and the final section discusses the relation between the flow field measurements and the performance measurements.
The six combustion chamber geometries are compared over the range of speeds, load conditions and air fuel ratios given in the method section. Com• parisons are made between all six configurations and by separation into the two groups, 'bathtub' and 'disc' defined in section 3.2.3.
5.2 Motored Tests
5.2.1 Motored Pressure Results
The motored pressure data taken by the fast acquisition system were used primarily to determine the gas properties required in the hotwire analysis of the in-cylinder flow field. These data also provided a powerful check on the phasing of the analogue signal of interest, either pressure or hotwire, and the acquisition trigger signal.
66 Chapter 5. Experimental Results and Discussion
The comparison of the motored pressure traces for the different chambers
also gives an indication of the similarity of the clearance volume of the cham•
bers. In their heat release analysis of engine pressure data Gatowski et al
[50] confirmed that the cylinder pressure, in addition to combustion effects,
responds to changes in the volume of the combustion chamber, heat transfer
to the walls and mass leakage.
Volume and compression ratio
Figures 5.1 to 5.4 show the motored pressure trace for the various chambers
at four speeds, for the 360 degrees of crank angle from BDC of the intake
stroke to BDC of the expansion stroke. Zero crank angles is assigned to TDC
of the compression stroke. These figures show that the maximum developed
pressure, at each speed, was similar for the different chamber configurations,
with the exception of the disc piston. Volumetric checks of the combustion
chamber cavities and clearance heights indicated that the clearance volume
was within 5% for all the chambers except the disc. The compression ratio
for these configurations was approximately 9.0:1. In situ depth measurements
of the disc configuration revealed a compression ratio closer to 10.0:1. Aside
from any difference in the compression ratios, a difference in heat transfer
from the larger surface area of the bowl-in-piston and bathtub type chambers
may also produce a decrease in pressure. This is supported by the lower peak
pressure of the squish jet piston. This chamber had the largest surface area,
estimated to be 30% greater than the bowl-in-piston without channels [18].
Figure 5.5 shows the effect of engine speed on the motored pressure for
the single slot chamber. This figure shows an increase in peak pressure with
increased speed due to the shorter time for heat transfer to the cylinder walls. Chapter 5. Experimental Results and Discussion
On all the pressure plots a high frequency fluctuation in the pressure trace is
seen at approximately 130 degrees BTDC, increasing in extent at the higher
speeds. This is the point of inlet valve closing and the fluctuation is attributed
to vibration of the valve. While this fluctuation had no effect on the overall
pressure trace, it proved detrimental to the temperature calculations based on
the pressure data. Changing the temperature calculation after valve closure
from an assumption of adiabatic compression to the use of the ideal gas law,
meant that a reference temperature could be taken from a spike. This section
of the pressure trace was smoothed prior to the temperature calculation.
Temperature
Figure 5.6 shows the calculated temperature at 33.3 rps for all chambers. The
temperature profile for the disc still shows some fluctuation at the end of the
filtering region, 60 degrees BTDC. The filtering routine was restricted to the
most critical region to prevent the loss of information in the area of greater
interest,around TDC.
Table 5.1 lists the compression and expansion coefficients averaged over
speed and RAFR used in the temperature calculations. The choice of these
coefficients is discussed in the analysis section 4.2.2.
5.2.2 Flow Field Results
The flow field measurements, taken approximately 2mm below the spark plug
position, have been represented by window ensembled mean velocities and
intensities per crank angle for the compression and expansion strokes. The
differences in the velocity and intensity profiles calculated using a stationary,
cycle-by-cycle ensembled method, and those calculated by non-stationary, Chapter 5. Experimental Results and Discussion
window ensembled, techniques is shown in Figure 5.7. The window averaging
and interpolating techniques used in the analysis have the affect of reducing
the spikes in the mean velocity without changing the magnitude of the over•
all profile. The difference in the analytical methods is more pronounced in
the intensity plot. Without the removal of the cyclic fluctuation from the
instantaneous velocity, the turbulent intensity may be grossly overestimated.
The positions of the inlet valve closing, IVC, and the exhaust valve opening,
EVO, are also shown on these plots.
Figures 5.8 and 5.9 show the effect of the piston chamber geometry on
the velocity and turbulent intensity at 33.3 rps. These and all flow field
measurements were taken at WOT conditions. Although these figure show
very different traces around TDC from the various chamber configurations,
they also show the common high velocity and intensity generated through
shear past the intake valve, relaxing after valve closure and decaying to TDC.
A rapid increase in velocity and intensity is also evident after the exhaust
valve opens.
Combustion chamber geometries
A detailed description of the effect of the different combustion chamber ge•
ometries on the in-cylinder flow field follows. Figures 5.10 to 5.17 compare
the six configurations in two groups at 20.0 rps and 33.3 rps. Measurements
at a higher speed, either 50.0 rps of 66.7 rps obtained for four of the cham•
bers are given in Figure 5.18 to 5.23. These later figures show that the mean
velocity and intensity generally increase with engine speed. This tendency is
clearly evident during the intake stroke though to TDC. However after TDC,
this trend is less obvious, showing some crossing over of traces such as the Chapter 5. Experimental Results and Discussion
castellated piston at 20.0 rps.
Bathtub chambers
Figures 5.10 and 5.12 show the mean velocity profiles for the bathtub and
modified bathtub piston. These figures show a decrease in, or the lack of
formation of, the high mean velocity prior to TDC, with a sustained relatively
high mean after TDC, for the modified pistons. Figures 5.Hand 5.13 show a
similar tendency in the intensity profiles. A possible reason for this tendency
is the addition of a slotted ridge to the standard flat piston, used with the
bathtub cylinder head, promotes early break down of the large scale intake
generated turbulent motion. The ridge may prevent the formation of large
scale rolhng motion within the chamber such as 'barrel' motion, a source of
energy at a latter stage.
These figures also show the predominant squish motion of the bathtub,
evident by a 'hump' displaced after TDC. The essentially one sided squish
area of the bathtub configuration, and the location of the hotwire probe at
one side, has been suggested as a plausible cause for the displacement of this
hump [74].
Disc and bowl-in-piston chambers
The non squish disc configuration and the enhanced squish and squish jet
bowl-in-piston type chambers are described next. The mean velocity and tur•
bulent intensity profiles for these pistons are compared in Figures 5.14 and
5.17. All chambers showed a decrease in the mean velocity form IVC reach•
ing a minimum at TDC. The squish action generated by the bowl-in-piston Chapter 5. Experimental Results and Discussion
chamber configuration produced a visible local increase in the mean at TDC,
however this was substantially lower than the other chambers.
While the disc chamber phenomenon! may be attributed to the higher
compression ratio, the disparity with the squish jet mean velocity profile was
unexpected. The performance and combustion characteristics of this piston,
discussed in the following sections, do not support such low values. Exami•
nation of the probe position in the bowl-in- piston measurements, shown in
Figure 3.7, suggests that the wire location of 4.7 mm below the cylinder head
may be outside the region of the plane squish flow. The destruction or ab•
sence of a main squish motion at TDC, ie a 'hump', by the added channels in
the squish jet piston was consistent with previous evaluations of this design
[17, 18].
scaling with mean piston speed
Figures 5.26 to 5.37 give the mean velocity and turbulent intensity profiles
scaled with the mean piston speed. The range of the scaled values at differ•
ent crank angle positions, for example TDC compared to 30 degrees ATDC,
illustrates the limitation of using single point measurements to determine
characteristic parameters from the flow field. The profiles for the mean ve•
locity generally 6how the scaled values in the region of TDC to be between
0.2 and 0.5. These are comprable with the findings of Bopp, Vafidis and
Whitelaw [22]. However, the scaled turbulent intensity values in the region of
TDC, ranging from 0.05 to 0.2 are substantially lower. This reduction may be
partly due to the rms intensity evaluated using window averaging techniques
being lower than ensemble averaged values. Differences in the geometry of the
chambers used and in the position of flow measurement are also significant Chapter 5. Experimental Results and Discussion
factors in the range of values obtained.
5.3 Fired Tests
5.3.1 General Performance Parameters
Engine performance parameters, calculated in accord with SAEJ1349, are
given in Tables 5.2to5.9 for stoichiometric and lean operation.
Full load performance
Tables 5.2 to 5.4 compare the different piston geometries at the three speed,
20.0, 33.3 and 50.0 rps for WOT full load operation. These tables show
marked differences in the brake power and thermal efficiency values for the
various chambers.
At all conditions the bowl-in-piston chamber gave the best performance
and the squish jet or castellated pistons the poorest. These tables also show
an improvement in the relative performance of the single slot piston with
speed. The brake specific fuel consumption (Bsfc) displayed a similar trend,
being lowest for the bowl-in-piston configuration and generally highest or
least economical for the squish jet piston.
Part load performance
Table 5.5 details part load operation at 2.5 bar bmep at 33.3 rps. This
measurement also shows that the bowl-in-piston chamber has the highest
brake power and efficiency and the castellated piston the lowest. However
the squish jet piston performance is now closer to that of the bowl without
channels. Chapter 5. Experimental Results and Discussion 73
Table 5.6 comparing three piston geometries at the part load condition
of 3.5 bmep at 50.0 rps again shows the bowl-in-piston produced the best
performance, the single slot piston comparable, and the castellated piston
gave the poorest. These part load cases also show that there is some variance
in maintaining the part load condition at a pre set BMEP.
For stoichiometric operation, Table 5.5 shows a 12% range. This however,
does not appear to be the cause of the difference in performance between the
chambers. The castellated piston exhibited relatively poor performance when
operating at both higher and lower values of BMEP than the other pistons.
In all cases the single slot piston performed as well as or better than the
castellated piston. For the lean cases the single slot chamber approached or
exceeded the standard bathtub chamber performance.
Brake thermal efficiency
A closer examination of the effect of piston chamber geometry on the brake
thermal efficiency at different air fuel ratios for full load operation is made
in Tables 5.7 to 5.9, and illustrated in Figures 5.38 to 5.43. A discussion of
these results follows.
Figure 5.39 shows a continuous rise in efficiency with leaner mixtures for
the single slot chamber, while the bathtub chamber displays constant or de•
creasing efficiency after a relative air fuel ratio of 1.2. This same trend is
shown at 20.0 rps, Figure 5.38 and 50.0 rps, Figure 5.39 with a pronounced
improvement of the single slot chamber exhibited at the higher speed com•
pared with the performance of the bathtub chamber. These figures also show
an improvement in efficiency with the castellated piston at the higher speeds,
though it was still the lowest of the six chamber designs tested. The data Chapter 5. Experimental Results and Discussion
for the bathtub chamber at 50.0 rps were taken from multispeed tests rather
than from a single speed versus relative air fuel ratio set. These data however
were consistent over the range examined.
The next group of Figures 5.41 to 5.43 illustrate the results for the disc
and bowl-in-piston type chambers. These figures show the same trend in
brake thermal efficiency, at increasingly lean operation as discussed for the
performance tables. At all speeds the bowl-in-piston chamber had the highest
efficiency, increasing with leaner mixtures, the disc chamber had a constant,
lower efficiency and the squish jet piston the lowest.
Finally comparing the bathtub and disc group of chambers at 33.3 rps,
shows the bathtub chambers to be similar to the disc chamber.
Ignition advance
The effect of chamber geometry on the amount of ignition advance required for
MBT operation is also shown in Tables 5.7 to 5.9. This is a useful comparison
because reduced ignition advance is an indication of faster burn rates. In
general ignition advance shows less pronounced trends between the bathtub
type chambers than the disc bowl-in-piston group. Figures 5.44 to 5.46 show
a similar advance required for the single slot and the castellated pistons,
increasing slightly with leaner mixtures, with the bathtub chamber requiring
more advance at the higher speeds. The reduced advance required by the
bowl-in-piston type chambers suggest a possible benefit of these types of
chambers in the extension of the knock limit. Where slower burn rates require
very early advance, the chance of preheating of the end gases leading to knock
is more prevalent. Chapter 5. Experimental Results and Discussion 75
5.3.2 Fired Pressure Results
Fired pressure traces
Figure 5.51 to 5.68 show the ensembled fired pressure trace versus the crank
angle position for the 360 degrees from BDC before the compression stroke to
BDC after the expansion stroke. These plots are divided into the two piston
groupings, 'bathtub type' and 'disc bowl-in-piston types', for three speeds
and two air fuel ratios at full load. The part load cases for 2.5 BMEP at 33.3
rps are also shown in two groupings for clarity. The smaller data set for 3.5
BMEP at 50.0 rps for three pistons is shown in Figure 5.67 and 5.68.
Figure 5.50 shows typical fired and motored pressure traces for the single
slot piston at 33.3 rps and WOT stoichiometric operation. The pressure trace
is given over two complete revolutions from BDC of the expansion stroke and
through the exhaust, intake, compression and expansion strokes. Compres•
sion TDC is taken as the reference of zero degrees. The motored pressure
shows the peak within 2 degrees of TDC, confirming proper phasing ot the
signal. The fired peak pressure is approximately 17 crank angle degrees after
TDC for this condition. •
A more quantitative comparison of the effect of combustion chamber con•
figuration on the pressure developed is shown by the characteristic parameters
of IMEP, peak pressure and angle of occurrence of peak pressure. The full load
cases are given in Tables 5.10 to 5.12 and the part load cases in Tables 5.13
and5.12 Chapter 5. Experimental Results and Discussion
Full load operation
In this section the trends are first highlighted in a comparison of the pressure
traces and then discussed in relation to the cycle statistics.
The two sets of Figures 5.51 to 5.56 and 5.57 to 5.62 for full load operation
exhibit the same trends in peak pressure for the bathtub and disc groups.
The chambers with multijets, ie., squish jet and castellated pistons, show a
lower peak pressure than the equivalent non-jet pistons, at all conditions.
Figures 5.53 and 5.54 show that at the mid speed the single slot developed a
comparable peak pressure to the standard bathtub chamber. However at the
higher speed, Figures 5.55 and 5.56 show the bathtub developed the highest
peak pressure.
Figures 5.59 and 5.60 show that the standard disc and bowl-in-piston
pistons were similar at the mid speed. This was contrary to expectations for
this non squish case but may be explained by the higher compression ratio.
No measurements were taken at 50.0 rps, however a set of data taken at
66.7 rps showed very high pressures and intermittent knock at stoichiometric
operation.
Part load operation
The fired pressure traces for the two modified bathtub pistons are compared
at fixed part load operating conditions 2.5 BMEP at 33.3 rps in Figures 5.63
and 5.64. The standard bathtub chamber proved unstable for 'fast' data
acquisition at this condition. However general performance parameters were
taken by the 'slow' acquisition technique as shown in Table 5.3. Again the
single slot piston gave a higher peak pressure than the multijet, castellated Chapter 5. Experimental Results and Discussion
piston at the both stoichiometric and lean operation. This trend is also
displayed in the disc group, shown in Figure 5.65 and 5.66 where the plain
squish, bowl-in-piston chamber exhibits the highest peak pressure. The non
squish disc piston had the lowest peak pressure and broadest trace at lean
operation.
Comparing the chambers at stoichiometric operation, the standard disc,
single slot and squish jet chambers produced similar traces. A marked dif•
ference however is seen in the bowl-in-piston chamber with the highest peak
pressure and the castellated piston with the lowest. The bowl-in-piston and
squish jet chambers also gave a more peaked pressure trace suggesting faster
burn rates.
Part load operation at 3.5 BMEP at 50.0 rps shows the same ordering for
stoichiometric and lean operation. At both conditions the bowl-in-piston
shows the highest peak pressure with the castellated piston exhibiting a
smaller difference compared to the single slot piston at lean running.
Full load IMEP
The full load IMEP statistics for the various chambers given in Tables 5.10
to 5.12 exhibit pronounced trends in the mean IMEP and lesser trends in the
coefficient of variance data. Comparing the bathtub chambers only, the single
slot chamber consistently gave the highest IMEP and the castellated piston
the lowest. The differences in IMEP ranged from 2-5% at stoichiometric
operation to 6-8% at lean operation over the speed range examined.
These tables show a less obvious trend in the coefficient of variance (COV)
of IMEP. At lean operation the single slot piston displayed the least variability
with the castellated piston having a lower COV than the bathtub at the Chapter 5. Experimental Results and Discussion
lowest speed, 20.0 rps, only. At stoichiometric operation the single slot piston
displayed the least variation at the higher engine speed of 50.0 rps only. Over
the speed range tested the differences in COV between the chambers ranged
from 7-30% at stoichiometric and 6-50% at lean operation.
Similarly comparing the IMEP for the disc group, the disc piston had the
highest mean IMEP and the squish jet the lowest. The COV information
produced a mixed result. The disc chamber gave the lowest COV for stoi•
chiometric operation and the highest for lean operation at 20.0 rps. At 33.3
rps the disc piston again displayed the least variation at stoichiometric wtih
comparable values at lean operation.
The tables generally indicate the highest COV for the squish jet piston at
lean operation and high speed with a 50% difference between the six pistons.
Part load IMEP
Table 5.11 compares the pistons at the slower part load condition and shows
an increased difference in the COV for the castellated and disc pistons from
the single slot and bowl-in-piston types. The modified bathtub chambers
show increased COV's at part load with the differences between the pistons
being less marked at lean operation. IMEP was again highest for the single
slot chamber. The largest difference is shown in the disc chamber COV at
stoichiometric operation where it is an order of magnitude greater than the
bowl-in-piston types.
Comparing the IMEP values for all the chambers at all conditions shows
that the bathtub type chambers generally developed higher IMEPs than the
disc or bowl-in-piston type chambers. Chapter 5. Experimental Results and Discussion
Angle of occurrence of peak pressure
The angle of occurrence of peak pressure is also given in Tables 5.10 to 5.14.
For all conditions of speed, load and relative air fuel ratio, the maximum
spread of peak position for all chambers was eight crank angle degrees. The
latest peaks occurred approximately 20 degrees ATDC.
Barton et al [36] used the angle of occurrence as an indication of flame
speed. His studies however, were based on the same ignition advance for all
tests. This does not hold for the different values of ignition advance required
for MBT timing, and hence the position of the peak occurrence based on
TDC is an indirect indication of burn rate only.
The coefficient of variance of peak position provides a measure of the
phasing variation. This variance is dependent on the reference position, eg.,
measuring the peak pressure angle of occurrence from TDC the COV ranges
from 15% to 70%. Hence the significance of the COV is debatable. The stan•
dard deviation of peak position however, does give an indication of phasing
variation. These tables show that trends in the peak position are not consis•
tent with IMEP or performance trends. This last result confirms Matekunas'
[45] view that the position of peak pressure occurrence is less useful than
IMEP as a measure of engine drivability.
The remainder of this section discusses the general trends found in the
pistons with and without jet action in relation to the findings from previous
studies. Chapter 5. Experimental Results and Discussion
Previous studies
The full and part load results for the peak pressure in this study indicate
higher values for the plain squish or single slot piston chambers. These results
are in line with comparative squish and squish jet results of Dymala-Dolesky
[18]. At all operating speeds and conditions Dymala-Dolesky found that
the chambers without slots had the highest peak pressures. However, the
same consistency was not evident for the IMEP values. Dymala-Dolesky
concluded that the best IMEP was obtained for the piston with eight holes
directed upwards, (the piston termed 'squish jet' in this study), and that the
bowl-in-piston had the lowest IMEP. The current results show the bowl-in-
piston to have a similar or higher IMEP than the squish jet.
In examining the effect of compression ratio, Dymala-Dolesky tested an
alternative bowl-in-piston chamber with a marginally higher compression ra•
tio. He obtained IMEP values with this chamber similar to the squish jet
chamber with eight channels. However, the effect of minor changes in the
chamber volume noted by Dymala-Dolesky, has not been confirmed in the
combustion geometry or compression checks performed in this work. Other
factors such as excessive leakage past the piston rings or wearing in of the pis•
ton may account for the original low values of IMEP measured in the previous
study.
5.3.3 Mass Fraction Burned Results
The results of characterizing the combustion process through a mass fraction
burn analysis are discussed in this section. The mass fraction burned data
calculated on a crank angle basis are displayed in three ways. Figures 5.69 Chapter 5. Experimental Results and Discussion
to 5.84 show the cumulative mass fraction burned against crank angle posi•
tion. Tables 5.15 to 5.23 list the durations of the initial and main combustion
periods determined from arbitrary percentage limits, and measured in terms
of crank angle degrees. Lastly Figures 5.89 to5.98 present bar graph pre•
sentations of mass fraction burned ratios showing the initial and main burn
durations for each chamber configuration relative to the slowest or standard
bathtub chamber.
The general observed trends are discussed in terms of the mass fraction
burned plots while quantitative comparisons are made from the statistical
results of the duration calculations.
Mass fraction burned traces
Figure 5.69 shows the cumulative mass fraction burned per crank angle for 60
degrees BTDC to 60 degrees ATDC. The simple mass burn analysis, described
in the previous chapter, assumes the start of combustion at spark and the end
of combustion where the pressure rise due to combustion is zero. The mass
fraction burned is therefore scaled between 0 and 1 from start to end of
combustion. For the bathtub slot chamber at full load 33 rps and RAFR of
1.27, the figure shows the start of combustion at spark 38 degrees BTDC,
with a slow initial burn period followed by a rapid increase in mass fraction
burned and a final leveling off at approximately 45 degrees ATDC. The 1%,
5% and 90% mass fraction burned percentages are also shown. The initial
burn period is taken from spark to 1% or 5% and the corresponding main
burn period from l%-90% or 5%-90%. Chapter 5. Experimental Results and Discussion
Full load operation
Figures 5.70 and 5.71 compare all six piston chambers for stoichiometric and
lean conditions at WOT 33.3 rps. While gross differences in the slope are
visible, eg. the single slot piston compared to the bowl-in-piston chamber,
the different values of ignition advance make direct comparisons unclear. The
distinct difference between the bathtub type pistons and the bowl-in-piston
type pistons supports the division of the results into these two groups. These
figures also indicate a greater variance between the combustion chambers at
lean operation.
The effect of engine speed on the burn rate is shown in Figure 5.72 for
the bathtub chamber at WOT and lean operation. This figure shows the
increase in ignition advance required as the speed was increased. The figure
also illustrates that the mass fraction burned curve for a faster burn at higher
speed does not necessarily result in a steeper slope when plotted against crank
angle (ca). For example the main combustion duration at 50.0 rps for this
chamber was 40.3 ca or 2.2 ms, while the similar crank angle period 41.8 ca
at 20.0 rps is 5.8 ms.
The raw pressure data for the 40.0 and 50.0 rps standard bathtub cases
were provided by A.Kapil [77] of the AFL Group from measurements taken
over the same period as this study. These data were collected for 100 cycles.
Figures 5.73 to 5.78 for the bathtub type chambers show a slight increase
in the rate of burn for the single and castellated pistons over the bathtub
chamber. This improvement was most pronounced for lean operation at 20.0
and 33.3 rps. Figure 5.77 for stoichiometric operation at 50.0 rps shows
similar slopes for all three chambers, the spacing of the plots resulting from Chapter 5. Experimental Results and Discussion 83
the bathtub chamber being advanced a further six degrees. In all cases the
single slot chamber shows a marginal improvement over the castellated piston.
The corresponding Figures 5.83 to 5.84 for the disc group show the pro•
nounced effect of the plain squish motion. In all cases the bowl-in-piston and
squish jet chambers produced an increase in the burning rate shown by the
reduced ignition advance and steeper slopes. Similar to the bathtub types
this affect was emphasised at lean operation. These figures also indicate that
the bowl-in-piston without the squish jets produces a faster main combustion.
Part load operation
Five of the chambers are compared at the fixed 2.5 BMEP 33.3 rps part load
condition in Figures 5.85 and 5.86. The bathtub chamber as mentioned was
unstable at this condition. Again these figures show that the bowl-in-piston
type chambers have a shorter burn duration than the bathtub chambers with
the bowl-in-piston having the shortest. This comparison also shows that the
disc chamber and the single slot chamber were similar, with the castellated
piston producing the slowest burn. Similar trends are seen in Figures 5.87
and 5.88 for the 50.0 rps part load case.
The remainder of this section covers the selection of the initial burn period
limits and a discussion of the duration results.
Combustion duration
The initial and main burn durations were determined for two ranges, using
1% and 5% mass fraction burned as the cut off for the initial burn period. A
comparison of the two arbitrary limits show the same general trends in the ef•
fect of the combustion chamber configuration on the combustion. Differences Chapter 5. Experimental Results and Discussion
in the initial burn period for the different piston shapes are more pronounced
when 1% is used, particularly for the single slot and castellated pistons.
For example in Tables 5.17 and 5.18 comparing the initial burn periods
for the bathtub chambers, at 33.3 rps full load, shows a 32% range between
chambers when 1% initial burn was used and a 22% range when 5% was
used. In separating the initial flame initiation and kernel development from
the main burn period, the attempt is made to find a link between the type
and timing of the turbulent motion generated, and the combustion process.
This is discussed further in the evaluation of the measuring technique and
uncertainties.
Mass fraction burned ratio
Calculation of a representative mass fraction burned ratio has been performed
on the 5% initial burn and subsequent main burn period 5%-90% only. Where
possible the different configurations are referenced to the standard bathtub
chamber. Where this chamber was unavailable the 'slowest' chamber is used.
The similarity of the standard bathtub and disc chamber durations allows a
comparison on a two group basis as before.
Figures 5.89 and 5.90 for full load operation at 20.0 rps show that the
modified bathtub pistons marginally decreased the initial and main combus•
tion durations at stoichiometric operation with a more pronounced effect on
the main burn period at lean operation. These figures also show that while
the bowl-in-piston and squish jet chambers have a dramatic effect on both
the initial and main combustion periods, the effect on the initial period de•
creased with lean operation. At lean operation the disc chamber also showed Chapter 5. Experimental Results and Discussion
a decreased main burn period. At the higher speed of 33.3 rps shown in Fig•
ure 5.91 and 5.92 the same trend in reduction of the main burn period by
squish action is displayed. The reduced initial burn period observed with the
castellated piston was not repeated at lean operation.
The 50.0 rps full load case given in Figures 5.93 and 5.94 show some
reversal of the effect of the castellated piston on the main burn, that is,
a negligible effect at both stoichiometric and lean operation. In all cases
the bowl-in-piston produced the shortest initial and main burn periods, and
the single slot chamber generally gave shorter durations than the castellated
piston.
Finally, the part load results presented in Figures 5.95 to 5.97 show the
greater effect of plain squish or a single jet over the multi squish jet action in
reducing the main combustion period. While all the squish pistons reduced
the main burn period compared to the non-squish disc chamber, the single slot
piston also produced a decrease in the initial burn period at lean operation.
In comparing the different combustion chambers tested in this study with
each other and with other designs, the chamber geometry is only one of the
variables. The location of the 6park plug in reducing the maximum distance
traveled by the flame is also an important variable. The combustion con•
figurations used in this study may be roughly equated to those used in a
comparative study by Heywood [61], for example, comparing the bathtub
chamber to the 'hemi' chamber with side spark plug location, the single slot
and castellated pistons to the 'open' chamber with side spark and the disc
chamber and bowl-in-piston chamber as given. Within these limits these re•
sults are consistent with Heywood's study, which showed that the geometry
had the greatest impact on the main burn period. The central location of Chapter 5. Experimental Results and Discussion
the spark was also shown to be a major factor in reducing the combustion
duration supported by the more centrally located spark in the bowl-in-piston
types showing an improvement of over 50% compared to the single slot side
spark maximum improvement of 15%.
The following section provides a discussion of the uncertainties contained
in the above results and an evaluation of the experimental technique.
5.4 Experimental Uncertainties and Technique
The uncertainty associated with the results of this investigation may be di•
vided into two areas. Errors involved in the equipment sensitivity, accuracy
and repeatability are first discussed. Uncertainties inherent in the experi•
mental analysis techniques are then discussed under the same headings as the
experimental work.
5.4.1 Flow Measurement
Pressure measurement
Both the pressure transducer and the amplifier were repeatedly calibrated
and had a sensitivity error less than 1%. Digitalization error of the pressure
signal is also less than 1%. The largest source of error lies in the choice of
a reference pressure, ie, the inlet manifold pressure assigned at BDC. This
however is a scaling error and of little concern in a comparative analysis. The
pressure measurement is assigned an error of 2%. Chapter 5. Experimental Results and Discussion
Hot wire measurement
The estimate of the error associated with the flow field measurements is 30%.
This is a combination of the uncertainty in the hotwire measurements and in
the assumptions of isotropic and homogeneous turbulence used in the separa•
tion of a fluctuating velocity term from a mean in the absence of a definitive
mean flow. The complex flow environment of the engine cylinder, the lack of
directional sensitivity of the probe, the high temperature gradients between
the fluid and the cylinder walls and the scaling of the calibration constants
using an analytical model, all contribute to the uncertainty of the flow mea•
surements. Sensitivity analysis on the effect of temperature on the hot wire
calibration by Dohring [75] estimated the error associated with this analyti•
cal model at 23%. The attempt was made to minimize the thermal gradients
present, by maintaining the cylinder coolant at a maximum of 80°C and plac•
ing the probe at the furthest distance from the surface while maintaining near
spark position. The probe was approximately 5 mm from the cylinder head.
Recent studies by Heywood [14] show the thickness of the thermal boundary
. layer at TDC of the order of 2mm.
The effects of the high error in the flow measurements are somewhat re•
duced by their systematic nature in this comparative study.
A final comment on the hotwire determination of the flow field deals with
the restriction of a single point measurement in representing the cylinder flow.
For example, the placement of the probe at 5 mm below the cylinder head
suggests that the low measurements for the bowl-in-piston chamber are due to
the probe being below the main squish action, not a reduction of turbulence
with this chamber. This explanation is supported by the fired performance Chapter 5. Experimental Results and Discussion
results of the bowl-in-piston chamber and previous studies.
5.4.2 Performance Measurements
Engine parameters
Performance measurements obtained by the slow data acquisition system were
dependent on analogue signals from a number of different instruments. As
described in the Chapter 3, extensive checks were carried out after each engine
rebuild prior to operation of the engine timing, throttle and torque positions.
The conversion of the analogue signals by the data translation system was also
periodically checked and found to be within 3% error. This was consistent
with similar measurements by Dymala-Dolesky three years earlier.
The largest error in the performance parameters came from the dynamic
instability of the torque system, resulting in a 5% error. The collected samples
of 400 data were taken after stable conditions had been maintained for a
minimum of ten minutes. On misfire or knock occurrence the data collection
was aborted.
Variations in the set operating conditions for the different combustion
configurations over the extended two year period of this investigation were
less than 1% for speed, 3% in air fuel ratio at lean operation, and for part
load operation 12% in BMEP.
Fired pressure analysis
Similar to the motored pressure the error in the pressure data is set to 2%.
Volume assignment for IMEP was assessed using the same clearance volume
for all chambers except the disc chamber. The error in the clearance volume Chapter 5. Experimental Results and Discussion 89
and subsequent compression ratio is 3%-5%.
The selection of the sample size has been discussed in the method section.
Statistical sampling errors were generally lower for the large sample sizes. No
strong memory element was observed in the IMEP, in line with Belmont et
al[54] who showed that there was little memory effect in the presence of high
cyclic variability and lean operation.
Figure 5.99 shows the IMEP variation over 200 consecutive cycles for the
single slot chamber at WOT, MBT and lean operation. In this case, and in all
fired pressure measurements the cyclic variability discussed is for stable oper•
ation, that is without misfire cycles. Similarly Figure 5.100 shows the IMEP
variation over 44 non consecutive cycles for the bathtub chamber at the same
operating conditions. Referring to the values in Table 5.11, the mean and
standard deviation (std dev) are also displayed on the plot. In this case there
is a slight increase in the standard deviation for the smaller sample. The
error associated with the sample size for the single slot is: mean ± 0.9, stan•
dard deviation ± 0.64 and coefficient of variance ±0.1, and for the bathtub:
mean ± 2.5, standard deviation ± 1.8 and coefficient of variance ± .14.
Mass fraction burned
The main approximations made in the mass fraction burned calculation were
the calculation of the polytropic oefficient and determination of the end of
combustion, Rassweiler and Withrow [49] found that a variation in the poly•
tropic coefficient between 1.25 and 1.35 had little effect. A sensitivity study on
the coefficient of compression (7C), comparing an ensembled averaged value
and individual cycle values, produced a smooth representative trace from Chapter 5. Experimental Results and Discussion
individually calculated coefficients only. The compression coefficient was cal•
culated from an average of values before spark. The use of Rassweiler and
Withrow's simple method was supported by Amann's [19] comparison with
more complex computer models.
In selecting 1% and 5% mass fraction burned for this study, the initial
burn period was also calculated for 2%, 3% and 10%. A recent sensitivity
study by Kapil [77] showed that below 2% mass burn there is a sharp increase
in standard deviation of the initial burn period associated with experimental
error and measuring techniques. He subsequently used extrapolated values
from 5% to get a theoretical standard deviation at zero burn time. However
the upturn in the cyclic variability need not necessarily be assigned to ex•
perimental error as it appears repeatable and may be a phenomena of cyclic
variation in mass fraction burned. Young [4] and Yatsamoto [46] both used
1% as the initial burn period to separate the 'ignition delay' (a somewhat
misleading term), from the main burn period. Other researchers have used
5% and 10% [17, 61].
Statistical sample size error
Finally in comparing the data taken over 44, 100 and 200 cycles a statistical
check was used to acertain whether there was a significant difference between
values.
Testing the 'no change'or null hypothesis (Ho) at a 5% level of significance
showed that a difference in the means of IMEP between two samples less than
5-10 bar caused 'no rejection' of the null hypothesis. Similarly over all sample
sizes and conditions a difference less than 50-100 bar in the peak pressure
caused 'no rejection'. A test at the 5% level of significance for a 'no change' Chapter 5. Experimental Results and Discussion 91
hypothesis in the standard deviation showed that a difference less than 30%
between two 44 cycle samples, less than 18% for one 44 and one 200 cycle
sample, or less than 5% for two 200 cycle samples, caused 'no rejection' of
the null hypothesis.
The effect of the number of cycles on the sample error of all the statistical
calculations, ie., IMEP, peak pressure and burn durations is indicated in the
number of significant figures given in the appropriate tables. The sample error
was based on the 'large sample' approximation for maximum likelyhood esti•
mates of mean, standard deviation and coefficient of variance. The equations
for these calculations were taken from Bury's statistical text [83].
The statistical analysis was used as a general check on the data obtained
and does not constitute a detailed analytical investigation.
5.5 Turbulence, Combustion and Performance
The remainder of this section examines the relation between the flow field
measurements and the performance measurements obtained in this study.
The general findings for the bathtub group of pistons illustrate that the
addition of squish jet action to the standard bathtub chamber enhances the
mean velocity and turbulent intensity of the in-cylinder flow. The subsequent
improvement in performance was confirmed by increased brake thermal effi•
ciency and indicated mean effective pressure (IMEP). The results show that
the effect of the directed squish motion was stronger when concentrated in
one jet,that is, in the single slot piston.
Examination of the MBT ignition advance for the bathtub group confirms
that the squish effect is present in the main burn period. Squish action Chapter 5. Experimental Results and Discussion
occurred after 20 degrees BTDC, with the ignition advance for the bathtub
pistons generally in the region 30 degrees BTDC.
The effect of the large squish area of the bowl-in-piston type chambers
in reducing in both the initial and main burn periods was illustrated in the
ignition timing of 15-20 degrees BTDC for these chambers.
Two significant differences in the turbulence and performance results for
the bowl-in-piston chambers are apparent. First, the low mean velocity and
turbulent intensity measurements for the plain bowl-in-piston were not re•
flected in low performance data. Second, the high flow field measurements
for the squish jet piston did not result in high performance. However, a
comparison with the results of previous studies and an examination of the
velocity measurement location suggests the first anomaly is a deficit of the
measurement recorded not of the physical phenomenon.
The turbulence and performance measurements for the squish jet piston,
however, were consistent with previous studies. Possible reasons for the ob•
served reduction in performance may be related to the effect of the high
level of turbulence and increased surface area. As noted previously, excess
turbulence can cause the disruption of the early flame kernel by convection
away from the spark point towards the chamber walls causing quenching.
The increased surface area of the chamber due to the channel passages also
contributes to heat loss and quenching.
The potential for improved part load performance through strong squish
action was clearly shown for the single slot and bowl-in-piston pistons. These
pistons gave a 2%-3% increase in thermal efficiency over the standard bathtub
and squish chambers. The coefficient of variance of IMEP for these concen•
trated squish action pistons was a third to a half lower than the multijet Chapter 5. Experimental Results and Discussion
designs, which is an indication of improved drivability.
The specific focus of the experimental work in this project has been to ex•
pand the information available on squish and squish enhanced turbulence in
affecting and promoting fast-lean operation using alternative fuels. Through•
out the study, conclusions drawn from this investigation have been reviewed
under a range of engine configurations and operating conditions. The conclu•
sions drawn from this investigation and the directions identified for further
research , are presented in the final chapter. Chapter 6
Conclusions and Recommendations
6.1 Introduction
The objective of this research was to investigate the influence of combustion chamber design on turbulence enhancement in the achievement of fast-lean operation of a spark ignition engine. The specific focus of the experimental work was to examine the effect of squish induced jet action.
A modified flat piston was designed for use with a bathtub cylinder head where the jet action was formed by a single slot or slots in a raised ridge on the piston surface. In the further investigation of bowl-in-piston chambers these jets were produced by adding channels from the top surface to the side of the bowl inclined towards the spark position.
The evaluation of six combustion chamber configurations was conducted in four stages:
1. Flow field measurements by hotwire at the spark location were taken
and subsequently analysed using a window average technique. Mean
velocities and turbulent intensities were obtained by this method.
2. Performance measurements for the engine fueled by natural gas over the
speed range 20-50 rps for stoichiometric and lean operation at full load
and part load were made.
94 Chapter 6. Conclusions and Recommendations
3. Combustion pressure histories were measured over the same range. Char•
acterizing parameters of Imep, peak pressure and angle of occurrence of
peak pressure were also used to evaluate the engine performance and as
a measure of the cyclic variability with each chamber.
4. Mass fraction burned curves were calculated using a simple rate of pres•
sure rise model to evaluate the effect of chamber design on the initial
and main burn durations. These durations for two phases were defined
as 0-1% and 1-90% and 0-5% and 5-90% mass fraction burned.
6.2 Conclusions
The conclusions drawn from this investigation may be summarized under the
principal areas of study:
6.2.1 Turbulence Studies
• The mean velocity and turbulence intensity at TDC increases with in•
creased engine speed.
• The addition of a ridge with one or more slots to the standard bathtub
piston reduces the mean velocity and turbulent intensity before TDC.
• The single slot or castellated wall produces or maintains a higher level
of turbulence after top dead centre, than the flat piston.
• The increased squish area of the bowl-in-piston and squish jet pistons
produces a higher mean velocity and turbulent intensity than the smaller
squish area of the bathtub type chambers. The bowl-in-piston measure•
ments were confirmed by previous studies [17, 18]. Chapter 6. Conclusions and Recommendations 96
• The addition of slots to the single slot ridge reduces the effect of the main
squish and jet action. Similar addition of channels to the bowl-in-piston
chamber reduces the squish effect.
• The squish jet action of forcing the flow into the lower portion of the
bowl as observed by Dymala-Dolesky [18] was confirmed by these mea•
surements.
6.2.2 Performance and Combustion Studies
• Enhanced turbulence by combustion chamber geometry is most effective
in improving performance at lean operation.
• The larger squish area and more centrally located spark configurations
produce the greatest reduction in the initial and main combustion du•
ration.
• Squish is most effective in the main burn period, occurring approxi•
mately 20 degrees before TDC, with ignition timing for the chambers
tested generally around 30 degrees before TDC.
• The single jet action of the single slot piston, directed towards the spark
is most effective in improving the efficiency at high speed and lean mix•
tures.
• High turbulence intensity is not necessarily beneficial to engine perfor•
mance when accompanied by increased heat loss and cyclic variability.
• The reduced ignition advance requirements for the modified bathtub
pistons indicate their design potential to extend the knock limit.
• The addition of the squish jet action has the greatest potential for im•
proving engine drivability at part load operation by reduction of the Chapter 6. Conclusions and Recommendations
cyclic variability.
In stating these conclusions, it is important to note that few trends were
consistent over all speed and load conditions for the range of specific engine
operating conditions tested in this investigation. Conclusions drawn from
some previous studies have been limited to specific operating conditions. The
following section contains recommendations on expanding the basis of com•
parison and test conditions following the range of engine and operating vari•
ables included in this investigation.
6.3 Recommendations
The trends identified from this experimental investigation provide direction
for further work on the analysis of the flow field developed and the interaction
of the jet motion in combustion chamber design. The limits of single point
HWA measurements in an engine have been well documented and confirmed
in this study. It is suggested that further studies of the 'jet effect' should be
conducted in a combustion bomb or rapid compression machine, ideally with
laser Doppler anemometry at multi positions.
The promising improvements of the single slot piston over the standard
bathtub configuration at high speed and lean operation suggest further that
studies should examine leaner mixtures and higher speeds within the limits of
the available apparatus. The bathtub piston modification was selected as the
most practical modification for this investigation, because it required changes
to the piston only. This however restricted the squish area available. Further
experiments using a modified bowl-in-piston chamber and the required step
in the cylinder head should be carried out to achieve the advantages of both Chapter 6. Conclusions and Recommendations
the large squish area action and jet slot action.
Finally it is recommended that exhaust emission studies be carried out
to ascertain whether the advantages of these fast burn squish jet pistons are
mitigated by the increased surface areas, and possible quenching effects of the
channels leading to unburned deposits. Bibliography
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[18] Dymala-Dolesky, R., "The Effects of Turbulence Enhancement on the Performance of a Spark- Ignition Engine", M.A.Sc. Thesis UBC. 1986. [19] Amann, C.A., "Cylinder Pressure Measurement and its use in Engine research", SAE Paper 852067, 1985. [20] Winsor, R.E., and Patterson, D.J., "Mixture Turbulence- A Key to Cyclic Combustion Variation", SAE Paper 730086, 1973. [21] Witze, P.O., "Measurements of Spatial Distribution and Engine Speed Dependence of Turbulent Air Motion in an IC Engine", SAE Paper 770220,1977. [22] Bopp, S., Vafidis, O, and Whitelaw, J.H., "The Effect of Engine Speed On The TDC Flow field In A Motored Reciprocating Engine", SAE Paper 860023, 1986. [23] Daneshyar, H., and Fuller, D.E., "Definition and Measurement of Tur• bulence Parameters in Reciprocating IC Engines", SAE Paper 861529, 1986. [24] Hall, M.J., and Bracco, F.V., "A Study of Velocities and Turbulence In• tensities Measured in Firing and Motored Engines", SAE Paper 870453, 1987. [25] Ikegami, M., Siaji M.,, and Nishimoto, K., "Turbulence Intensity and Spatial Integral Scale During Compression and Expansion Strokes in a Four-Cycle Reciprocating Engine", SAE Paper 870372, 1987. [26] Dent, J.C., and Salama, N.S., "The Measurement Of the Turbulence Characteristics in an Internal Combustion Engine Cylinder", SAE Paper 750886, 1975. [27] Haghgooie, M., Kent, J.C., and Tabaczynski, R.J., "Turbulent Time- Scale Measurements in a Spark Ignition Engine Using Hot wire Anemom- etry and fast Response Ion Probes", Symposium on Flows in IC Engines ASME WAM 1982. [28] Rask, R.B., "Laser Doppler Anemometry Measurement in an Internal Combustion Engine," SAE Paper 790094, 1979. [29] Fraser, R.A., Felton, P.G., Bracco, F.V., and Santavicca, D.A, "Prelim• inary Turbulence Length Scale Measurements in a Motored IC Engine", SAE Paper 860021, 1986. [30] Witze, P.O., "A Critical Comparision of Hot- Wire Anemometry and Laser Doppler Velocimetry for IC Engine Applications", SAE Paper 800132,1980. [31] Matsuoka, S., Yamaguchi, T., and Umemura, V., "Factors Influencing the Cyclic Variations of Combustion of Spark-Ignition Engine", SAE Paper 710586, 1971. [32] Gatowski, G.A., Heywood, J.B., and Deleplace, O, "Flame Photographs in a Spark-Ignition Engine", Combustion and Flame Vol 56 pp 71-81, 1984. [33] Keck, J.C., Heywood, J.B., and Noske, G., "Early Flame Development and Burning Rates in Spark Ignition Engines and Their Cyclic Variabil• ity", SAE Paper 870164, 1987. Bibliography
Witze, P.O., et al. "Measurements and Predictions of the Pre- Combustion Fluid Motion and Combustion Rates in a Spark Ignition Engine", SAE Paper 831697, 1983. Martin, J.K., Witze, P.O., and Borgnakke, C, "Combustion Effects on the Preflame Flow Fields in a Research Engine", SAE Paper 850122, 1985. Barton, R.K., Kenemuth, et al. "Cycle-by-Cycle Variations of A SI En• gine — A Statistical Analysis", SAE Paper 700488, 1970. Gosman, A.D., "Multidimensional Modelling of Cold Flows and Turbu• lence in Reciprocating Engines", SAE Paper 850344, 1985. Soltau, J.P., "Cylinder Pressure Variations in Petrol Engines", Proceed• ings of the Institution of Mechanical Engineers No. 2, 1960-1961. Patterson, D.J., "Cylinder Pressure Variations, A Fundamental Com• bustion Problem", SAE Paper 660129, 1966. Hansel, J.G., "A Turbulent Combustion Model of Cycle-to-Cycle Com• bustion Variations in Spark-Ignition Engines," Combustion Science and Technology, vol 2 pp 223- 225, 1970. Hancock, M.S., Buckingham, D.J., and Belmont, M.R., "The Influence Of Arc Parameters on Combustion in a Spark-Ignition Engine", SAE Paper 860321, 1986. Anderson, R.W., "The Effect Of Ignition System Power on Fast Burn Engine Combustion", SAE Paper 870549, 1987. Kalghatgi, G.T., "Spark Ignition, Early Flame Development and Cyclic Variation In IC Engines", SAE Paper 870163, 1987. Lancaster, D.R., and Kreiger, R.B., et al. "Effects of Turbulence on Spark Ignition Engine Combustion", SAE Paper 760160, 1976. Matekunas, F.A., "Modes and Measures of Cyclic Variability", SAE Transaction vol 92 pp 1139, SAE Paper 830337, 1983. Yamamoto, H., and Misumi, M., "Analysis of Cyclic Combustion Varia• tion in a Lean Operating SI Engine", SAE Paper 870547, 1987. Nagayana, I., Araki, Y., and Lioka, Y., "Effects of Swirl and Squish on SI Engine Combustion and Emmision", SAE Paper 770217, 1977. Kuroda, H., and Nakajima, Y., "The Fast Burn with Heavy EGR, New Approach for Low NOx and Improved fuel Economy", SAE Paper 780006, 1978. Rassweiler, CM., and Withrow, L., "Motion pictures of Engine Flames Correlated with Pressure Cards", SAE Vol 42 no 5 1938. Gatowski, J.A., Heywood, J.B., et al. "Heat Release Analysis of Engine Pressure Data", SAE Paper 841359, 1984. Kreiger, R.B., and Borman, G.L., "The Computation of Apparent Heat Release for Internal Combustion Engines", ASCE 66-WA/DGP-4, 1966. Bibliography 102
[52] Amann, C.A., "Combustion In The Spark-Ignition Engine", Keynote address IMechE International conference on combustion in engines," May 1988. [53] Cole, J.B., and Swords, M.D., "An Investigation of the Ignition Process in a Lean-Burning Engine using Conditionally Sampled Laser Doppler Anemometry," SAE Paper 800043, 1980. [54] Belmont, M.R., Hancock, M.S., and Buckingham, D.J., "Statistical As• pects of Cyclic Variability", SAE Paper 860324, 1986. [55] Daily, J.W., "Cycle-to-Cycle Variations: A Chaotic Process?", SAE Pa• per 870165, 1987. [56] Ma, T.H., "Effect of Cylinder Charge Motion on Combustion", C81/75 IMechE, 1975. [57] Witze, P.O., "The Effect of Spark Location on Combustion in a Variable- Swirl Engine", SAE Paper 820044, 1982. [58] Witze, P.O., and Vilchis,F.R., "Stroboscobic Laser Shadowgraph study of the Effect of Swirl on Homogeneous Combustion in a Spark-Ignition Engine", SAE Paper 810226, 1981 [59] Sheppard, C.W., and Bradley, "Limitations to Turbulence Enhanced Burning Rates in Lean Burn Engines," IMECHE C46/88, 1988. [60] Saxena, V., and Rask R.B., "Influence of Inlet Flows on the Flow Field in an Engine", SAE Paper 870369, 1987. [61] Heywood, J.B., "Combustion Chamber Design for Optimum Spark- Ignition Engine Performance", International Journal of Vehicle Design , vol 5 No.3 pp 133-147, 1984. [62] Gruden, D.O., "Combustion Chamber Layout for Modern Otto En• gines",SAE Paper 811231, 1981. [63] Overington, M.T., and Thring, R.H., "Gasoline Engine Combustion- Turbulence and The Combustion Chamber," Ricardo Consulting En• gineering Ltd. SAE Paper 810017, 1981. [64] Shimoda, M., et al. "Effect of Combustion Chamber Configuration on In-Cylinder Air Motion and Combustion Characteristics of CI Diesel Engine", SAE Paper 850070, 1985. [65] Jane, P.A.H., "The Development of a Direct Injection Diesel Combus• tion System for Low Noise Emmisions and Mechanical Loading", C66/88 IMechE 1988. [66] Evans, R.L., "Internal Combustion Engine Squish Jet Combustion Chamber", USA Patent, No 4,572,123 Feb.25, 1986. [67] Evans, R.L., and Cameron, O, "A New Combustion Chamber for Fast Burn Applications," SAE Paper 860319, 1986. [68] Nakamura ,H., Ohinouye, T., et al. Mitsubishi "Development of a New Combustiono System (MCA-JET) in Gasoline Engine", SAE Paper 790016, 1979. Bibliography 103
[69] Jones, A. "UBC Ricardo Hydra Engine Test Facility", AFL-86-08, 1986. [70] Ricardo Consultant "The Ricardo/Cussons Standard Hydra Engine and Test Bed", 1985. [71] Brown, "Methods for Evaluating Requirements And Errors In Cylinder Pressure Measurement", SAE Paper 670008, 1967. [72] Benson, R.S., and Pick, R., 'Recent Advances In Internal Combustion Engine Instrumentation With Particular Reference to High-Speed Data Acquisition and Automated Test Bed,,' SAE Paper 740695, 1974. [73] Vines, R.F.,"The Platinum Metals and their Alloys", The International Nickel Company, Inc.,New York, New York, 1941. [74] Boisvert, J., "Turbulent Combustion of Gas-Air Mixtures In a Spark Ignition Engine," M.A.Sc Thesis U.B.C. AFL-86-05 1986. [75] Dohring, K., "The Relative Effects of Intake and Compression Stroke Generated Turbulence In I.C. Engine Duration", M.A.Sc. Thesis UBC AFL-86-01 1986. [76] Lancaster D.R., Krieger, R.B. and Liemesch, "Measurement and Analy• sis of Engine Pressure Data", SAE Paper 750026, 1975. [77] Kapil, A. "Cycle-to-Cycle Variations in Spark Ignition Engines", M.A.Sc Thesis, UBC, 1988. [78] Collis, D.C., and Williams, M.J. "Two- Dimensional Convection from Heated Wires at Low Reynolds Number", J. of Fluid Mechanics Vol 6 pp 357-384, 1959. [79] Davies, P.O.A.L., and Fisher, M.J., "Heat Transfer From Electrically Heated Cylinders", Proc. Roy. Soc. A., Vol 280 pp 486-526, 1964. [80] Rask, R.B., "Comparision of Window Smoothed- Ensembled and Cycle- by-Cycle Data Reduction Techniques for Laser Doppler Anemometry Measurements of In-Cylinder Velocity", ASME 1981. [81] Fransler, T.D., "Laser Velocimetry Measurements of Swirl and Squish Flows in an Engine with a Cylindrical Piston Bowl", SAE Paper 850124, 1985. [82] Catania, A.E., and Mittica, A., "A Contribution To The Definition And Measurement Of Turbulence In A reciprocating IC Engine", 85DPG12, 1985. [83] Bury, K.V., Statistical Models in Applied Science John Wiley and Sons 1975. Tables 104
ENGINE: Number of cylinders 1 Bore 80.26 mm Stroke 88.9 mm Swept Volume 0.45 litres Maximum Speed 90 rps Maximum Power 15 kW
Compression ratio (nominal) 9 : 1
VALVE ARRANGEMENT Overhead cam shaft verticle lift 9 mm IVO: Inlet Valve Opens 12° BTDC IVC: Inlet Valve Closes 56° ABDC EVO: Exhaust Valve opens 56° BBDC EVC: Exhaust Valve Closes 12° ATDC
IGNITION SYSTEM Type Lumenition Coil Lucas SP 12 Timing range 70° BTDC to 20° ATDC Spark Plug Champion N6Y or AGYC
Table 3.1: Ricardo Hydra Gasoline (or Gaseous fuel) Engine Specifications. Tables 105
PISTON o: PRESSURE •: HOTWIRE 44 cycles SPEED 20.0 (rps) 33.3 (rps) 50.0 (rps) 66.7 (rps) std bathtub 0 0 0 single slot 0 • 0 • 0 • castellated 0 • 0 • o bowl/piston 0 • 0 • 0 • 0 squish jet 0 • 0 • 0 std disc 0 • 0 • • 0 • • • Table 3.2: Motored operating conditions for Pressure and Hotwire measurements at WOT.
PISTON PRESSURE o: 44 cycles • : 200 cycles SPEED 20.0 (rps) 33.3 (rps) 50.0 (rps) 66.7 (rps) RAFR 1.00 1.27 1.00 1.27 1.00 1.27 1.00 1.27 std bathtub 0 0 0 0 0 0 single slot castellated • • • • • • bowl/piston • • • • • • squish jet •0 •0 •o •o •0 •0 std disc 0 0 0 0 o 0
Table 3.3: Fired operating conditions for Pressure measurements at MBT and Full Load (WOT).
PISTON PRESSURE o: 44 cycles •: 200 cycles SPEED 33.3 (rps) 50.0 (rps) BMEP 2.5 (bar) 3.5 (bar) RAFR 1.00 1.27 1.00 1.27 std bathtub single slot castellated • • • • bowl/piston • • • • squish jet •o •0 • • std disc 0 0
Table 3.4: Fired operating conditions for Pressure measurements at MBT and Part Load. Tables 106
MOTORED PRESSURE HOTWIRE ANALYSIS ANALYSIS
HOTWIRE HOTWIRE HW-Cal
ISAAC2VAX ISAAC2VAX
PRESS- ANAL
EXTRACTION EXTRA CTION 1 SMTH--AVE
TEMP--1800
HW-ANAL
TURBULENCE
Table 4.1: Motored data Analysis program flow chart. Tables 107
FIRED PRESSURE FIRED PRESSURE PERFORMANCE ANALYSIS: 200 cycles ANALYSIS: 44 cycles ANALYSIS
HOTPRES2 HOTPRES DATAQ
ISAC22VAX ISAAC2VAX
FIRERUN FIRERUN-44 CRUNCH
MASSBURN-ALL MASSBURN-ALL
Table 4.2: Fired data Analysis program flow charts. Tables 108
PISTON Compression coefRceint jc Expansion coefficient 7e std bathtub 1.34 1.3.5 single slot 1.33 1.35 castellated 1.32 1.35 bowl/piston 1.30 1.35 squish jet 1.27 1.35 std disc 1.37 1.35
Table 5.1: Compression and Expansion coefficients for the motored condition. Tables 109
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. std bathtub 3.20 7.10 256 28.9 25.4 24 1.02 single slot 3.09 6.87 276 26.9 24.6 24 1.03 castellated 2.70 5.99 298 24.9 21.5 22 1.03 bowl/piston 3.40 7.60 240 30.8 27.1 11 1.02 squish jet 2.75 6.10 294 25.2 22.0 16 1.03 std disc 3.11 6.92 262 28.3 24.8 19 1.04
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. std bathtub 2.60 5.94 248 29.9 21.3 29 1.30 single slot 2.80 6.16 250 29.6 22.1 28 1.30 castellated 2.36 5.27 285 26.0 18.9 31 1.30 bowl/piston 2.96 6.57 226 32.8 23.5 21 1.30 squish jet 2.33 5.20 284 26.1 18.6 24 1.31 std disc 2.77 6.17 243 30.5 22.0 29 1.30
Table 5.2: Engine performance as per SAEJ1349 for different piston geometries for stoi• chiometric and lean RAFR at MBT, WOT, and 20.0 rps. Tables 110
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. std bathtub 5.30 7.12 254 29.2 25.5 35 1.02 single slot 5.18 6.90 272 27.2 24.7 30 1.01 castellated 5.09 6.77 272 27.3 24.2 26 1.01 bowl/piston 5.98 7.94 241 30.7 28.4 12 1.01 squish jet 4.81 6.40 288 25.8 22.9 20 1.01 std disc 5.64 7.50 248 29.9 26.9 32 1.02
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. std bathtub 4.55 6.12 245 30.3 21.9 38 1.28 single slot 4.80 6.50 234 31.6 23.1 31 1.28 castellated 4.47 5.96 251 29.5 21.3 33 1.29 bowl/piston 5.27 7.01 219 33.7 25.1 20 1.28 squish jet 4.17 5.54 272 27.3 19.8 27 1.31 std disc 4.78 6.39 238 31.2 22.9 33 1.28
Table 5.3: Engine performance as per SAEJ1349 for different piston geometries for stoi• chiometric and lean RAFR at MBT, WOT, and 33.3 rps.
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. single slot 7.98 7.08 262 28.4 25.4 33 1.01 castellated 7.49 6.67 279 26.5 23.9 31 1.01 bowl/piston 8.77 7.75 245 30.2 27.8 16 1.02 squish jet 7.05 6.29 299 24.8 22.5 21 1.01
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. single slot 7.40 6.50 230 32.3 23.4 37 1.27 castellated 6.94 6.17 249 29.8 22.1 37 1.27 bowl/piston 7.82 7.00 222 33.4 24.9 22 1.27 squish jet 6.14 5.47 279 26.6 19.6 28 1.27
Table 5.4: Engine performance as per SAEJ1349 for different piston geometries for stoi• chiometric and lean RAFR at MBT, WOT, and 50.0 rps. Tables 111
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. std bathtub 2.02 2.70 349 21.3 9.7 35 1.05 single slot 1.93 2.56 348 21.3 9.2 31 1.05 castellated 1.79 2.38 371 20.0 8.5 31 1.03 bowl/piston 1.90 2.51 308 24.0 9.0 23 1.04 squish jet 1.90 2.54 339 21.9 9.1 26 1.05 std disc 1.84 2.48 373 19.9 8.9 32 1.06
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. std bathtub 2.20 3.00 355 20.9 10.5 35 1.27 single slot 1.94 2.56 331 22.4 9.2 33 1.31 castellated 1.91 2.55 364 20.3 9.1 34 1.30 bowl/piston 1.92 2.54 291 25.4 9.1 24 1.31 squish jet 1.92 2.54 325 22.8 9.1 29 1.27 std disc 1.75 2.35 350 21.1 8.4 32 1.25
Table 5.5: Engine performance as per SAEJ1349 for different piston geometries for stoi• chiometric and lean RAFR at MBT, 2.5 bmep, and 33.3 rps.
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. single slot 3.88 3.45 314 24.4 12.4 37 1.02 castellated 3.97 3.53 321 23.1 12.6 36 1.02 bowl/piston 3.78 3.36 305 24.3 12.0 24 1.03
bp bmep bsfc Vth Torq. Ign.Adv RAFR PISTON kW bar g/kWhr % Nm deg. single slot 3.88 3.46 281 26.3 12.4 41 1.28 castellated 4.01 3.56 300 24.7 12.7 40 1.28 bowl/piston 3.92 3.48 277 26.7 12.4 25 1.28
Table 5.6: Engine performance as per SAEJ1349 for different piston geometries for stoi• chiometric and lean RAFR at MBT, 3.5 bmep, and 50.0 rps. Tables 112
RAFR 1.02 1.06 1.12 1.17 1.22 1.27 1 1.32 PISTON IGNITION ADVANCE (degrees btdc) std bathtub 24 26 27 28 29 29 31 single slot 20 24 27 29 29 28 28 castellated 22 25 28 29 29 31 32 bowl/piston 11 14 15 18 21 21 20 squish jet 16 19 21 22 23 24 25 std disc 19 21 26 28 28 29 30
RAFR 1.02 1.06 1.12 1.17 1.22 1.27 1.32 PISTON BRAKE THERMAL EFFICIENCY (%) std bathtub 28.9 29.7 29.6 30.1 29.7 29.9 29.3 single slot 26.9 28.0 28.6 29.2 29.4 29.6 29.6 castellated 24.9 25.4 25.5 26.0 26.0 26.0 26.0 bowl/piston 30.8 31.7 32.1 32.2 32.9 32.8 32.9 squish jet 25.2 25.5 26.2 26.1 25.8 26.1 26.0 std disc 28.3 29.4 30.0 30.4 30.6 30.5 30.5
Table 5.7: Ignition Advance and Brake Thermal Efficiency for different piston geometries for RARF^l.00-1.35 at MBT, WOT, and 20.0 rps. Tables 113
RAFR 1.02 1.06 1.12 1.17 1.22 1.27 1.32 PISTON IGNITION ADVANCE (degrees btdc) std bathtub 35 36 37 36 38 38 40 single slot 30 30 30 31 31 31 31 castellated 26 30 31 31 32 33 34 bowl/piston 12 15 18 19 19 20 21 squish jet 20 24 26 26 27 27 29 std disc 32 36 36 37 33 33 34
RAFR 1.02 1.06 1.12 1.17 1 1.22 1.27 1.32 PISTON BRAKE THERMAL EFFICIENCY (%) std bathtub 29.2 30.0 30.0 30.5 30.4 30.3 30.1 single slot 27.2 28.8 29.6 30.5 31.1 31.6 31.7 castellated 27.3 28.5 29.2 29.6 29.8 29.5 29.3 bowl/piston 30.7 32.3 32.9 33.3 33.4 33.7 33.9 squish jet 25.8 26.7 26.9 27.1 27.2 27.3 27.1 std disc 30.0 30.8 31.5 31.5 31.8 31.2 30.9
Table 5.8: Ignition Advance and Brake Thermal Efficiency for different piston geometries for RARF «1.00-1.35 at MBT, WOT, and 33.3 rps.
RAFR 1.02 1.06 1.12 1.17 1.22 1.27 1.32 PISTON IGNITION ADVANCE (degrees btdc) single slot 33 32 33 35 35 37 38 castellated 31 32 34 36 35 37 38 bowl/piston 16 19 21 21 21 22 23 squish jet 21 24 26 26 28 28 31
RAFR 1.02 1.06 1.12 1.17 1 1.22 1 1.27 1.32 PISTON BRAKE THERMAL EFFICIENCY (%) single slot 28.4 29.6 30.7 31.4 32.2 32.3 32.7 castellated 26.5 28.2 29.0 29.7 29.8 29.8 29.7 bowl/piston 30.2 31.8 32.5 33.2 33.2 33.4 33.7 squish jet 24.8 25.5 26.1 26.2 26.5 26.6 26.5
Table 5.9: Ignition Advance and Brake Thermal Efficiency for different piston geometries for RARF wl.00-1.35 at MBT, WOT, and 50.0 rps. Tables 114
IMEP t1 tr COV CT COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RA FR=1.27 std bathtub 818.7 12.8 1.58 701.5 20.4 2.94 single slot 821.7 14.1 1.72 764.2 13.6 1.79 castellated 808.5 13.0 1.61 726.2 14.1 1.95 bowl/piston 788.1 14.6 1.86 697.1 13.8 1.98 squish jet 768.9 10.8 1.41 673.9 13.3 1.99 std disc 826.9 8.0 0.98 743.8 19.0 2.58
PEAK tT COV tT COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 3957 318 8.1 3237 402 12.6 single slot 4032 316 7.8 3819 344 9.0 castellated 4030 263 6.6 3802 337 8.9 bowl/piston 4418 738 16.7 4361 201 4.6 squish jet 3527 124 3.6 3169 168 5.3 std disc 3918 388 10.0 3761 441 11.9
IPEAK tT COV tT COV c.a. c.a. % c.a. c.a. % PISTON RAFR=1.00 RAFR=1.27 std bathtub 16.9 2.1 12.7 16.2 2.1 13.3 single slot 20.4 3.0 14.8 19.0 3.0 15.9 castellated 18.6 2.3 12.4 16.8 2.9 17.3 bowl/piston 12.6 5.8 46 12.0 1.4 11.7 squish jet 15.5 1.4 9.0 14.3 1.8 12.5 std disc 19.0 2.6 13.7 16.6 2.1 12.6
Table 5.10: Imep, peak pressure and angle of occurance of peak pressure for different piston geometries for stoichiometric and lean RAFR at, MBT, WOT, and 20.0 rps. Tables 115
IMEP f1 COV COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RA FR=1.27 std bathtub 859.9 16.6 1.95 763.9 28.8 3.81 single slot 878.7 12.7 1.45 813.3 26.0 3.20 castellated 841.8 14.4 1.72 757.1 31.8 4.21 bowl/piston 867.9 13.6 1.57 745.5 24.8 3.34 squish jet 803.7 19.6 2.46 726.4 23.5 3.27 std disc 875.6 12.4 1.44 777.8 25.6 3.33
PEAK cr COV /* cr COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 4331 341 8.0 3729 378 10.3 single slot 4396 316 7.2 3965 436 11.0 castellated 3930 292 7.4 3631 450 12.4 bowl/piston 3878 943 24.4 4371 259 5.9 squish jet 3383 371 11.1 3204 245 7.7 std disc 4224 459 11.0 3990 597 15.1
IPEAK cr COV cr COV ca. ca. % ca. ca. % PISTON RAFR=1.00 RA FR=1.27 std bathtub 14.1 2.1 14.9 13.5 1.8 13.3 single slot 16.9 2.7 16.0 17.7 3.2 18.1 castellated 19.5 2.8 14.3 17.8 3.1 17.4 bowl/piston 14.0 9.8 70 13.1 1.8 13.7 squish jet 15.5 4.4 28 14.9 2.3 15.4 std disc 16.0 2.0 12.5 14.4 2.4 16.6
Table 5.11: Imep, peak pressure and angle of occurance of peak pressure for different piston geometries for stoichiometric and lean RAFR at, MBT, WOT, and 33.3 rps. Tables 116
IMEP V- cr COV f1 tr COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RA FR=1.27 std bathtub 884.1 17.0 1.92 800.3 36.0 4.53 single slot 906.7 18.8 2.08 841.0 36.7 4.37 castellated 864.0 17.5 2.03 790.7 36.6 4.64 bowl/piston 888.6 15.7 1.77 775.4 28.8 3.72 squish jet 772.5 18.2 2.38 722.8 42.5 5.94
PEAK f1 COV cr COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RA FR=1.27 std bathtub 4910 276 5.7 4316 553 12.9 single slot 4448 335 7.6 4209 433 10.3 castellated 4009 392 9.8 3811 460 12.1 bowl/piston 4333 458 10.6 4346 300 6.9 squish jet 3345 176 5.3 3298 247 7.6
IPEAK cr COV COV ca. ca. % ca. ca. % PISTON RAFR=1.00 RAFR=1.27 std bathtub 12.1 1.9 15.7 12.8 2.4 18.8 single slot 16.4 2.7 16.5 15.6 2.6 16.7 castellated 16.0 3.5 21.9 16.9 3.0 17.8 bowl/piston 13.9 4.6 33.1 10.4 3.6 35 squish jet 15.5 2.5 16.1 12.7 3.5 27
Table 5.12: Imep, peak pressure and angle of occurance of peak pressure for different piston geometries for stoichiometric and lean operation at, MBT, WOT, and 50.0 rps. Tables 117
IMEP COV COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub unstable unstable single slot 422.6 11.3 2.7 414.3 24.1 5.8 castellated 400.0 27.6 6.9 394.3 31.7 8.0 bowl/piston 381.0 6.2 1.6 384.9 7.78 2.0 squish jet 397 11.7 3.0 414 12.7 3.1 std disc 418 127 30 400 31 7.9
PEAK CT COV COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub unstable unstable single slot 1586 207 13.0 1517 180 11.9 castellated 1395 212 15.2 1372 179 13.0 bowl/piston 2009 133 6.6 1999 112 5.6 squish jet 1539 329 21.6 1707 170 10 std disc 1698 326 19 1524 230 15
IPEAK o~ COV V- COV ca. ca. % ca. ca. % PISTON RAFR=1.00 RAFR=1.27 std bathtub unstable unstable single slot 21.7 4.7 21.7 15.5 5.1 33 castellated 13.5 7.4 55 12.8 7.2 56 bowl/piston 15.4 2.7 17.5 16.9 1.4 8 squish jet 14.0 8.1 58 17.5 3.1 17.8 std disc 17.5 6.4 36 13.9 4.3 31
Table 5.13: Imep, peak pressure and angle of occurance of peak pressure for different piston geometries for stoichiometric and lean RAFR at, MBT, 2.5 bmep, and 33.3 rps. Tables 118
IMEP f1 cr COV cT COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 single slot 539 10.2 1.9 526 20.1 3.8 castellated 547 22.2 4.06 544 33.7 6.2 bowl/piston 506 8.5 1.7 501 13.3 2.7
PEAK cr COV V- or COV kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 single slot 2349 251 10.7 2325 312 13.4 castellated 2271 281 12.4 2126 325 15.3 bowl/piston 2523 127 5.0 2517 171 6.8
IPEAK V- tr COV tT COV c.a. c.a. % c.a. c.a. % PISTON RAFR=1.00 RAFR=1.27 single slot 18.5 2.8 15.1 17.1 2.5 14.6 castellated 19.9 3.3 16.6 16.6 4.1 24.7 bowl/piston 12.2 3.3 27 15.7 2.5 15.9
Table 5.14: Imep, peak pressure and angle of occurance of peak pressure for different piston geometries for stoichiometric and lean RAFR at, MBT, 3.5 bmep, and 50.0 rps. Tables 119
INITIAL a COV cr COV 0-01% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 10.2 1.2 11.9 14.1 1.6 11.3 single slot 9.5 0.9 9.7 13.1 1.4 10.4 castellated 9.4 1.0 10.6 13.5 1.8 13.6 bowl/piston 5.4 1.4 27 10.2 1.4 13.5 squish jet 7.8 0.9 12.2 12.0 1.5 12.9 std disc 11.3 1.4 12.6 15.0 1.6 10.7
MAIN t1 cr COV f1 cr COV 01-90% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 35.1 3.2 9.1 44.1 6.4 14.7 single slot 33.8 2.3 6.9 38.4 2.7 7.0 castellated 33.0 2.1 6.5 37.4 2.7 7.2 bowl/piston 19.5 2.5 13 22.4 3.0 13 squish jet 28.5 5.7 20 33.1 7.2 22 std disc 34.0 3.3 9.8 37.6 4.5 12.2
Table 5.15: Initial (0-01% massburned) and Main (01-90% massburned) combustion durations for different piston geometries for stoichiometric and lean RAFR at MBT, WOT, and 20.0 rps. Tables 120
INITIAL COV CT COV 0-05% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 12.2 1.4 11 16.3 1.7 10.6 single slot 11.5 1.0 8.8 15.7 1.5 9.3 castellated 11.3 1.1 9.4 16.1 1.8 11.1 bowl/piston 6.3 1.2 19 11.4 1.3 11.1 squish jet 9.4 1.0 10.4 14.4 1.5 10.5 std disc 13.1 1.7 13.3 17.2 1.9 11.2
MAIN COV cr COV 05-90% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 33.1 3.0 9.2 41.8 6.3 15.2 single slot 31.8 2.3 7.1 35.8 2.6 7.3 castellated 31.1 2.1 6.6 34.8 2.6 7.4 bowl/piston 18.6 2.4 13 21.2 3.1 14.6 squish jet 26.9 5.7 21 30.8 7.1 23 std disc 32.2 3.0 9.5 35.3 4.4 12.6
Table 5.16: Initial (0-05% massburned) and Main (05-90% massburned) combustion duration for different piston geometries for stoichiometric and lean RAFR at MBT, WOT, and 20.0 rps. Tables 121
INITIAL f1 tT COV tT COV 0-01% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 15.8 2.1 13.2 18.4 2.5 14 single slot 13.9 1.4 9.8 15.1 2.0 13.1 castellated 11.4 2.0 18 18.0 2.8 15.5 bowl/piston 8.8 1.3 14.5 10.7 3.0 28 squish jet 12.0 1.6 14 15.2 1.9 12.9 std disc 16.2 2.7 17 17.5 3.4 20
MAIN cr COV f1 cr COV 01-90% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 40.4 4.4 11.0 47.0 5.7 12.2 single slot 36.0 2.2 6.1 39.9 3.2 7.9 castellated 36.8 2.8 7.6 39.9 3.5 8.7 bowl/piston 24.4 5.7 24 26.2 7.8 30 squish jet 26.9 4.2 16 31.9 6.2 20 std disc 38.8 5.0 13.1 43.0 6.7 15.9
Table 5.17: Initial (0-01% massburned) and Main (01-90% massburned) combustion duration for different piston geometries for stoichiometric and lean RAFR at MBT, WOT, and 33.3 rps. Tables 122
INITIAL tT COV tT COV 0-05% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 18.3 1.7 9.2 21.2 2.1 10.1 single slot 16.4 1.5 8.9 18.1 1.9 10.3 castellated 14.6 1.5 10.4 20.8 2.4 11.5 bowl/piston 10.2 1.2 11.4 12.4 2.4 19 squish jet 13.7 1.3 9.5 17.8 1.8 9.9 std disc 18.6 2.3 12.5 20.8 3.5 17
MAIN tT COV tT COV 05-90% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 37.9 4.5 11.9 44.2 8.2 12.4 single slot 33.5 2.1 6.3 36.9 2.9 7.8 castellated 33.6 2.2 6.7 37.0 3.1 8.5 bowl/piston 23.1 5.8 25 24.4 7.9 32 squish jet 25.2 4.2 17 29.2 6.4 22 std disc 36.4 4.8 13.2 39.7 6.9 18
Table 5.18: Initial (0-05% massburned) and Main (05-90% massburned) combustion duration for different piston geometries for stoichiometric and lean RAFR at MBT, WOT, and 33.3 rps. Tables 123
INITIAL COV 0~ COV 0-01% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 19.2 3.2 17 21.3 3.3 16 single slot 15.9 2.2 14 13.9 8.5 61 castellated 15.3 3.8 25 19.9 10.7 54 bowl/piston 11.5 1.9 17.0 11.3 3.2 28 squish jet 12.2 1.3 10.7 15.1 3.1 21
MAIN cr COV f1 cr COV 01-90% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 39.8 4.6 11.5 43.0 5.4 12.6 single slot 38.4 2.7 7.2 44.5 8.6 19.4 castellated 38.3 3.9 10.0 44.4 9.0 20.2 bowl/piston 22.1 5.1 23 23.9 8.4 35 squish jet 26.8 3.6 14 30.3 7.3 24
Table 5.19: Initial (0-01% massburned) and Main (01-90% massburned) combustion duration for different piston geometries for stoichiometric and lean RAFR at MBT, WOT, and 50.0 rps. Tables 124
INITIAL cr COV tT COV 0-05% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 22.1 2.0 9.1 24.0 2.6 10.9 single slot 19.2 1.7 8.9 20.3 4.0 19.8 castellated 18.2 4.1 23 23.1 9.1 39 bowl/piston 13.0 1.6 12.2 13.1 2.8 22 squish jet 14.7 1.2 8.4 17.9 2.3 13
MAIN tT COV cr COV 05-90% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub 36.9 4.1 11.2 40.3 5.2 12.8 single slot 35.1 2.6 7.5 38.1 4.3 11.3 castellated 35.5 4.1 11.5 41.2 7.8 18.9 bowl/piston 20.6 5.2 25 22.1 8.4 38 squish jet 24.2 3.4 14 27.5 7.0 26
Table 5.20: Initial (0-05% massburned) and Main (05-90% massburned) combustion duration for different piston geometries for stoichiometric and lean RAFR at MBT, WOT, and 50.0 rps. Tables 125
INITIAL V- COV cr COV 0-01% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub unstable unstable single slot 18.8 2.9 15.4 20.5 6.0 29.3 castellated 21.0 4.7 22.6 23.0 5.9 25.9 bowl/piston 13.9 1.4 9.8 16.0 1.5 9.2 squish jet 16.9 2.8 16.5 18.3 3.7 20.7 std disc 20.6 4.2 20.5 24.6 38.8 16.0
MAIN cr COV a COV 01-90% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 std bathtub unstable unstable single slot 45.5 5.0 11.1 53.7 8.4 15.6 castellated 52.0 9.2 17.6 53.3 10.1 19.0 bowl/piston 25.9 1.6 6.4 28.5 2.1 7.5 squish jet 30.6 4.0 13.1 37.2 6.0 16.3 std disc 48.1 7.2 15.2 51.8 6.7 13.1
Table 5.21: Initial (0-01% massburned) and Main (01-90% massburned) combustion duration for different piston geometries for stoichiometric and lean RAFR at MBT, 2.5 bmep, and 33.3 rps. Tables 126
INITIAL t1 cr COV cr COV 0-05% kPa kPa % kPa kPa % PISTON RA FR=1.00 RAFR=1.27 std bathtub unstable unstable single slot 22.2 2.2 10.0 24.1 5.0 21.0 castellated 24.2 4.4 18.1 26.8 4.9 18.5 bowl/piston 15.6 1.1 7.2 17.8 1.2 7.1 squish jet 18.9 2.2 11.7 20.7 2.2 10.8 std disc 23.71 4.3 18.3 27.3 2.9 10.7
MAIN • COV tT COV 05-90% kPa kPa % kPa kPa % PISTON RA FR=1.00 RAFR-1.27 std bathtub unstable unstable single slot 42.1 4.6 11.0 50.1 7.8 15.7 castellated 48.8 8.9 18.3 49.5 9.8 19.8 bowl/piston 24.2 1.6 6.6 26.7 2.1 7.7 squish jet 28.6 4.0 14.3 34.8 6.3 18.3 std disc 45.0 6.3 14.2 49.0 5.8 12.1
Table 5.22: Initial (0-05% massburned) and Main (05-90% massburned) combustion duration for different piston geometries for stoichiometric and lean at RAFR MBT, 2.5 bmep, and 33.3 rps. Tables 127
INITIAL cr COV cr COV 0-01% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 single slot 18.3 3.6 19 17.2 6.6 38 castellated 21.7 8.5 39 27.5 7.8 28 bowl/piston 13.9 3.3 24 18.3 4.5 25
MAIN cr COV cr COV 01-90% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 single slot 44.3 4.0 8.9 51.8 6.7 13.0 castellated 44.1 6.3 14 49.7 7.8 15.8 bowl/piston 26.8 5.1 19 29.8 5.5 18.4
Table 5.23: Initial (0-01% massburned) and Main (01-90% massburned) combustion duration for different piston geometries for stoichiometric and lean RAFR at MBT, 3.5 bmep, 50.0 rps.
INITIAL cr COV /* cr COV 0-05% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 single slot 22.7 2.3 10.1 24.1 5.4 23 castellated 24.2 8.6 36 30.7 6.2 20 bowl/piston 16.7 3.0 18 20.1 4.6 23
MAIN V- cr COV A* cr COV 05-90% kPa kPa % kPa kPa % PISTON RAFR=1.00 RAFR=1.27 single slot 40.0 3.3 8.3 44.9 6.1 13.5 castellated 41.7 6.8 16 46.6 7.3 15.8 bowl/piston 24.0 5.0 21 28.0 5.7 20
Table 5.24: Initial (0-05% massburned) and Main (05-90% massburned) combustion duration for different piston geometries for stoichiometric and lean RAFR at MBT, 3.5 bmep, and 50.0 rps. 5 B COOLING n NAT. GAS =£c $ "ATWATEI R cn GASOLINE =£=
control ^ signals AVL — OPTICAL OIL COOLING HYDRA TACHO IETER MODULE. ENGINE DYNAMOMETER sh.n CONVERTER CABINET
Instilment TORQUE signalj WATER TT r» out LOAD CELL control signals ,n TEST CELL power at In or f out CONTROL ROOM
IBM A/D SYSTEM CONTROL "PT -IN COMPUTER TRANSFORMER CONSOLE analog slgnalsV . control *~ signals out ^digital "out ) ^ tnstruawnt * signals in CIRCUIT BOARD HOUSING POWER •—conditioned SUPPLY Instrument signals " signals out out G 1
Figure 3.1: Ricardo Hydra engine, dynamometer and control systems layout. to oo Figures 129
Figure 3.2: Ricardo Hydra MKIII Gasoline (or gaseous fuel) Engine cross-sectional and longitudinal views. DETAIL OF COMBUSTION CHAMBER Figures 131
SECTION T-T
Figure 3.4: Single slot and castellated piston geometri Figures 132
Piston no.7
CR « 9.0:1
Figure 3.5: Bowl-in-piston and squish jet piston geometries. Figures 133
Figure 3.6: Instrumentation layout for the Ricardo engine test cell. Figure 3.7: Hot wire probe location through the spark plug entry for the bathtub and flat cylinder heads. Figures 135
RS232 Un* to PC
Figure 3.8: Acquisition hardware arrangement with fast pressure data acquisition hook up. Figures 136
Figure 4.1: Comparison of the effect of window size on the turbulent intensity profile for the bathtub chamber at WOT, 33.3 rps. Figures 137
-90 -60 -30 0 30 60 CRANK ANGLE DEGREES FROM TDC
Figure 4.2: Polytropic coefficent calculated from the ensembled pressure for single slot chamber at MBT, WOT and 33.3 rps for RAFR=1.27. Figures 138
2000 std bathtub single slot caste 11ated bowl/piston squish jet ' '• std disc ,iff \\ 1500 if X 1 V 'ii V '/''// 1000 \ \ 500 \
jJL-— 0 .
1 1 1 1 1 1 1 1 1 1 1 -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180
CRANK ANGLE DEGREES FROM TDC
Figure 5.1: Motored pressure profiles for different chamber geometries at WOT, 20.0 rps.
Figure 5.2: Motored pressure profiles for different chamber geometries at WOT, 33.3 rps. Figures 139
Figure 5.3: Motored pressure profiles for different chamber geometries at WOT, 50.0 rps.
CRANK ANGLE DEGREES FROM TDC
Figure 5.4: Motored pressure profiles for different chamber geometries at WOT, 66.7 rps. Figures 140
Figure 5.5: Motored pressure profiles for the single slot piston at WOT, for three speeds; 20.0, 33.3, and 50.0 rps.
800
std bathtub single slot 700 castellated bowl/piston \ squish Jet i \ std disc if% 600 $'if % \ i \ 500 //'/ •/ \ /' \ 400
300
-180 -150 -120 -90 -60 -30 30 60 90 120 150 180
CRANK ANGLE DEGREES FROM TDC
Figure 5.6: Motored temperature profiles for different chamber geometries at WOT, 33.3 rps. Figures 141
20
CRANK ANGLE DEGREES FROM TDC
10
CRANK ANGLE DEGREES FROM TDC
Figure 5.7: Window ensembled and cycle ensembled mean velocity and turbulent inten• sity profiles for the single slot piston at WOT, 33.3 rps. Figures 142
Figure 5.8: Mean velocity profiles for different chamber geometries at WOT, 33.3 rps.
sta; bathtub single slot
-180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC
Figure 5.9: Turbulent intensity profiles for different chamber geometries at WOT, 33 3 rps. Figures 143
20
Figure 5.10: Mean velocity profiles for the 'bathtub' group of chambers at WOT, 20.0 rps.
5 _, '. ,
CRANK ANGLE DEGREES FROM TDC
Figure 5.11: Turbulent intensity profiles for the 'bathtub' group of chambers at WOT, 20.0 rps. Figures 144 20
18 std bathtub single slot castellated 16
14 .
1 1 1 1 1 1 1 1 1 1 r r~ -180 -150 -120 -90 -60 -90 0 30 60 90 120 150 CRANK ANGLE DEGREES FROM TDC
Figure 5.12: Mean velocity profiles for the 'bathtub' group of chambers at WOT, 33.3 rps.
5
Figure 5.13: Turbulent intensity profiles for the 'bathtub' group of chambers at WOT, 33.3 rps. Figures 145
20
18 — std disc — bowl/piston — squish Jet 16
14
12
H 10 o o a•J 8 > 6
4
2
0 i 1 1 1 1 1 1 —r r -180 -150 -120 -90 -60 -30 0 30 60 .90 120 150 CRANK ANGLE DEGREES FROM TDC Figure 5.14: Mean velocity profiles for the 'disc' group of chambers at WOT, 20.0 rps.
std d1sc bowl/piston squish Jet
I _| , "I-1-" j !——--( j. .fl j— -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 CRANK ANGLE DEGREES FROM TDC
Figure 5.15: Turbulent intensity profiles for the 'disc' group of chambers at WOT, 20.0 rps. Figures 146
Figure 5.16: Mean velocity profiles for the 'disc' group of chambers at WOT, 33.3 rps.
Figure 5.17: Turbulent intensity profiles for the 'disc' group of chambers at WOT, 33.3 rps. Figures 147
20
18 20.,0-srps 33,: 3 rps 66.7 rps
r- H o o >
30 60 90 120 150 CRANK ANGLE DEGREES FROM TDC Figure 5.18: Mean velocity profiles for the bathtub chamber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps.
'. 20.0 rps \ 33.3 rps \ 66.7 rps
I i i i i i i i 1 1 r 1 -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 CRANK ANGLE DEGREES FROM TDC Figure 5.19: Turbulent intensity profiles for the bathtub chamber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps. Figures 148
20
18 20.0 rps 33.3 rps 50.0 rps 16 .
14
12
30 60 90 120 150
CRANK ANGLE DEGREES FROM TDC
Figure 5.20: Mean velocity profiles for the castellated chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps.
Figure 5.21: Turbulent intensity profiles for the castellated chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps. CRANK ANGLE DEGREES FROM TDC
Figure 5.23: Turbulent intensity profiles for the squish jet chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps. Figures 150 20
18 — 20.0 rps -•' 33.3 rps — 66.7 rps 16
14
12
V IVC H 10 o EVO • o 8 > 6
4
2
0
-180 -150 -120 -90 -60 -30 0 30 60 90 CRANK ANGLE DEGREES FROM TDC
Figure 5.24: Mean velocity profiles for the disc chamber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps.
20.0 rps 33.3 rps —66.7 rps
1 I i I I I I I l I r l -180 -150 -120 -90 -60 -30 0 30 60 90 120 150
CRANK ANGLE DEGREES FROM TDC
Figure 5.25: Turbulent intensity profiles for the disc chamber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps. Figures 151 2.5
-180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC
Figure 5.26: Mean velocity profiles scaled with mean piston speed for the bathtub cham• ber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps.
20.0 rps 33.3 rps 66.7 rps 0.8
0.6 _
0.4
0.2
CRANK ANGLE DEGREES FROM TDC
Figure 5.27: Turbulent intensity profiles scaled with mean piston speed for the bathtub chamber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps. Figures 152
1 1 h 1 1 1 1 1 1 T r i I -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.28: Mean velocity profiles scaled with mean piston speed for the single slot chamber at WOT, for two speeds; 20.0 and 33.3 rps.
l
5 o i !
1 f- i 1 1 1 1 1 1 1 1 i r 1 -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 Figure 5.29: Turbulent intensitCRANy Kprofile ANGLsE scaleDEGREEd witS FROh meaM TDn Cpisto n speed for the single slot chamber at WOT, for two speeds; 20.0 and 33.3 rps. -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180
CRANK ANGLE DEGREES FROM TDC
Figure 5.30: Mean velocity profiles scaled with mean piston speed for the castellated chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps.
l
CRANK ANGLE DEGREES FROM TDC
Figure 5.31: Turbulent intensity profiles scaled with mean piston speed for the castellated chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps. Figures 154 2.5
20.0 rps 33.3 rps
a a a a. co IVC Z 1.5 EVO o H co Z < s
H O O 0.5 H >
II -180 -120 -90 -60 120 150 180 CRANK ANGLE DEGREES FROM TDC
Figure 5.32: Mean velocity profiles scaled with mean piston speed for the bowl-in-piston chamber at WOT, for two speeds; 20.0 and 33.3 rps.
D 20.0 rps a 33.3 rps B OH 0.8 . cn z o H
0.6 CRANK ANGLE DEGREES FROM TDC Figure 5.33: Turbulent intensity profiles scaled with mean piston speed for the bowl-in-piston chamber at WOT, for two speeds; 20.0 and 33.3 rps. Figures 155 1 1 1 1 1 1 1 1 1 1 r i I -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.34: Mean velocity profiles scaled with mean piston speed for the squish jet chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps. g 0.2 . a ^> D CQ -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.35: Turbulent intensity profiles scaled with mean piston speed for the squish jet chamber at WOT, for three speeds; 20.0, 33.3 and 50.0 rps. 1 1 1 1 1 1 1 1 1 1 r 1 I -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.36: Mean velocity profiles scaled with mean piston speed for the disc chamber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps. -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.37: Turbulent intensity profiles scaled with mean piston speed for the disc chamber at WOT, for three speeds; 20.0, 33.3 and 66.7 rps. Figures 157 34 o—° std bathtub single slot B---e castel lated 32 o Z 30 a h. 28 a 20 1 T 1.1 1.15 1.2 1.25 1.3 1.35 1.4 1.05 RELATIVE AIR FUEL RATIO Figure 5.38: Brake thermal efficiencies for stoichiometric to lean operation for the 'bath• tub' group of chambers at MBT, WOT and 20.0 rps. 34 std bathtub s 1 ng 1 e slot B---Q castel lated 32 >« . A-" o -e z 30 a -B .A' B-- — B- c o ..-•J..B- a a 28 a •J < S 26 a s H 24 a 22 n 20 I I T" 1.05 1.1 1.15 1.2 1.25 1.3 1.35 1.4 RELATIVE AIR FUEL RATIO Figure 5.39: Brake thermal efficiencies for stoichiometric to lean operation for the 'bath• tub' group of chambers at MBT, WOT and 33.3 rps. Figures 158 34 . std bathtub single slot B---e castel lated 32 ...A* o z a 30 o E 28 . a < S 26 a: « B H a 24 < oS 03 22 20 1 I I I 1 1 1 1 1.05 1.1 1.15 1.2 1.25 1.3 1.35 1.4 RELATIVE AIR FUEL RATIO Figure 5.40: Brake thermal efficiencies for stoichiometric to lean operation for the 'bath• tub' group of chambers at MBT, WOT and 50.0 rps. 34 . std disc A---O bowl /piston B---B squish Jet 32 o>> z a 30 i—oi E a 28 a < S 26 cS __.„.-G B B o a . — •-Br' x H 24 . a 20 1 I I I I 1.4 1.05 1.1 1.15 1.2 1.25 1.3 1.35 RELATIVE AIR FUEL RATIO Figure 5.41: Brake thermal efficiencies for stoichiometric to lean operation for the 'disc' group of chambers at MBT, WOT and 20.0 rps. Figures 159 34 32 o z 30 . a a 28 B B -B <: B- S 26 as a s 24 . a 0—0 std disc bowl /piston CD 22 squish Jet 20 1 1 1 r— —I r~ 1 1.05 1.1 1.15 1.2 1.25 1.3 1.35 1.4 RELATIVE AIR FUEL RATIO Figure 5.42: Brake thermal efficiencies for stoichiometric to lean operation for the 'disc' group of chambers at MBT, WOT and 33.3 rps. 34 32 . .---***** >• v z A' — a 30 a a 28 a < 2 26 . oS a x H 24 a o—e std disc bowl /piston as 22 . B---B squish Jet CD 20 1 i i 1 1 1 1 l 1.05 1.1 1.15 1.2 1.25 1.3 1.35 1 RELATIVE AIR FUEL RATIO Figure 5.43: Brake thermal efficiencies for stoichiometric to lean operation for the 'disc' group of chambers at MBT, WOT and 50.0 rps. Figures 160 45 std bathtub single slot 40 castel lated o o H CD 35 cn W a o; a 30 . a a o z 25 D < Z go 20 . t—t H Z 15 o 1 1 1 1.05 1.1 1.15 1.2 1.25 1.3 1.35 1.4 RELATIVE AIR FUEL RATIO Figure 5.44: Ignition advance for stoichiometric to lean operation for the 'bathtub' group of chambers at MBT, WOT and 20.0 rps. 45 o—0 std bathtub single slot 40 . castel lated O Q t-i CO. 35 . CO .-0 a -B-" B : ; rGr& ' A--- - - A a 30 oA -B--"-"o'-•"- A * a Q a o z 25 $ Q < Z 20 o H .2 15 1.05 1.1 1.15 1.2 1.25 1.3 1.35 1.4 RELATIVE AIR FUEL RATIO Figure 5.45: Ignition advance for stoichiometric to lean operation for the 'bathtub' group of chambers at MBT, WOT and 33.3 rps. Figures 161 45 o—° std bathtub a---*, sing 1 e slot 40 . castel lated o Q H 35 CQ co cc o a 30 . Q o z 25 $ D Figure 5.46: Ignition advance for stoichiometric to lean operation for the 'bathtub' group of chambers at MBT, WOT and 50.0 rps. 40 Q—° std disc A---A bowl /piston 35 Q---0 squish Jet o a H CQ 30 acn a OS o 25 a D / a / B--~~" o / _--B-"""' z JB .A ---A-... 20 ""•A Q < Z >• o2 15 .-A' r—( A'" Z 10 1 1 1 1 1 1 1 1.05 1.1 1.15 1.2 1.25 1.3 1.35 1.4 RELATIVE AIR FUEL RATIO Figure 5.47: Ignition advance for stoichiometric to lean operation for the 'disc' group of chambers at MBT, WOT and 20.0 rps. Figures 162 40 35 30 _.-e er' B-- 25 20 . a—0 std disc 15 bowl /piston Q-'-B squish jet 10 —i 1 n 1 1 1 1.3 1.35 1.4 1.05 1.1 1.15 1.2 1.25 RELATIVE AIR FUEL RATIO Figure 5.48: Ignition advance for stoichiometric to lean operation for the 'disc' group of chambers at MBT, WOT and 33.3 rps. 40 35 30 25 Br' 20 tT 0—° std disc 15 bowl/piston a s--"61 squish jet 10 1 1 1 1 1 1 1 1 1.05 1.1 1.15 1.2 1.25 1.3 1.35 1.4 RELATIVE AIR FUEL RATIO Figure 5.49: Ignition advance for stoichiometric to lean operation for the 'disc' group of chambers at MBT, WOT and 50.0 rps. Figures 163 5000 fired pressure motored pressure 4000 3000 2000 1000 T -540 -450 -360 -270 -180 -90 180 CRANK ANGLE DEGREES FROM TDC Figure 5.50: Ensembled fired and motored pressure profiles over four strokes for the single slot chamber at WOT and 33.3 rps. Fired trace for MBT and RAFR=1.00. Figures 164 5000 std bathtub single slot castellated 4000 « CL. a 3000 tf a cu Q 2000 a j CQ S a co Z 1000 a 1 1 1 1 1 I I I -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.51: Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT, and 20.0 rps for RAFR=1.00. 5000 std bathtub single slot castellated 4000 . o a. Jt a tf 3000 D cn cn a tf e. Q a 2000 •a CQ S CaO Z a 1000 . ~~1 1 -180 -150 -120 -90 -60 60 90 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.52: Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT, and 20.0 rps for RAFR=1.27. Fig uies 165 5000 -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.53: Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT, and 33.3 rps for RAFR=1.00. 5000 std bathtub single slot castellated 4000 . a ai 3000 . D a ai OH D 2000 a pa S a in Z 1000 a 1 1— -90 -60 -30 0 30 60 90 180 180 -150 -120 CRANK ANGLE DEGREES FROM TDC Figure 5.54: Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT and 33.3 rps for RAFR=1.27. Figures 166 5000 -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.55: Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT, and 50.0 rps for RAFR=1.00. 5000 -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.56: Fired pressure profiles for the 'bathtub' group of chambers at MBT, WOT, and 50.0 rps for RAFR=1.27. Figures 5000 std disc bowl/piston -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.57: Fired pressure profiles for the 'disc' group of chambers at MBT, WOT, an 20.0 rps for RAFR=1.00. 5000 . std disc bowl/piston -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.58: Fired pressure profiles for the 'disc' group of chambers at MBT, WOT, an 20.0 rps for RAFR=1.27. Figures 168 5000 std disc bowl/piston 1 I I I I I I 1 1 1 1 1 -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.59: Fired pressure profiles for the 'disc' group of chambers at MBT, WOT, and 33.3 rps for RAFR=1.00. 5000 : , 4000 a 0. M a 3000 0. D cn cn § O. O 2000 . H J CQ s a 1000 cn Z a "i r -180 -150 -120 -90 -60 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.60: Fired pressure profiles for the 'disc' group of chambers at MBT, WOT, and 33.3 rps for RAFR=1.27. Figures 169 5000 bowl/piston squish Jet '4000 . a On M a 3000 . D co CO a cc a, 2000 Q H a S 63 co 1000 . 2 H I I -180 -150 -120 90 -60 -30 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.61: Fired pressure profiles for the 'disc' group of chambers at MBT, WOT, and 50.0 rps for RAFR=1.00. 5000 bowl/piston squish Jet 4000 . a OH M a ai 3000 . D co co a ai a, Q a 2000 oa S a CO 2 1000 . a -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.62: Fired pressure profiles for the 'disc' group of chambers at MBT, WOT, and 50.0 rps for RAFR=1.27. Fig ures 17.0 2500 single slot castellated 2000 « Cu Jt a 1500 C6 cn cn cc ft. 1000 . Q B CQ S 500 B cn Z a -180 -150 -120 -90 -60 -30 0 30 60 90 180 CRANK ANGLE DEGREES FROM TDC Figure 5.63: Fired pressure profiles for the 'bathtub' group of chambers at MBT, Bmep=2.5, and 33.3 rps for RAFR=1.00. 2500 single slot castellated 2000 a.a Jt a c£ 1500 cn acn ns OH 1000 Q a CQ S a 500 CO Za —1 1 1 1 1- -1 1 1 1 -180 -150 -120 -90 -60 -30 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.64: Fired pressure profiles for the 'bathtub' group of chambers at MBT, Bmep=2.5, and 33.3 rps for RAFR=1.27. Figures 171 2500 std disc bowl/piston squish jet 2000 0 0. « 1500 . OH P to a OH 1000 OH Q w CQ H 500 cn Z w I I 1 -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.65: Fired pressure profiles for the 'disc' group of chambers at MBT, Bmep=2.5, and 33.3 rps for RAFR=1.00. 2500 std disc bowl/piston squish Jet 2000 a Cu J* a ai 1500 . D co CO 63 ai OH Q a 1000 •J oa 63 CO Z 500 a CRANK ANGLE DEGREES FROM TDC Figure 5.66: Fired pressure profiles for the 'disc' group of chambers at MBT, Bmep=2.5, and 33.3 rps for RAFR=1.27. Figures 172 3000 s.lngle slot castellated 2500 bowl/piston •* 2000 & Jo cn cn § 1500 Q a CD 1000 500 . 1 1 1 1 1 1 1 r 120 150 180 -180 -150 -120 -90 -60 -30 0 30 60 90 CRANK ANGLE DEGREES FROM TDC Figure 5.67: Fired pressure profiles for three different chamber geometries MBT, Bmep=3.5, and 50.0 rps for RAFR=1.00. 3000 single slot castellated 2500 . bowl/piston U\ 2000 . \ 1500 \ fi \ Ii \ 1000 Ii \ \-\.- Vy, 500 r-^ -180 -150 -120 -90 -60 -30 0 30 60 90 120 150 180 CRANK ANGLE DEGREES FROM TDC Figure 5.68: Fired pressure profiles for three different chamber geometries at MBT, Bmep=3.5, and 50.0 rps for RAFR=1.27. Figures 173 I I I i I I I I l -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.69: Mass fraction burned curve for the bathtub chamber at MBT, WOT and 33.3 rps for RAFR=1.27 r I l I I 1 1 1 "6° -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.70: Mass fraction burned curves for different chamber geometries at MBT, WOT and 33.3 rps for RAFR=1.00 Figure 5.71: Mass fraction burned curves for different chamber geometries at MBT, WOT and 33.3 rps for RAFR=1.27 Figures 175 Figure 5.72: Mass fraction burned curves for the bathtub chamber at MBT, WOT and RAFR=1.27 for five speeds; 20.0, 33.3, 40.0, 50.0 and 66.7 rps. CRANK ANGLE DEGREES FROM TDC Figure 5.73: Mass fraction burned curves for the 'bathtub' group of chambers at MBT WOT, and 20.0 rps for RAFR=1.00. -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.74: Mass fraction burned curves for the 'bathtub' group of chambers at MBT WOT, and 20.0 rps for RAFR=1.27. I I I 1 1 1 1 1 1 -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.75: Mass fraction burned curves for the 'bathtub' group of chambers at MBT, WOT, and 33.3 rps for RAFR=1.00. I I I I i i i -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.76: Mass fraction burned curves for the 'bathtub' group of chambers at MBT, WOT, and 33.3 rps for RAFR=1.27. Figures 178 Figure 5.77: Mass fraction burned curves for the 'bathtub' group of chambers at MBT, WOT, and 50.0 rps for RAFR=1.00. I I i I i i i 1 -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.78: Mass fraction burned curves for the 'bathtub' group of chambers at MBT, WOT, and 50.0 rps for RAFR=1.27. Figures 179 -60 -45 -30 -15 0 15 30 45 CRANK ANGLE DEGREES FROM TDC Figure 5.79: Mass fraction burned curves for the 'disc' group of chambers at MBT, WOT, and 20.0 rps for RAFR=1.00. -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.80: Mass fraction burned curves for the 'disc' group of chambers at MBT, WOT, and 20.0 rps for RAFR=1.27. Figures 180 -60 -45 -30 -15 0 15 30 45 CRANK ANGLE DEGREES FROM TDC Figure 5.81: Mass fraction burned curves for the 'disc' group of chambers at MBT, WOT, and 33.3 rps for RAFR=1.00. -60 -45 -30 -15 0 15 . 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.82: Mass fraction burned curves for the 'disc' group of chambers at MBT, WOT and 33.3 rps for RAFR=1.27. Figures 181 bowl/piston squish Jet 0.8 Q H Z oi D m 0.6 . z o H O •< 0.4 . CC Eh in cn <: 2 0.2 1 1 1 1 1 1 1 -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.83: Mass fraction burned curves for the 'disc' group of chambers at MBT, WOT, and 50.0 rps for RAFR=1.00. bowl/piston squish jet 0.8 . Q a z OH D m 0.6 . z o H T" T -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.84: Mass fraction burned curves for the 'disc' group of chambers at MBT, WOT, and 50.0 rps for RAFR=1.27. Figures 182 ! I 1 —i 1 1 1 1 1 1 -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.85: Mass fraction burned curves for five different chamber geometries at MBT, Bmep=2.5, and 33.3 rps for RAFR=1.00. Figure 5.86: Mass fraction burned curves for five different chamber geometries at MBT, Bmep=2.5, and 33.3 rps for RAFR=1.27. Figures 183 60 CRANK ANGLE DEGREES FROM TDC Figure 5.87: Mass fraction burned curves for three different chamber geometries MBT, Bmep=3.5, and 50.0 rps for RAFR=1.00. -60 -45 -30 -15 0 15 30 45 60 CRANK ANGLE DEGREES FROM TDC Figure 5.88: Mass fraction burned curves for three different chamber geometries at MBT, Bmep=3.5, and 50.0 rps for RAFR=1.27. Figures 184 1.2 Initial burn:0 - 5% main burn :5 - 90% O i—i 0.6 a o « z rt 0.6 . p « z yo o.4 o < ai fa 0.2 . cn cn •< s n i i i i r bathtub single castle bowl sq.Jet disc Figure 5.89: Mass fraction burned ratio bar graphs, relative to the bathtub chamber, for MBT, WOT, and 20.0 rps for RAFR=1.00. 1.2 Initial burn:0 - 5% 1 main burn :5 - 90% 5 0.8 . CC Q a 0.6 z ai P zco o r>H 0.4 Figure 5.90: Mass fraction burned ratio bar graphs, relative to the bathtub chamber, for MBT, WOT and 20.0 rps for RAFR=1.27. 5 Oct (2 c i I-l -l CD n cn MASS FRACTION BURNED RATIO cn MASS FRACTION BURNED RATIO CD CD O o to o o H o o o H co lO CO p p a a CO M» co CO ^ CO P • p o co n CO •1 I-I o" o t3 a cn a cr O cr i-i n I-I a a a CD > CD ex. a. Tl n p cr to P n 0Q 00 i-« •i P P T3 a- a* co cn H CD i-l CD CD CD a- a- CD CD cr cr p p a a cr cr a<•>* p cr cr CD OO O ii CTT Figures 186 Figure 5.93: Mass fraction burned ratio bar graphs, relative to the bathtub chamber, MBT, WOT, and 50.0 rps for RAFR=1.00. 1.2 Initial burn:0 - 5% 1 main burn :5 - 90X § :0-8 ai a z. 0.6 ai D m z O 0.4 H O < CH CL. 0.2 < S 1 1 bowl sq.Jet disc i i r bathtub single castle Figure 5.94: Mass fraction burned ratio bar graphs, relative to the bathtub chamber, for MBT, WOT, and 50.0 rps for RAFR=1.27. Figures 187 1.2 1 . o.a tf5 Q a z 0.6 tf D CQ z o Ha o.4 o <: cn 0.2 cn «s! s bathtub single castle bowl sq.Jet disc Figure 5.95: Mass fraction burned ratio bar graphs, relative to the disc chamber, for MBT, Bmep=2.5, and 33.3 rps for RAFR=1.00. 1.2 Initial burn:0 - 5% main burn :5 - 90J> 1 . P 0.8 { |--- i i i r bathtub single castle bowl sq.Jet disc Figure 5.96: Mass fraction burned ratio bar graphs, relative to the disc chamber, for MBT, Bmep=2.5, and 33.3 rps for RAFR=1.27. Figures 188 1.2 initial burn-.O - 5% 1 main burn :5 - 90* H 0.8 • Figure 5.97: Mass fraction burned ratio bar graphs, relative to the castellated chamber, MBT, Bmep=3.5, and 50.0 rps for RAFR=1.00. 1.2 Initial burn:0 - 5% mmai n burn :5 - 90* H 0.8 cd D a z 0.6 cd CQ Z 2 0.4 j u <: cd a co 0.2 co •< 2 i 1 r 1 1 r bathtub single castle bowl sq.Jet disc Figure 5.98: Mass fraction burned ratio bar graphs, relative to the castellated chamber, for MBT, Bmep=3.5, and 50.0 rps for RAFR=1.27. Figures 189 900 875 «3 a OS 850 D co co a 825 OS o. a > 800 H aO a 775 fa a z < a 750 . IS IMEP per cycle Q mean a 725 mean+std dev mean-std dev p—i o 700 z "I I I I 1 1 140 160 180 200 20 40 60 80 100 120 CYCLE NUMBER Figure 5.99: Indicated mean effective pressure per cycle for the single slot chamber at MBT, WOT, and 33.3 rps for RAFR=1.27: 200 cycles Figure 5.100: Indicated mean effective pressure per cycle for the bathtub chamber at MBT, WOT, and 33.3 rps for RAFR=1.27: 44 cycles Appendix A Instrument Specification and Calibration This appendix contains the manufacturers calibration curves for: • Kistler piezo-electric pressure transducers; 6121 • Meriam Laminar flow element 50MW20-1.5 (Fuel flow rate) • Meriam Laminar flow element 50MC2-4F (Air flow rate) Specifications for the Kistler pressure transducer and sample calibration data are also presented. Pressure transducer Quasi static calibration of the pressure transducer and amplifier was car• ried out in accord with Ricardo consultants recommendations [70]. Calibra• tion was used to confirm the manufacturers constants and to confirm faulty operation suspected from the operating pressure trace. The calibration curves for the three pressure transducers used during this investigation are given in Figures A.l to A.3. The high input resistance of the Kistler 5004 amplifier unit, set on 'long' response makes it suitable for a quasi static calibration procedure using a dead weight tester. The amplifier unit was set to the sensitivity and range used during the fired operation, viz., sensitivity « 14 pC/bar and gain 10. After resetting the amplifier a weight was suddenly applied and the response 190 endix A. Instrument Specification and Calibration MODEL 6121 RANGE UNITS Range 0 ... 250 bar Calibrated partial ranges 0 ... 25 bar 0 ... 2.5 bar Overload 350 bar Sensitivity w -14 pC/bar Natural frequency > 55 kHz Frequency response ±1% 6 kHz Linearity, all ranges < ± 1.0 % FSO Hysteresis < 1.0 % FSO Acceleration sensitivity: axial < 0.003 bar/g transverse < 0.0002 bar/g Shock resistance 2000 g Thermal sensitivity shift: 20 ... 350 °C < ± 3 % 200 ± 50 °C « ± 1 % Calibrated in range 20 ... 350 °c Operating temperature range -196 ... 350 "C Transient temperature error < 0.02 bar Table A.l: Pressure transducer specifications for Kistler model 6121. Appendix A. Instrument Specification and Calibration 192 Pressure (psi) Amp output (mvolt) Pressure (bar) Charge output (pC) 15 128 1.03 17.9 35 253 2.41 35.4 55 415 3.79 58.1 75 533 5.15 74.6 105 775 7.24 108.5 125 910 8.58 127.4 155 1140 10.69 159.6 205 1450 14.14 203.0 Table A.2: Pressure transducer calibration data for Kistler model 6121. No. 317205 voltage measured on a digital oscilloscope. Calibration data for the pressure transducer No. 317205 is given in Table A.l and indicated on Figure A.l. The charge produced by the transducer is calculated from: Q = V * Cg* k where V is the amplifier output in volts, Cg is the value of the range capacitor in pF and A; is the sensitivity setting of the amplifier expressed as a part of unity, eg.; sensitivity of 1.4 mechanical units per volt is equivalent to 0.14. Laminar flow elements The laminar flow elements calibration curves are given in Figures A.4 and A.5. Corrections for standard air pressure and temperature (21.1 °C,101.3 kPa) were calculated using the manufacturers supplied tabulated values. These values were approximated to linear equations for use in the DATAQ and CRUNCH programs. The viscosity of natural gas, used in these equations was determined from the approximate gas composition given in Appendix C. Appendix A. Instrument Specification and Calibration 193 Druckaufnehmer Capteur de pression Pressure transducer Type 6121 SN 317205 Kalibrierter Bereicti Betriebstemperaturbereich Gamme elalonnee iDar] Gamme de temp, d'utiitsalion f*C) -60.-350 Calibrated range 0...25O 0-.25 0-S.5 Operating temperature range Emplindlicfttieil Kalibriert bet Sensibilite [pC/bar] -n Etalonne a 20 *C Sensitivity -1t,0 .o -14,0 Calibrated at by Sh Date 3.12.8? Unearitat _ . 1 bar s 10' N •_, ^1"' = 1,019...at = 14,50...psi Linearite <±%FSO 1al = 1kpcm B 1kgfcrrr' = 0,960665 bar 0,3 o,3 0,3 1psi=0,06894...bar Figure A.l: Kistler and laboratory calibration curves for pressure transducer Model 6121 No. 317205, (used for single slot castellated, bowl-in-piston and fired squish jet tests). Appendix A. Instrument Specification and Calibration 194 Daickaufnehmer Capteur de pression Pressure transducer Type 6121 SN 282737 Kafibrierter Bereich Betriebsiemperaturbereich Gamine etalonnee [bar] Gamme de temp, d'utilisation PC] -80-350 Calibrated range 0-250 0_.25 0-2.5 Operating temperature range EmpfincJIicfikeit Kalibriert b*i Sensible [pC/Oar] -H,7 -H,6 Etafonne a 20 *C Sensriiviry Calibrated a) by Sh .Date 15.12.86 Lineantat 1 bar = 10* N • rrr* = 1,019...al = 14.50...psi 1 Linearite ' Figure A.2: Kistler calibration curve for pressure transducer Model 6121 No. 282737, (used for standard bathtub and disc chamber tests). Appendix A. Instrument Specification and Calibration 195 Druckaufnehmer Capteur de pression Pressure transducer Type 6121 SN 317125 Katibrierter Bereicfi Betriebatemperaturtwrettfi Gamme etalonnee [bar] Gamme de temp, d'utilisatlon [*C] -S0...350 Calibrated range 0...2S0 0...25 0...2.5 Operating temperature range Emptindltcnkett Kalibriert bei Sensibilite [pC/bar] -14.8 -14.5 -14.5 Etalonne d 20 9C Sensitivity Calibrated at by Sh Date 11.8.88 Linearital _ „ 1 bar * 10' N • itr' = 1.019...at = u,50...psi Linearite <+%FSO 0.3 0.3 0.3 1at= 1kpcm-'= 1kgtcrrr' = 0,980665 bar 1psi-0.06894...bar 0 25 50 75 100 125 150 175 200 225 250 0 2.S 5 7.5 10 12,5 15 17.5 20 22.5 25 0 0.2S 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 Abhftngigkett der Empfindlfc*ikect [ . , [ . :, . I ] , ; , | j j • — ' I i . j . —4- 3i, von der Temperatur —f 1 '• : .l~^±^s|= ' , "[ ^—.^1 • """"] Sensibilite en fonction r*"""l^ • • | ; [ \ '• | • --- I* ' -1^ oC/bar de ta temperature J "*"~ ~t' , I I ! I ' , I J_ • ... j —-—j- —- j . . . • Sensitivity versus temperature | • • ..... j . •••_—] I ...... I . .1 1 Figure A.3: Kistler calibration curve for pressure transducer Model 6121 No. 317125, (used for the motored squish jet test only). Appendix A. Instrument Specification and Calibration 195 Druckaufnehmer Capteur de pression Pressure transducer Type 6121 SN 317125 Kalibrierter Bereicti Setriebstemperaturbereich Gamme etalonnee [bar] Gamme de temp, d'utilisation f*C) -80...350 Calibrated range 0...250 0...25 0...2.5 Operating temperature range Empfindlidikeit Kaltbrtert bei a Sensibilile [pC/bar] -H.B -1"i.5 -Ti.5 Etalonne a 20 C Sensitivity Calibrated at by Sh Date 11.8.88 Linearitat 1bar = 10' N mr' = 1.019...at = 14.50...psi , Linearite <+T»rSO 0.3 0.3 0.3 1at=1kpcm-'=1kgtcm- = 0.980665 bar 1psi=0.06894...bar 0 25 50 75 100 125 150 175 200 225 250 0 2.5 5 7.5 10 12,5 15 17.5 20 22.5 25 0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 Abhangigkert der Empfindlichkeit U.. . "j ; i.•..••TT . 1 . | ; ; • ' | ' • I ' I ' ; j ' . ' ' ] 3'-* ; von der Temperatur I ; : t ; : •! ,,' • ' T *—-— j; ; ' . ' -.^~J Senaibilrte en tonction L^rrrr.jr. I 1.11 : ; ' 1 ' .. ITT. , 7^|TTZL^T4 -^.8 pC/bar de la temperature ~ + — - 4 -!• • •• 4- — | •—-|~~' •-•{• — •• —f- -3'.. Sensitivity versus temperature 350 CC Figure A.3: Kistler calibration curve for pressure transducer Model 6121 No. 317125, (used for the motored squish jet test only). Appendix A. Instrument Specification and Calibration 196 Figure A.4: Meriam calibration curve for Model 50MW20-1.5 No. S-4875-l,( used for Natural gas flow rate). Appendix A. Instrument Specification and Calibration 197 Figure A.5: Meriam calibration curve for Model 50MC2-4F No. S-4875-2,( used for Air flow rate). Appendix B Hot Wire Anemometry Specification and Calibration This appendix presents the specifications for the hotwire probe and bridge unit and typical calibration data. Calibration of the hotwire system was carried out in a small wind tunnel for velocities between 0.5 and 16 m/s in accord with DISA recommendations. The internal probe and cable resistance were compensated for, through the rezeroing of the bridge unit for the short-circuited cable, and by subtracting the internal probe resistance from the measured resistance value. Typical values of sensor resistance after annealing for 6-8 hours at the operating tem• perature were 10.8 fi ambient and 12.1 fi operating. The resistance drop during the annealing process, before a stable value was reached, was approx• imately 1 fi. Balance of the bridge was obtained by applying a square wave to the probe exposed to the maximum velocity in the wind tunnel and adjusting the cable compensation for a clear response signal. A pitot tube and inclined alcohol manometer provided flow calibration data against a mean voltage unit. These data were then used in the calibration program HW-Cal to obtain the analytical model constants. The exponential constant n varied from 0.36 to 0.66 over the wires used , showing a good agreement with the initial theory of King [78]. Sample data for wire No. 1 used for the bowl-in-piston chamber flow field 198 Appendix B. Hot Wire Anemometry Specification and Calibration 199 PROBE: Model TSI 1226 No. 65202 Internal Resistance o.57 fi Wire material TSI Platinum Iridium, PI2.5 Diameter 6.3 pm Length 1.5 mm Thermal coeff. of resistance .00089/°C Operating temperature 600 °C BRIDGE: Model DISA Type 55M10 CTA Standard FILTER: Model DISA Type 55D25 Auxiliary unit Low pass filter 20 kHz High pass filter off Table B.l: Hot wire anemometry equipment and specifications. measurements is given in Table B.2 and illustrated in Figure B.l. The am• bient and operating resistance of this wire were 11.34 fi and was 16.96 fi respectively. The calibration constants were obtained from a curve fitting routine. Appendix B. Hot Wire Anemometry Specification and Calibration 200 A Pressure (Pa) Voltage (V) Reynolds No. Nusselt No. 1.4 3.44 0.63 .571 4.2 3.69 1.09 .664 8.2 3.86 1.52 .732 14.4 4.00 2.02 .789 20.4 4.10 2.40 .832 31.4 4.22 2.98 .884 44.6 4.33 3.55 .932 59.5 4.41 4.10 .971 76.0 4.49 4.64 1.01 95.5 4.57 5.20 1.05 112 4.62 5.63 1.07 145 4.71 6.41 1.12 172 4.77 6.98 1.15 189 4.81 7.32 1.17 Table B.2: Hotwire calibration data for wire No. 1, yielding the calibration constants: 4=0.1575; B=0.4908; and n=0.360. Appendix B. Hot Wire Anemometry Specification and Calibration 201 HOT WIRE CALIBRATION : #1 24.6.88 A-.157516, B».490792, N..360255 1.4 REYNOLDS NUMBER Figure B.l: Hotwire calibration curve for Wire No. 1 Ramb = H-34^/2^ = 11.9511, (used for the bowl-in-piston chamber tests). Appendix C BC Natural Gas Properties This appendix presents calculations of the properties of BC Natural gas, used in the firing tests. The molecular weight, heating values, viscosity and sto• ichiometric air fuel ratio are calculated from the typical composition given in Table C.l. The figures given are those used in the original calculation of constants for the DATAQ and CRUNCH programs and used in previous AFL Group studies. This gas composition was compared to monthly averages from BC Hydro over the test period and found to be within 1 %. This led to a 1 % variation in stoichiometric air fuel value. COMPOSITION VOLUME % Methane 94.00 Ethane 3.30 Propane 1.00 Iso-Butane 0.15 N-Butane 0.20 Iso-Pentane 0.02 N-Pentane 0.02 Hexane 0.01 Nitrogen 1.00 Carbon Dioxide 0.01 Water content: 3-4 lbs/mft3 Table C.l: Composition of BC Natural Gas. The molecular weight and heating values, shown in Tables C.2 and C.3, 202 Appendix C. BC Natural Gas Properties 203 COMPONENT Vol Fraction Mol. Wt. Mass (kg/kmol) Methane CH4 0.940 16.040 15.078 Ethane C2H$ 0.033 30.070 0.992 Propane C$H& 0.010 44.097 0.441 Butane C4H10 0.004 58.124 0.232 Nitrogen N2 0.010 28.013 0.280 Carbon Dioxide 0.003 44.010 0.132 co2 TOTALS 1.000 17.156 Table C.2: Molecular weight of BC Natural gas. COMPONENT Mass (%) HHV (kJ/kg) LHV (kJ/kg) Methane CHA 0.879 * 55496 = 48781 * 50010 = 43959 Ethane C2H$ 0.058 * 51875 = 3008 * 47484 = 2754 Propane CzHg 0.026 * 50343 = 1309 * 46353 = 1205 Butane C^Hxo 0.013 * 49500 = 644 * 45714 = 640 Nitrogen N2 0.016 0 0 0 0 Carbon Dioxide C02 0.008 0 0 0 0 TOTALS 1.000 53742 48558 Table C.3: Higher and Lower Heating values of BC Natural gas. were determined using a convenient approximation, that is, hydrocarbons of higher order tha C^H^o were included with the butane figures. From the average molecular weight of 17.156 the gas constant and density are determined by: R 8.3143 R = 0.4846. kJ/kgK MW 17.156 P = z*R*T at 21.1°C, 101.3 kPa, z « 1 this gives: p = 209/7/(°K) kg/m3. The viscosity of the natural gas at 0°C may be obtained from a similar Appendix C. BC Natural Gas Properties 204 COMPONENT Vol.Frac. Mol.Wt. Viscosity ViMW? Vi MWi Mi Methane CHA 0.940 16.040 102.6 386.29 3.765 Ethane C2H6 0.033 30.070 84.8 15.35 0.181 Propane C3H.S 0.014 48.105 75.0 7.28 0.020 Carbon Dioxide C02 0.003 44.01 139.0 2.78 0.020 Nitrogen N2 0.010 28.013 166.0 8.80 0.053 TOTALS 1.000 420.50 4.116 Table C.4: Viscosity calculations for BC Natural gas. simplified composition using: 1 _ E Vt * Pi* MWj7 pgas — 1 where t/j is the volume fraction, MW{ is the molecular weight and pi is the viscosity, of component i. In this approximation butane and higher com• ponents are included with propane. From the above the viscosity at 0°C is determined by: 420.5 Vga, = = 102.16/xpoise at 0°C 4.11b The viscosity at other temperatures is obtained from: Tf pga. = 9.879 * (T + 163.17) The viscosity of Natural gas at 21.1 °C is 108.96 /zpoise. The stoichiometric air fuel ratio may be determined from the equation for the complete combustion of one mole of BC Natural gas approximated by: .94(7 H4 + .033C2#6 + MC3H6 + .004C4Hlo+ 1.055CO2 + 2.039ff2O + .OIJV2 + .003CO2 + 2.072O2+ (7.791 + .01)AT2 2.072(3.76)JV2 Appendix C. BC Natural Gas Properties 205 The stoichiometric air fuel ratio is then determined by: AFR = 2.072(1 + 3.76). 28.97 .94(16) + .033(30) + .01(44) + .004(58.1) + 0.01(28) + .003(44) Appendix D Pressure Filtering Methods This appendix presents examples of the filtering method used on the motored and fired pressure profiles. Filtering of the motored pressure signal was performed in the region of intake valve closing to remove the transient effect caused by the valve motion, detrimental to the temperature calculations. Initially low pass filtering using a 'sixth order butterworth' method was considered. To effectively remove the pressure spikes at low speeds a cut off frequency of 6 Khz was required. The pressure data at this speed was digitalized at 36 kHz. This method however causes a phase shift over the whole range or a step in the data if applied to a specific data. Figure D.l illustrates the effect of filtering at 6 kHz on motored pressure obtained at 50.0 rps. Figure D.2 shows the best approximation obtained using a replacement method varying amounts of end tension. This was also deemed unsatisfactory. The preferred method used in the preceding study involved an averaging and smoothing routine over the 600 points (180 to 60 degrees BTDC). Fig• ure D.3 compares the effect of averaging over a 30 point (6 degree) and 60 point (12 degree) windows. The 6 degree window was used over all speeds for the motored pressure. Similarly an averaging and smoothing method was applied to the fired pressure in the region 150 to 55 degrees BTDC, using a 12 degree window. 206 Appendix D. Pressure Filtering Methods 207 Figure D.4 shows an expanded view of the pressure profile for the single slot piston at 33.3 rps before and after smoothing. Appendix D. Pressure Filtering Methods 208 2000 NO FILTERING FC 6KHZ (-150,-60) 1500 1000 500 . T ~l— "T "1— -180 -150 -120 -9-90 -60 •30 CRANK ANGLE DEGREES FROM TDC Figure D.l: Motored pressure trace at 50.0 rps low pass filtered at 6 kHz over the entire range and over an initial region. Appendix D. Pressure Filtering Methods 209 2000 J NO SMOOTHING P(END POINTS)*.03 P(END P0INTS)=.02 a CM -* a 1500 J a D CO aCO a 1000 J Q a j m aS / CO 500 / & w 1 1- -180 -150 -120 -90 -60 -30 CRANK ANGLE DEGREES FROM TDC Figure D.2: Motored pressure trace at 50.0 rps with the region between 150 and 60 degrees BTDC replaced with a section under end tension Appendix D. Pressure Filtering Methods 210 500 AVE 30 POINTS NO END TENSION AVE 60 POINTS NO ENO TENSION NO SMOOTHING OR FILTERING I 1 1 1 •150 -120 -90 . -60 -30 CRANK ANGLE DEGREES FROM TDC Figure D.3: Motored pressure trace at 50.0 rps with the region between 180 and 60 degrees BTDC averaged and smoothed over 6 and 12 degree windows. Appendix D. Pressure Filtering Methods 211 1 I i i i i 1 -180 -150 -120 -90 -60 -30 0 CRANK ANGLE DEGREES FROM TDC 400 50 . -180 -150 -120 -90 -60 -30 0 CRANK ANGLE DEGREES FROM TDC Figure D.4: Expanded fired pressure traces at 33.3 rps, with and without averaging and smoothing applied over 12 degree windows in the region between 150 and 55 degrees BTDC.