the use of hydrogen as a fuel for compression ignition engines

the use of hydrogen as a fuel for compression ignition engines
THE USE OF HYDROGEN AS A FUEL FOR
COMPRESSION IGNITION ENGINES
Thesis by
Jorge M.G. Antunes
In Partial Fulfilment of the Requirements
for the Degree of
Doctor of Philosophy
Date of Submission
September 2010
ii
Keywords: diesel engine, compression ignition, hydrogen, injection,
simulation, pulsed injection, emissions, HCCI, Miller cycle,
Abstract
The objective of this research was to investigate the applicability of
hydrogen as a fuel for compression ignition engines. The research indicates
that hydrogen is a suitable fuel for “compression ignition” (CI) engines,
“fumigated diesel” (FD), “homogeneous charge compression ignition”
(HCCI) and “direct injection of hydrogen” (DIH2).
Peculiarities of the various modes of operation with hydrogen were
investigated using a high speed commercial direct injection diesel engine,
Deutz 1FL 511 with a compression ratio of 17:1, as well as a simulation
model to assist with on the understanding of certain phenomena that were
impossible to reproduce due to the engine and transducers physical
limitations.
Instrumentation with high-speed data acquisition was designed and
installed to measure crankshaft speed and position, airflow rate, inlet air
pressure and temperature, fuel consumption, brake power, cylinder
combustion
pressure,
and
exhaust
gas
temperature.
The
design,
construction and characterization of a pulse controlled hydrogen injection
system for HCCI and DIH2 was carried out and discussed.
In this research, special attention was paid to characterize and identify the
operating parameters that control the hydrogen combustion in a CI engine.
High rates of engine cylinder pressure rise were found when using hydrogen
and some form of control solution is required. Simulation and engine tests
were carried out to characterize and identify new design approaches to
control such high rates of pressure rise, culminating in the proposal of a
pulsed injection methodology, and also the use of the Miller cycle to
mitigate the observed high rates of pressure rise. A number of possible
iii
innovative solutions and measures, making the hydrogen engine operation
reliable and safe are also presented.
iv
Acknowledgments
This research work was carried out at the TecnoVeritas - Services of
Engineering and Systems Technology Ltd. laboratory, without which it
would not be possible.
Profound thanks are due to my family for their unconditional support, and
love given throughout my student life, most in particular during the present
research work, for bearing my bad mood when research work looked like
being impossible, and the discouragement was beating hard.
During the development of this research at the School of Marine Science
and Technology, I received a valuable help, advice and financial support
from my friend and supervisor Professor A. P. Roskilly.
I wish to thank, Mr. João Reis, for the Saturdays and nights spent with me
around the engine and Dr Yao Dong Wang for his patience and support.
Thank you GOD! There is more to come! I will be counting on you!
To: Margarida, Joana, Anucha, Gabi, Mia e Jaime.
v
vi
List of papers
Antunes J. and Roskilly A., (2004). “The use of H2 on compression
ignition engines”,
3rd European Congress on economics and management of energy in
industry, Lisbon.
Antunes J. and Roskilly A., (2006). “Opportunities & advantages of the
use of hydrogen on board ships – new concepts”,
X International Naval Engineering Conference, Lisbon.
Antunes J. and Roskilly A. R.Mikalsen, (2008). “An investigation of
hydrogen-fuelled HCCI engine performance and operation”,
Published by the International Journal of Hydrogen Energy.
Antunes J. and Roskilly A. R.Mikalsen, (2008). “An experimental study of a
direct injection compression ignition hydrogen engine”,
Published by International Journal of Hydrogen Energy.
Antunes J. and Roskilly A. R.Mikalsen, (2009). “The hydrogen-fuelled
HCCI engine performance and operation”,
Paper presented at the Conference HYPOTHESIS VIII Hydrogen Power
Theoretical and Engineering Solutions - International Symposium April
2009.
Antunes J. and Roskilly A. R.Mikalsen (2009) “Conversion of large-bore
diesel engines for heavy fuel oil and natural gas dual fuel operation”.
Paper submitted in 2009, waiting for publication by CIMAC.
vii
viii
Contents
Abstract ............................................................................................................................. iii
Acknowledgements .................................................................................................. v
List of papers ............................................................................................................... vii
List of figures ............................................................................................................... xv
List of tables ............................................................................................................... xxv
Nomenclature............................................................................................................xxvii
1 Introduction.................................................................................................................1
1.1 Alternative power generating systems ................................................................................2
1.1.1 Stationary power generation ...............................................................................2
1.1.2 Propulsion systems .............................................................................................3
1.2 Internal combustion engines ...............................................................................................4
1.2.1 The use of hydrogen in CI engines .....................................................................4
1.3 Contribution to existing research ........................................................................................6
2 Hydrogen engine research review .............................................................9
2.1 Hydrogen utilisation as an engine fuel ................................................................................9
2.2 Spark ignition hydrogen engine operation ........................................................................11
2.3 Improvement of CI engines performance: bi-fuel experience ...........................................12
2.3.1 Early work on bi-fuel CI engine operation .........................................................13
2.3.2 Bi-fuel operation with hydrogen induction .........................................................14
2.4 Compression ignition hydrogen engines ...........................................................................17
2.4.1 The DIH2 engine ................................................................................................18
2.4.2 DIH2 engine specific problems ..........................................................................27
2.4.3 The HCCI hydrogen engine ..............................................................................27
2.4.4 HCCI hydrogen engine specific problems .........................................................29
2.5 Fundamental hydrogen engine specific properties ...........................................................30
ix
2.5.1 Comparison of hydrogen versus methane combustion.....................................31
2.5.2 Heat transfer in hydrogen fuelled engines ........................................................34
2.5.3 Exhaust heat losses in hydrogen fuelled engines .............................................37
2.6 Hydrogen engine safety ....................................................................................................39
2.7 Conclusions .......................................................................................................................40
3 Engine performance analysis through experimentation.........43
3.1 Engine experimental setup ...............................................................................................43
3.1.1 Compression ignition engine .............................................................................43
3.1.2 Air supply system modifications ........................................................................47
3.1.3 Fuel system .......................................................................................................47
3.1.4 Exhaust gas system ..........................................................................................48
3.1.5 Test rig instrumentation and data acquisition system .......................................49
3.1.6 Data acquisition system hardware ....................................................................50
3.1.7 Data acquisition system software......................................................................53
3.1.8 Engine speed and crank angle measurement ..................................................57
3.1.9 Cylinder pressure transducer ............................................................................58
3.1.10 Air mass flow transducer and measurement ...................................................58
3.1.11 Hydrogen mass flow measurement.................................................................59
3.1.12 Lambda oxygen transducer and measurement ..............................................59
3.2 Hydrogen fuel injection systems .......................................................................................59
3.2.1 Material considerations when using hydrogen ..................................................59
3.2.2 Hydrogen HCCI injection system ......................................................................61
3.2.3 Hydrogen direct injection system ......................................................................62
3.2.4 Injector hydraulic power pack............................................................................65
3.2.5 Alternative DIH2 injector design ........................................................................67
3.3 Hydrogen injector test rig ..................................................................................................70
3.3.1 Pressure vessel for injector testing ...................................................................70
3.3.2 Pulse width modulation control circuit ...............................................................73
3.3.3 Static and dynamic characterisation of HCCI and DIH2 injectors .....................76
3.3.3.1 Static performance test results for the HCCI injector ........................77
3.3.3.2 Static performance tests of the DIH2 injector ....................................80
3.3.3.3 – HCCI dynamic injector response ....................................................83
3.3.3.4 DIH2 dynamic injector response ........................................................83
3.4 Hydrogen injection engine control system ........................................................................85
3.4.1 Low level hydrogen injection control loop .........................................................86
3.4.2 High level hydrogen injection control loop ........................................................90
3.5 Conclusion ........................................................................................................................95
x
4 Experimental testing of hydrogen engine operation:
results and analysis .........................................................................................97
4.1 Objectives of engine testing and methodology .................................................................97
4.1.1 Testing procedures............................................................................................97
4.1.2 Engine operation and safety .............................................................................98
4.1.3 Instrumentation set up and operation................................................................98
4.2 Data logging and treatment ...............................................................................................99
4.2.1 Cylinder pressure measurement data ...............................................................99
4.2.2 Cylinder pressure sampling rate .....................................................................100
4.3 Methodology of engine testing ........................................................................................103
4.4 Diesel fuel operation characterisation .............................................................................104
4.5 Dual fuel operation ..........................................................................................................107
4.5.1 Experimental setup..........................................................................................108
4.5.2 Test results ......................................................................................................109
4.5.2.1 Combustion and energy efficiency ..................................................109
4.5.2.2 Exhaust gas emissions....................................................................111
4.6 HCCI operation characterization .....................................................................................114
4.6.1 Inlet air temperature and ignition control .........................................................114
4.6.2 Operating characteristics and performance ....................................................116
4.6.3 Emissions ........................................................................................................123
4.6.4 Operational stability .........................................................................................124
4.7 DIH2 operation characterization ......................................................................................125
4.7.1 Auto-ignition of the hydrogen jet .....................................................................126
4.7.2 Engine tests.....................................................................................................127
4.7.3 Emissions formation ........................................................................................131
4.8 Efficiency calculations and comparison ..........................................................................132
4.8.1 Comparison of thermal efficiencies .................................................................134
4.9 Uncertainty of measured variables .................................................................................136
4.9.1 Quantification of uncertainty............................................................................136
4.9.2 Uncertainty in thermal efficiency .....................................................................137
4.9.3 Measurement of thermal efficiency .................................................................138
4.9.4 Uncertainty in engine power calculation .........................................................139
4.9.5 Uncertainty in fuel mass flow rate ...................................................................140
4.9.6 Uncertainty in the volumetric flow measurement ............................................141
4.9.7 Uncertainty associated with other measurements ..........................................144
4.10 Conclusion ....................................................................................................................144
xi
5 Modelling and simulation .............................................................................149
5.1 Modelling of hydrogen HCCI and DIH2 engines..............................................................149
5.1.1 Modelling objectives ........................................................................................150
5.1.2 HCCI and DIH2 engine model differences .......................................................151
5.2 Simulation program description ......................................................................................152
5.2.1 High level program structure ...........................................................................152
5.2.2 Model parameters ...........................................................................................153
5.2.3 Engine parameters ..........................................................................................154
5.2.4 Ambient conditions ..........................................................................................154
5.3 Engine cycle calculation ..................................................................................................154
5.3.1 Piston crank mechanism .................................................................................156
5.3.2 Heat losses......................................................................................................159
5.3.3 Cylinder valves modelling ...............................................................................164
5.3.4 Ignition delay ...................................................................................................176
5.3.5 Mass f low calculation .....................................................................................177
5.4 Engine simulation program structure ..............................................................................181
5.4.1 Simulation program interfaces ........................................................................185
5.5 Hydrogen injectors modelling ..........................................................................................190
5.5.1 HCCI injector design considerations ...............................................................190
5.5.2 DIH2 injector design.........................................................................................193
5.5.3 Under-expanded gas flow in the proximity of a nozzle hole ...........................197
5.5.4 Injector hydraulic actuation modelling .............................................................201
5.5.4.1 Assumptions for the hydraulic injector simulation ...........................207
5.5.5 DIH2 injector dynamic simulation ....................................................................208
5.6 Summary .........................................................................................................................215
6 Performance analysis through simulation
.....................................217
6.1 Hydrogen HCCI model analyses.....................................................................................217
6.1.1 Validation and evaluation of the HCCI model .................................................218
6.1.2 Hydrogen HCCI engine operation analysis .....................................................220
6.1.3 Problems associated with HCCI operation .....................................................221
6.1.3.1 Hydrogen slip during the valve overlap period ................................221
6.1.3.2 High rates of pressure rise. .............................................................222
6.1.3.3 Power limitation of the HCCI engine ...............................................224
6.1.4 Possible design improvements using simulation ............................................225
6.1.4.1 Combustion control and dependence of the inlet air temperature ..225
6.2 Hydrogen direct injection engine model validation .........................................................228
6.2.1 DIH2 engine design and operational analysis .................................................230
6.2.2 Control of MRPR and engine optimisation ......................................................232
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6.2.3 Comparison and conclusions regarding the simulated continuous and
pulsed injection...............................................................................................252
6.3 Effect of valve timing (Miller cycle) on DIH2 engine performance. ..................................253
6.3.1 Effect of the Miller cycle on the DIH2 engine ...................................................254
6.4 DIH2 injector dynamic simulation ....................................................................................256
6.4.1 Effect of the inertia of the moving parts on the injector dynamic response ....257
6.4.2 Effect of duty cycle on the injector dynamic response ....................................261
6.4.3 Effect of the injector actuation frequency on the dynamic response ..............266
6.4.4 Effect of the hydraulic pressure on the injector dynamic response ................272
6.4.5 Effect of the static force (pre-load) on the injector dynamic response ............276
6.4.6 Summary of injector design analyses .............................................................279
6.4.7 Possible injector design improvements ...........................................................280
6.5 Summary ........................................................................................................................280
7 Conclusions and recommendations .......................................283
7.1 Summary of the results ...................................................................................................285
7.2 Hydrogen as a fuel for CI engines, further considerations..............................................288
7.3 Recommendations for further work .................................................................................297
7.3.1 Feasibility of the hydrogen fuelled CI engine ..................................................298
7.3.2 Engine mechanical loading and controllability ................................................298
7.3.3 Compression ratio adjustment ........................................................................300
7.3.4 Control of the inlet air temperature..................................................................300
7.3.5 Internal exhaust gas recirculation ...................................................................302
7.3.6 Pulsed injection for DIH2 ............................................................................................................................304
References ..............................................................................................................311
Appendix A Commercial dual fuel engine developments ........321
Appendix B Development of a dual fuel combined heat
and power research facility ...................................................................337
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xiv
List of Figures:
2.1: Engine brake thermal efficiency as a function of brake power for different
hydrogen flow rates. ............................................................................ 15
2.2: Exhaust gas smoke levels as a function of engine load for varying hydrogen
injection rates. .................................................................................. 16
2.3 Internal and external cylinder charge formation. ............................................ 18
2.4: Hydrogen engine and sub-systems experimental setup ..................................... 20
2.5: Engine performance under direct injection hydrogen operation. ......................... 21
2.6: Cycle-to-cycle variations in cylinder pressure under DIH2 operation. .................... 21
2.7: Indicated thermal efficiency as a function of brake mean effective pressure for
different injector nozzle designs. ............................................................ 22
2.8: Effect of ambient gas temperature on hydrogen ignition delay ........................... 23
2.9: Effect of the ambient gas density () on hydrogen ignition delay. ....................... 24
2.10: Effect of O2 concentration on ignition delay. ............................................... 24
2.11: Effect of fuel temperature on ignition delay. .............................................. 25
2.12: Effect of ambient gas O2 concentration on H2 combustion and pressure rise
with ambient temperature 1000 K. ........................................................... 26
2.13: Effect of ambient air temperature and O2 concentration on the rate of heat
release. ........................................................................................... 26
2.14: Influence of ignition timing on hydrogen and methane combustion. ................... 31
2.15: Influence of ignition timing on components of thermal efficiency. ..................... 32
2.16: Influence of excess air ratio on the components of thermal efficiency. ............... 33
2.17: Comparison of hydrogen and methane combustion. ....................................... 35
2.18: Coefficients in the new proposed heat transfer equations for different
operating conditions. ........................................................................... 36
3.1: Cross section of the test diesel engine. ....................................................... 44
3.2: Engine test rig. .................................................................................... 46
3.3: Inlet manifold heater and mass flow meter. ................................................. 47
3.4: Diesel oil flow meter. ............................................................................ 48
3.5: Hydrogen direct injection engine test rig schematic diagram. ............................ 49
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3.6: Hydrogen Homogenous Charge Compression Ignition engine test rig schematic
diagram. ......................................................................................... 50
3.7 Data acquisition block diagram. ................................................................. 52
3.8: Visualization Display menu. ..................................................................... 53
3.9: Analysis Display menu. ........................................................................... 55
3.10: Cylinder pressure against crank angle analysis display. ................................... 55
3.11: Configuration display menu.................................................................... 56
3.12: Diagnosis display menu. ........................................................................ 57
3.13: Absolute encoder fitted at the camshaft end of the test engine. ....................... 57
3.14: Air mass flow meter Bosch HFM5. ............................................................ 58
3.15: Fast-acting solenoid injection valve and hydrogen pressure gauge, fitted on
the engine. ....................................................................................... 61
3.16: HCCI Hydrogen injection solenoid activated ball valve. .................................. 62
3.17: Simplified diagram of the HCCI hydrogen injection solenoid activated ball
valve. ............................................................................................. 62
3.18: Solenoid controlled hydraulic DIH2 cross. ................................................... 63
3.19: Solenoid controlled hydraulic DIH2 injector. ............................................... 63
3.20: Solenoid controlled hydraulic DIH2 injector installation.................................. 64
3.21: Hydraulic power pack and DIH2 injector. .................................................... 66
3.22: DIH2 hydraulic actuating and inert gas (nitrogen) purging system. ................... 66
3.23: Hydraulically controlled and actuated hydrogen injector. ............................... 68
3.24: Schematic diagram of the hydraulically controlled and actuated injection
system. ........................................................................................... 69
3.25: Injector test vessel.............................................................................. 70
3.26: Injector testing rig. ............................................................................. 71
3.27: Photograph of the DIH2 injector under test. ................................................ 73
3.28: Pulse width modulation control circuit. ..................................................... 74
3.29: Basic pulse width modulation control circuit characteristic curve. ..................... 75
3.30: Variable pulse width modulation control circuit. .......................................... 75
3.31: Variable pulse width modulation control circuit bread board. .......................... 76
3.32: Flow rate (mg/injection) as a function of the PWM and supply pressure for
the HCCI injector. .............................................................................. 80
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3.33: Flow rate (mg/injection) as a function of the PWM and average. ....................... 82
3.34: Oscilloscope traces showing time delay measurements for the DIH2 injector.
Supply pressure signal (0-100%). .............................................................. 85
3.35: Low level hydrogen control loop circuit board. ............................................ 86
3.36: Low level hydrogen injector control system hardware. ................................... 86
3.37: Main microcontroller program structure. .................................................... 87
3.38: Interrupt service routine program. ........................................................... 89
3.39: High level hydrogen injection control loop user interface. ............................... 92
3.40: Integration of high and low level hydrogen injector control loop with the data
acquisition system. .............................................................................. 94
4.1: In-cylinder pressure traces for Diesel, HCCI and DIH2 operation at 2200 rpm
and 5 kW. ........................................................................................103
4.2: Cylinder pressure diagram for diesel operation at (a) 5800 W, (b) 3380 W, (c)
1770 W engine load. 2200 RPM, Ta= 21ºC. .................................................105
4.3: Exhaust gas temperature, ignition pressure and maximum combustion pressure
at (a) 5800W, (b) 5080W, (c) 3380W, (d) 1770W. .........................................106
4.4: Diesel fuel consumption and brake thermal efficiency as a function of engine
load. (2200 RPM, Ta= 21ºC.) ..................................................................106
4.5: Open pressure diagram and its derivative of test engine operated in diesel
mode at 5.8 kW load. ..........................................................................107
4.6: Brake thermal efficiency as a function of load for various hydrogen flow rates.......110
4.7: Maximum combustion pressure for different hydrogen flow rates compared
with diesel-only operation. ...................................................................110
4.8: Comparison of exhaust gas temperatures between diesel and various hydrogen
flows. .............................................................................................111
4.9: Hydrogen slip into the exhaust gases for different hydrogen flow rates and
engine loads. ....................................................................................112
4.10: Comparision of the effect of hydrogen addition on the NOx emissions for
different engine loads. ........................................................................112
4.11: Particulate matter emissions compared for various hydrogem flowrates and
diesel operation. ...............................................................................113
4.12: Dependence of the RPR as a function of Tair inlet and λ. ................................114
4.13: Angle of ignition as a function of air inlet temperature, Tair. ...........................115
4.14: Brake thermal efficiency at constant speed (2200 rpm) for varying fuel air
ratios. ............................................................................................116
xvii
4.15: HCCI open cycle diagrams for different loads. ............................................ 117
4.16: Effect of the air inlet temperature on the excess air ratio and angle of
maximum pressure. (At constant speed of 2000 rpm, and mass flow rate
9g/minute of H2.).............................................................................. 118
4.17: Effect of the air inlet temperature on the maximum combustion pressure and
maximum rate of pressure rise. (At constant speed of 2000 rpm, and mass
flow rate 9g/minute of H2.) .................................................................. 119
4.18: Exhaust gas temperature and maximum combustion pressure as a function of
engine load (with constant Ta=90ºC at 2200 RPM). ....................................... 120
4.19: HCCI open pressure diagram and its derivative of test engine operated at
4.1kW load. ..................................................................................... 121
4.20: Required intake air temperature to sustain combustion as a function λ, and
speed for a 17:1 compression ratio. ........................................................ 122
4.21: Emissions at constant speed (2200 rpm) and air inlet temperature (100ºC) as a
function of air fuel ratio. ..................................................................... 123
4.22: Cylinder pressure-volume plots for H2 HCCI operation. .................................. 124
4.23: Effect of the end-of-compression temperature on the ignition delay of the
hydrogen jet. ................................................................................... 126
4.24: Cylinder pressure diagram and its derivative at 5.0 kW load. .......................... 128
4.25: Rate of pressure rise as a function of engine load for diesel and DIH2
operation. ....................................................................................... 129
4.26: Indicated thermal efficiency for different equivalence ratios φ and different
speeds............................................................................................ 130
4.27: Emissions as a function of engine load under DIH2 and DI Diesel operation. ......... 131
4.28: Engine energy flows considered for thermal efficiency calculation. .................. 132
4.29: Comparison of brake thermal efficiencies of the test engine for four operating
modes tested. ................................................................................. 135
4.30: Hydrogen and methane density as a function of pressure @ 300K ..................... 139
4.31: Hydrogen flow meter. ......................................................................... 141
5.1: High level structure of the simulation code. ................................................ 153
5.2: Engine cycle model structure. ................................................................. 155
5.3: is a representation of the piston crank mechanism. ....................................... 156
5.4: Engine cylinder heat losses. ................................................................... 160
5.5: Engine cylinder head and liner thermal image.............................................. 163
5.6: Valve apertures .................................................................................. 165
xviii
5.7: Valve geometry. ..................................................................................166
5.8: Valve lift characteristics as a function of crank angle .....................................169
5.9: Three adjacent points on a valve lift curve .................................................171
5.10: Specific lift characteristics of a poppet valve. ............................................172
5.11: Four stroke CI engine pressure volume cycle. .............................................180
5.12: Simulation program routines. .................................................................182
5.13: Simulation program human interface. ......................................................185
5.14: Pressure volume diagram. .....................................................................185
5.15: Open pressure diagram pressure as a function of the crank angle. ....................186
5.16: In-cylinder temperature as a function of the crank angle. ..............................186
5.17: Rate of pressure rise as a function of the crank angle. ..................................187
5.18: Rate of energy release as a function of the crank angle. ................................187
5.19: Rate of energy transfer fuel combustion and combustion chamber walls. ............188
5.20: Inlet and exhaust valve areas as a function of crank angle. .............................188
5.21: Variation of induced mass of air and exhaust gases as a function of rank
angle..............................................................................................189
5.22: Cylinder air mass flow rate and its variation with the crank angle. ....................189
5.23: HCCI Hydrogen injection valve. ..............................................................190
5.24: Cross section of the solenoid valve ..........................................................191
5.25: HCCI injection valve fitted on the engine. .................................................191
5.26: Isentropic jet development of hydrogen injection. .......................................194
5.27: Schematic diagram of the under-expanded jet behaviour at the nozzle hole
exit. ..............................................................................................197
5.28: Schematic diagram of the jet development model. ......................................200
5.29: Profile of jet penetration and half jet dispersion angle for an orifice with 1.0
mm diameter derived from experimental data. ...........................................201
5.30: Top - Injector hydraulic actuator free body diagram. Bottom Forces acting on
the needle .......................................................................................202
5.31: Cross section of solenoid actuated hydraulic valve. ......................................204
5.32: Definitions of angles of passages of the injector nozzle .................................207
5.33: Block diagram of hydraulic injector model .................................................209
5.34: Simulation model parameters ................................................................210
xix
5.35: Cylinder pressurization. ....................................................................... 210
5.36: Hydraulic actuator sub model. ............................................................... 211
5.37: Choked flow sub model........................................................................ 211
5.38: Critical flow calculation block. .............................................................. 212
5.39: Subsonic flow model. .......................................................................... 212
5.40: Solenoid sub model. ........................................................................... 213
5.41: Resultant force calculation sub model...................................................... 213
5.42: Overall valve flow model. .................................................................... 214
5.43: Inlet valve flow sub model. ................................................................... 214
5.44: Injection (exhaust) valve sub model. ....................................................... 215
6.1: Comparison between predicted and measured pressure traces and their
derivatives for the HCCI compression ignition engine.................................... 220
6.2: Simulated exhaust gas internal recirculation by reduction of valve overlap
period. ........................................................................................... 222
6.3: Simulated angle of ignition for different air inlet temperatures. ....................... 223
6.4: Dependence of the MRPR as a function of Tair inlet and λ. .............................. 223
6.5: Simulated relationship between the minimum cylinder air inlet temperature
required to maintain combustion and the engine compression ratio. ................. 226
6.6: Simulated effect of the air inlet temperature on the IMEP and indicated
power. ........................................................................................... 227
6.7: Comparison between predicted and measured pressure traces and their
derivatives for DIH2 mode operation. ....................................................... 230
6.8: Open cycle diagram for Study 1. .............................................................. 235
6.9: Rate of change of cylinder pressure for Study 1. ........................................... 235
6.10: Rate of energy release diagram for Study 1. .............................................. 236
6.11: Open cycle pressure diagram for Study 2. ................................................. 237
6.12: Rate of change of cylinder pressure diagram for Study 2. ............................... 237
6.13: Rate of change of cylinder pressure for Study 2. ......................................... 238
6.14: Open cycle diagram for Study 3 .............................................................. 239
6.15: Rate of change of cylinder pressure diagram for Study 3 ................................ 239
6.16: Rate of energy release diagram for Study 3 ............................................... 240
xx
6.17: Engine performance for pulsed injection, Study 4, frequency 10kHz, duty
cycle 40%. .......................................................................................241
6.18: Rate of change of cylinder pressure diagram for Study 4 ................................242
6.19: Rate of energy release diagram for Study 4 ................................................242
6.20: Open pressure diagram for Study 5. .........................................................244
6.21: Rate of change of cylinder pressure diagram for Study 5 ................................244
6.22: Rate of energy release diagram for Study 5 ................................................245
6.23: Open pressure diagram for Study 6 ..........................................................246
6.24: Rate of change of cylinder pressure diagram for Study 6 ................................246
6.25: Rate of energy release for Study 6...........................................................247
6.26: Open pressure diagram for Study 7 ..........................................................248
6.27: Rate of change of cylinder pressure diagram for Study 7 ................................248
6.28: Rate of energy release diagram for Study 7 ................................................249
6.29: Open pressure diagram for Study 8 ..........................................................250
6.30: Rate of change of cylinder pressure diagram for Study 8 ................................251
6.31: Rate of energy release diagram for Study 8 ................................................251
6.32: Miller cycle illustrations .......................................................................253
6.33: Relationship between thermal efficiency and hydrogen fuel rate for
conventional and Miller cycle inlet valve settings ........................................255
6.34: Relationship between indicated power and hydrogen fuel rate for
conventional and Miller cycle inlet valve settings ........................................255
6.35: Relationship between MRPR and hydrogen fuel rate for conventional and
Miller cycle inlet valve settings ..............................................................256
6.36: DIH2 Injector view ..............................................................................258
6.37: Injector actuator speed for an actuator and spring mass of 5 g ........................259
6.38: Injector actuator speed for an actuator and spring mass of 50g ........................259
6.39: Injector needle valve displacement for an actuator and spring mass of 5 g ..........260
6.40: Injector needle valve displacement for an actuator and spring mass of 50 g ........260
6.41: Actuator speed for a duty cycle of 5 ........................................................261
6.42: Actuator speed for a duty cycle of 10%. ....................................................262
6.43: Actuator speed for a duty cycle of 20% .....................................................262
6.44: Actuator speed for a duty cycle of 30% .....................................................263
xxi
6.45: Injector mass flow rate for a duty cycle of 5%. ........................................... 264
6.46: Injector mass flow rate for a duty cycle of 10%. .......................................... 264
6.47: Injector mass flow rate for a duty cycle of 20%. .......................................... 265
6.48: Injector mass flow rate for a duty cycle of 30%. .......................................... 265
6.49: Injector mass flow rate for a duty cycle of 50% ........................................... 266
6.50: Actuator speed with period of injection 0.024 s (5000 RPM) ............................ 267
6.51: Mass flow rate with period of injection 0.024 s (5000 RPM)............................. 267
6.52: Actuator speed with period of injection 0.03 s (4000 RPM) ............................. 268
6.53: Mass flow rate with period of injection 0.03 s (4000 RPM) .............................. 268
6.54: Actuator speed with period of injection 0.0333 s (3600 RPM) .......................... 269
6.55: Mass flow rate with period of injection 0.0333 s (3600 RPM) ........................... 269
6.56: Actuator speed with period of injection 0.0428 sec (2800 RPM) ....................... 270
6.57: Mass flow rate with period of injection 0.0428 s (2800 RPM) ........................... 270
6.58: Actuator speed with period of injection 0.0545 s (2200 RPM) .......................... 271
6.59: Mass flow rate with period of injection 0.0545 s (2200 RPM) ........................... 271
6.60: Actuator speed for 200 bar hydraulic pressure ............................................ 272
6.61: Actuator speed for 150 bar hydraulic pressure ............................................ 273
6.62: Actuator speed for 100 bar hydraulic pressure ............................................ 273
6.63: Actuator speed for 50 bar hydraulic pressure ............................................. 274
6.64: Mass flow rate for 50 bar hydraulic pressure .............................................. 274
6.65: Mass flow rate for 100 bar hydraulic pressure ............................................. 275
6.66: Mass flow rate for 150 bar hydraulic pressure ............................................. 275
6.67: Mass flow rate for 200 bar hydraulic pressure ............................................. 276
6.68: Relationship between hydrogen mass flow rate per injection and hydraulic
actuation pressure ............................................................................. 276
6.69: Actuator speed for a static force of 125 N ................................................. 277
6.70: Actuator speed for a static force of 250 N ................................................. 278
6.71: Actuator speed for a static force of 500 N ................................................. 278
6.72: Actuator speed for a static force of 750 N ................................................. 279
6.73: Relationship between the speed of response and the static spring load .............. 279
xxii
7.1: Main bearing dimensions. ...................................................................299
7.2: Ignition angle control through inlet air heating using an exhaust gases heat
exchanger ........................................................................................301
7.3: Combustion control through cylinder charge heating by recirculation of exhaust
gases. .............................................................................................302
7.4: Exhaust gas internal recirculation by reduction of valve overlap period. ..............303
7.5: Working principle of magnetostrictive materials ...........................................304
7.6: Basic electric actuating circuit. ...............................................................305
7.7: Strain magnetic field intensity of Terfenol-D ...............................................306
7.8: Various shapes of Terfenol-D ...............................................................307
7.9: Stress-strain comparison for various selected active materials ....................308
7.10: Terfenol-D temperature saturation strain. .............................................309
7.11: Terfenol-D based hydrogen injector. ....................................................310
xxiii
xxiv
List of Tables:
2.1: Comparison of hydrogen and methane physical properties. ............................... 10
3.1: Specification of the hydraulic engine pump. ................................................. 47
3.2: HCCI Injector flow rate (mg/injection) data for an average supply pressure of
2.47 bar. .......................................................................................... 78
3.3: HCCI Injector flow rate (mg/injection) data for an average supply pressure of
7.58 bar. .......................................................................................... 78
3.4: HCCI Injector flow rate (mg/injection) data for an average supply pressure of
10.96 bar.......................................................................................... 79
3.5: HCCI Injector flow rate (mg/injection) data for an average supply pressure of
14.5 bar. .......................................................................................... 79
3.6: DIH2 injector flow rate data for an average supply pressure of 60 bar. ................. 81
3.7: DIH2 injection flow rate data for an average supply pressure of 70 bar. ................ 81
3.8: DIH2 injection flow rate data for an average supply pressure of 80 bar. ................ 82
3.9 Low level hydrogen injector control loop microcontroller hardware functions. ....... 91
4.1: Cylinder pressure trace main harmonic components their frequencies and
amplitudes. ......................................................................................102
4.2: Maximum observed values of TEXH , PMAX and PIGN at (a) 5800 W, (b) 3380 W, (c)
1770 W engine load. (2200 RPM, Ta= 21ºC.) ...............................................104
4.3: Energy share ratios for hydrogen and diesel fuel at different engine loads for a
constant hydrogen flow of 6.0 dm3/min.. ..................................................109
4.4: Comparison of emissions for DI Diesel and H2 HCCI operation ............................124
4.5: Combustion characteristics as a function of injection timing and duration (2000
rpm, λ = 5.395).. ...............................................................................129
4.6: Comparison of engine energy balance and thermal efficiency at the maximum
reached power at a speed of 2200 RPM. ...................................................134
4.7: Units and values used for the determination of uncertainties. ...........................143
4.8: Summary of uncertainties associated with the transducers ...............................144
5.1: The under-expanded flow equations. .........................................................198
xxv
6.1: Comparison between simulated and measured results for HCCI mode of
operation. ....................................................................................... 219
6.2: Comparison between simulated and measured results for the DIH2 mode of
operation at 6.0 kW load. .................................................................... 229
6.3: Operating parameters of the engine for injection timing and duration
simulation studies. ............................................................................. 233
6.4: Engine performance for continuous injection, Study 1. ................................... 234
6.5: Engine performance for continuous injection, Study 2. ................................... 236
6.6: Engine performance for continuous injection, Study 3. ................................... 238
6.7: Engine performance for pulsed injection, Study 4, frequency 10kHz, duty cycle
40%. .............................................................................................. 241
6.8: Engine performance for pulsed injection, Study 5, frequency 10kHz, duty cycle
40%. .............................................................................................. 243
6.9: Engine performance for pulsed injection, Study 6, frequency 10kHz, duty cycle
40%. .............................................................................................. 245
6.10: Engine performance for pulsed injection, Study 7, frequency 10kHz, duty
cycle 40%. ....................................................................................... 247
6.11: Engine performance for pulsed injection, Study 8, frequency 10kHz, duty
cycle 40%. ....................................................................................... 250
6.12: DIH2 engine parameters for different injector profile and timing ...................... 254
6.13: Injector parameters for dynamic simulation ............................................... 257
7.1: Terfenol-D mechanical properties .................................................................... 307
xxvi
Nomenclature
A
Area [m2]
A
Orifice area [m2]
A
Area of throat at valve seat [m2]
A0
Supply orifice area [m2]
Ap
Actuator piston area [m2]
ASACK
Lateral area of the cone inside the sack volume [m3]
ATDC
After Top Dead Centre
B
Cylinder Bore [m]
BDC
Bottom Dead Centre
BTDC
Before Top Dead Centre
CAD
Crank angle degrees [º]
Cd
Valve Discharge coefficient [1]
cd
Coefficient of discharge [1]
Cp
Specific heat constant pressure [kJ/kg K]
Cv
Specific heat constant volume [kJ/kg K]
COV
Coefficient of variation
dMD
Diameter at Mach disc [m]
EGR
Exhaust Gas Recirculation
ER
Equivalence ratio [1]
g
Length of air gap [m]
H
Magnetic field intensity [A/m]
h
Heat transfer coefficient [W/m2K]
xxvii
I
Solenoid current [A]
IMEP
Indicated mean effective pressure [Pa]
k
Spring elastic constant [N/m]
k0
Flow coefficient [1]
krw
Flow coefficient [1]
ks
Spring constant [N/m]
ksp
Spring constant [N/m]
L
Stroke length [m]
LBARREL
Barrel length [m]
Lexh
Exhaust valve lift [m]
Linlet
Work inlet valve [m]
Lsteel
Magnetic circuit length in steel [m]
LHV
Lower Calorific Value [kJ/kg]
M
Mass [kg]
M
Mach number [1]
MMF
Magnetic motive force [N]
Mp
Net actuator mass [kg]
m
Mass [kg]

m
Mass flow rate [kg/s]
mt
mass of the needle actuator group [kg]
N
Number of turns
Ne
Engine speed [rpm]
P
Pressure [Pa]
P0
Pressure of the sack volume [m3]
xxviii
P0*
Pressure at nozzle exit [Pa]
P1
H2 upstream pressure [Pa]
P2
In cylinder pressure [Pa]
Pa
Atmospheric pressure [Pa]
Pa
Pressure at Mach disc [Pa]
Pc
Critical pressure [Pa]
Pcomp
Compression Pressure [Pa]
Pcp
Combustion chamber pressure [Pa]
pcr
Gas pressure at crevice [Pa]
Pe
Pressure at sack volume [Pa]
Pexp
Expansion Pressure (36ºATDC) [Pa]
P H2
Hydrogen supply pressure [Pa]
Phyd
Hydraulic pressure of brake [Pa]
PHYD
Hydraulic oil pressure [Pa]
Pmax
Maximum combustion pressure [Pa]
pmot
Motored pressure [Pa]
pr
Reference pressure [Pa]
Ps
Supply pressure [Pa]
Pt
Pressure at throat [Pa]
PWM
Pulse Width Modulation
Q
Heat flow [W]
Qstatic
Static mass flow rate [kg/s]
qnet
flow that makes the actuator to move upwords [kg/s]
R
Actuator radius [m]
xxix
R
Coil resistance [Ω]
R
Nozzle radius [m]
Re
Reynolds number [1]
RPR
Rate of Pressure Rise [bar/º], [Pa/ms]
r
Actuator stem radius [m]
r
Critical pressure ratio [1]
S
Stroke [m]
Sp
Average piston speed [m/s]
T
Torque [Nm]
T
Temperature [ºC]
T
Upstream H2 temperature [K]
T*
Temperature at nozzle exit [K]
T0
Temperature of hydrogen in the sack volume [K]
Ta
Ambient temperature [ºC]
Tair inlet
Air inlet temperature [ºC]
Tchamber
Gas temperature at the combustion chamber [K]
Tcr
Gas temperatures at the crevices [K]
Texh
Exhaust gas temperature [ºC]
Tign
Ignition temperature [ºC]
TMD
Temperature at Mach disc
Tr
Reference temperature [k]
t
Time [s]
U*
Speed at nozzle exit [m/s]
U0
Velocity of hydrogen in the sack volume [m/s]
UMD
Hydrogen speed at Mach disc [m/s]
xxx
u
Velocity [m/s]
V
Volume [m3]
v
Linear speed [m/s]
v
Velocity of the needle actuator [m/s]
Vd
Displaced volume [m3]
Vr
Reference volume [m3]
Vsol
Solenoid Voltage [V]
Vt
Velocity at throat [m/s]
X
Position [m]
X
Armature position [m]
X
Ball travel [m]
x
Displacement [m]
Xp
Actuator piston position [m]
α
Angle [rad]
β
Bulk modulus of the oil [Pa]
γ
Specific heat ratio Cp/Cv [1]
Δ
Difference
δ
Difference
θ
Crank angle [rad]
θ
Penetration cone angle[rad]
λ
Air fuel ratio [1]
λ
Excess air factor [λ=1/φ]
μ
Magnetic permeability [H/m]
μ0
Magnetic permeability of air [H/m]
ρ
Density [kg/m3]
xxxi
ρ*
Density at nozzle exit [kg/m3]
ρ0
Density of Hydrogen in the sack volume [kg/m3]
ρMD
Density at Mach disc [kg/m3]
φ
Magnetic flux [Wb]
φ
Equivalence ratio [1]
φ
Half angle of the injector needle tip cone [rad]
Φ dMD
Mach disc diameter [m]
ω
Angular velocity [rad/s]
xxxii
Chapter 1
Introduction
«A problem is a chance for you to do your best! »
Duke of Ellington
Compression ignition (CI) internal combustion engines have been on the
market for more than one hundred years, being and having been the “work
horse” of the power generation and transport industries. These engines are
known for their ability to burn a wide variety of fuels, from gasified biomass
to heavy fuel oils, and even pulverized coal. Despite a great deal of effort on
the development of other concepts of prime movers in recent years, industry
is still very much dependent on CI engines. The reasons for this are concerned
with fuel efficiency, reliability and running costs, in which CI engines often
provide superior performance. Despite much research on alternative
technologies, it is generally accepted that the internal combustion engine will
play a critical role in power generation for years to come.
Governments, scientific communities, industry and the general public have
become increasingly aware of environmental effects resulting from the
extensive use of hydrocarbon fuels as a source of energy. Use of hydrocarbon
fuels, whether derived from the crude oil or from vegetable oils, for land and
sea based transport and power generation results in the emission to the
atmosphere of significant quantities of carbon dioxide and other pollutants.
Therefore, the development of more efficient plants and the use of non
carbon based sources of energy are two fronts of development of new
environmentally friendly power generation. Much of this research work is
1
driven by the need for engines to comply with ever-tightening environmental
legislation imposed by governments worldwide, requiring drastic reductions in
emissions which pose health risks to humans, such as carbon monoxide, nitrous
oxides, volatile organic compounds and particulates.
The research work presented in this thesis was conducted to improve the
understanding of the practical options and respective design features which
should be considered if a CI engine is to be operated using hydrogen as the
fuel.
1.1 Alternative power generating systems
To comply with current and future environmental regulations, there is
currently an increasing interest into technologies that were developed in the
past and were abandoned because they were not economically attractive, and
new or unconventional technologies that make use of alternative fuels.
1.1.1 Stationary power generation
The majority production of electric power worldwide is based on fossil fuels.
For example, in the UK more than two thirds of electric power generation is
provided by gas or coal fired power stations [ref1]. In such plants a
considerable part of the energy supplied cannot be used, since it is rejected as
low temperature heat and transport to the final consumer centres is not
feasible. Much research is therefore being undertaken in the areas of
decentralised production of combined heat and power to reduce overall
losses. The production of heat and power closer of the consumers allows more
flexibility of the generating plant as well as direct use of the thermal power
available, therefore resulting in a increased total efficiency. However,
decentralised power production is not sufficient to achieve the required
2
reductions in carbon dioxide emissions set out in the Kyoto Protocol and the
Climate Change Bill 2007. Such drastic reductions can probably only be
obtained by increasing the use of renewable fuels.
1.1.2 Propulsion systems
The automotive industry is presently offering a number of technologies that
are claimed to provide near zero emissions. Both hybrid electric and fully
electric power trains are presently being offered. However, some other
technologies are under development, most prominently the use of hydrogen as
an energy carrier for use in fuel cells and internal combustion engines. Since
hydrogen can be produces from renewable sources, this has a high potential to
become classified as a zero carbon emission technology. The adoption of
hydrogen as a fuel by the car industry has been politically supported through
governmental funding of research into hydrogen technologies, but many
challenges still exist. The use of hydrogen as an automotive fuel continues to
be under intense research and development in many countries, in particular
Germany and the United States. It is widely agreed that if the problems
related to cost-effective hydrogen production, safe and compact on-board
hydrogen storage, fuel cell reliability and operational safety can be resolved,
then this technology has substantial potential.
The use of hydrogen on board marine vessels is another interesting
application, as large amounts of low temperature heat can be recovered and
potentially used to produce hydrogen. Hydrogen can then be stored and used
as a combustion improver or even as a main fuel for power production in
diesel engines, either at sea or, particularly, in port where exhaust emissions
regulations are stricter. The concept of on board production and use of
hydrogen as a fuel can be seamlessly integrated with the all-electric ship
concept.
3
1.2 Internal combustion engines
The principle design of today's internal combustion engines have a similar form
to
that
we
know
since
mid-19th
the
century.
Although
significant
improvements in engine performance have been achieved, those are mainly
due to developments in materials, manufacturing and control engineering.
Despite the fact that the first principles remain the same as 100 years ago,
new design challenges driven by the need for improved thermal efficiencies
and
lower
exhaust
emissions
are
now
being
studied
using
refined
computational models for each engine system, such as engine control, fuel
injection, knocking control, supercharging, etc.
1.2.1 The use of hydrogen in CI engines
There are various methods of using hydrogen as a fuel in CI engines and the
ones addressed in this research are:
a) Fumigation of hydrogen. This method is the easiest way of using hydrogen
in a CI engine and can be divided into two categories:
 fumigation of hydrogen in the inlet air manifold at a pressure slightly
above atmospheric pressure (typically about 300 mbar), with ignition
controlled by diesel fuel injection, and inlet port injection during the
time interval corresponding to the induction stroke, while the inlet
valve is open and the exhaust valve is closed; the cylinder charge
ignition being controlled by diesel fuel injection.
The main difference between these two methods is that the second one makes
better use of the hydrogen charge, since hydrogen is injected only during the
engine induction stroke and while the exhaust valve is already closed. As a
result, there is a reduction in the concentration of hydrogen in the exhaust
gases, since no hydrogen can pass through the cylinder during the valve
overlap period.
4
These two forms of hydrogen use in CI engines have similar characteristics, are
relatively simple to implement, and make use of modest hydrogen pressures.
Nevertheless, there are problems associated with risk of inlet manifold
explosions, engine power de-rating due to the displacement of intake air when
injecting hydrogen and the potential for hydrogen slip into the exhaust gases.
b) Homogeneous Charge Compression Ignition (HCCI).
Hydrogen HCCI can be achieved using a high compression ratio diesel engine,
typically above 20:1 is required. With such a high compression ratio, the final
temperature of compression will be sufficiently high to ignite the cylinder
charge. This method typically uses timed injection of hydrogen at a low
pressure in the engine inlet manifold. Hydrogen is injected only during the
engine induction stroke, while the exhaust valve is already closed. This
method has interesting characteristics, including a potential for high engine
thermal efficiency and extremely low exhaust gas emissions. There are,
however, some important problems which need to be solved, such as engine
load and speed control, mechanical component loading, and also the
possibility of air manifold explosions.
c) Direct injection.
The direct injection of hydrogen constitutes possibly the most promising
method of hydrogen use in CI engines.
The method can be divided into two slightly different concepts: one that
makes use of moderate hydrogen pressures and a second that makes use of
high hydrogen pressures. In both methods, the injection takes place only when
the cylinder valves are closed. In the first mode, the fuel is injected at low
pressure early in the compression process and ignition takes place only when
the final temperature of compression is reached, making the ignition angle
slightly erratic and difficult to control. With the high pressure direct injection
5
method, hydrogen is injected only when the final compression temperature is
above the self-ignition temperature of the hydrogen charge. (Similarly as in a
standard diesel engine.)
The main advantages of the high pressure direct injection are:

There is no power reduction due to displacement of intake air, hence a
direct injection hydrogen engine will have a higher maximum power
output compared with pre-mixed operation. The exhaust gas emissions
are well controlled, since the engine can be operated very lean, in this
way controlling in particular the NOx production. Because hydrogen is
injected only with the cylinder closed, no short-circuiting or air
manifold explosions are possible, providing safer engine operation.

Since hydrogen direct injection makes it possible to control the heat
input per cycle accurately, good engine load control is achievable.

Control of the ignition timing is achieved through controlling the start
of injection, and an optimized injection pattern can be used for each
engine load.
As will be shown later, there are difficulties essentially related with the
limitation of the rate of pressure rise, which must be limited to acceptable
levels to avoid mechanical damage.
1.3 Contribution to existing research
The amount of research reports describing the use of hydrogen as a fuel for CI
engines is very low compared to the vast amount of research on conventional
fuels. However, several authors have presented studies of the performance of
hydrogen as a fuel for spark ignition engines, and some reports exist
presenting results from compression ignition test chambers, though not
resulting in the establishment of general rules or concepts.
6
This thesis presents contributions to the existing state of the art in the form of
detailed investigation into the performance and operational characteristics of
Homogeneous Charge Compression Ignition (HCCI) and Direct Injection
Hydrogen (DIH2) engines.
The research work included the simulation and development of hydrogen
injection systems for HCCI and DIH2 operation. A single cylinder test engine for
hydrogen fuelled operation was developed, and extensive test results are
presented. A direct comparison of each mode of hydrogen operation, as well
as conventional diesel-fuelled operation, is presented, allowing an evaluation
of the potential advantages of hydrogen engine operation. Computational
simulation models have been used to investigate the general operational
characteristics more widely, including injection timing, knocking and thermal
efficiency for both modes of operation studied.
The HCCI and DIH2 operation modes of CI engines were evaluated in relation to
the need for future engine technology to allow efficient operation with
hydrogen as an alternative fuel for near zero carbon emissions power
generation. The thesis has the following structure:

Chapter 2 presents a detailed background study, thoroughly evaluating
the particular features of the use of hydrogen as a fuel for reciprocating
engines. Also, a review of reported hydrogen CI engines application and
their performance is presented.

Based on Chapter 2, Chapter 3 describes how the experiments were
designed. The various components of the experimentation test rigs that
were developed are presented, as well as the justification for the
chosen designs and some of the challenges encountered.

Chapter 4 presents the results of the tests carried out to set up the
hydrogen injection systems and the HCCI and DIH2 operation modes
respectively. In-cylinder process characterization of the two modes of
7
operation was carried out using a comprehensive data acquisition
system to gather the test data.

Chapter 5 presents full cycle simulation models for HCCI and DIH2
modes of operation. A model of the hydrogen direct injection system is
also presented.

Chapter 6 addresses engine control and thermodynamic performance of
both modes of hydrogen engine operation using the developed
simulation
framework.
Thermodynamic
performance
and
control
considerations resulting from variations in the input variables are
presented, and control strategies are investigated.

Finally in Chapter 7, the results of the research are summarised and
evaluated, and further work is suggested.
8
Chapter 2
Hydrogen engine research review
«None of us are as smart as all of us.»
(Japanese proverb)
This introductory chapter presents an overview of previous research into
hydrogen fuelled internal combustion engines. The characteristics of
hydrogen as an engine fuel is described and the main features of the HCCI
and DIH2 modes of engine operation are presented, establishing the
terminology for the rest of the thesis. Reported advantages and
disadvantages of both modes of operation are reviewed and analysed. The
hydrogen injection technology required and its relationship with the
hydrogen combustion characteristics and engine performance and control
are also introduced.
2.1 Hydrogen utilisation as an engine fuel
In the 17th century, Robert Boyle reported that “combustible air” was
obtained when iron was dissolved in sulphuric acid and hydrochloric acid.
Later, Henry Cavendish recognised the nature of the detonation of this gas
and was able to isolate hydrogen (TUV Suddeutschland, 2003). The
designation of the gas originates from the terms “hydrogène” or
“hydrogenium” which were coined by Lavoisier in the 18th century (Colin,
2001).
The use of hydrogen to produce mechanical work has been attempted by
numerous researchers and inventors. In 1820, the vicar and scientist
William Cecil had developed the first model of a hydrogen internal spark
ignition (SI) combustion engine, and described in a paper how this engine
could be built (Cecil, 1820). In the 1920s, a German engineer, Rudolf Erren,
9
developed a hydrogen gas detonation engine (SI) based on a two stroke
cycle process. This engine was patented in Germany in 1929 and later also
in the United Kingdom (Erren, 1932). Oemichen (1942) reported engine
efficiencies around 50%, converting more than 1000 SI engines to use
hydrogen fuel.
Table 2.1: Comparison of hydrogen and methane physical properties
(Karim, 2003).
Property Density (0.1 MPa, 300 K) [kg/m3]
Stoichiometric composition in air [% by volume]
Stoichiometric fuel/air ratio (mass basis) [1]
Higher heating value [MJ/kg]
Lower heating value [MJ/kg]
Higher heating value [MJ/m3]//3)
Lower heating value [MJ/m3]
Combustion energy (stoich. mixture) [MJ/kg]
Kinematic viscosity (300 K) [mm/s2]
Diffusion coefficient into air (NPT) [cm2/s]
Flammability limits [% by volume]
Minimum ignition energy [mJ]
Laminar flame speed (NTP) [m/s]
Adiabatic flame temperature [K]
Autoignition temperature [K]
Quenching gap (NTP) [mm]
Hydrogen
0.082
29.53
0.029
141.7
119.7
12.10
10.22
3.37
110
0.61
4-75
0.02
1.90
2318
858
0.64
Methane 0.717
9.48
0.058
52.68
46.72
37.71
33.95
2.56
17.2
0.189
5.3-15.0
0.28
0.38
2190
813
2.03
In Table 2.1, some of the key properties of hydrogen that are relevant to its
use as an engine fuel are compared with the corresponding properties of
methane (the main component of natural gas). It is evident that hydrogen
is lighter than methane and requires less air by volume stoichiometric
combustion, while it requires a higher relative mass of air. Its heating value
on mass basis is higher than methane but on volume basis it is much lower.
There is a significant difference between its higher and lower heating
values, since the product of combustion in air is only water. However, its
energy release during combustion per unit mass of stoichiometric mixture is
one of the highest among all the fuels. Hydrogen has a high diffusion
coefficient, benefiting fuel-air mixing and the combustion process. Further
characteristics which influence the behaviour of hydrogen as a fuel for
reciprocating engines include its wide flammable range when mixed with
10
air, permitting extremely lean as well as rich combustion. The amount of
energy required to ignite hydrogen is very low when compared with
methane. This, together with its fast flame speed, results in a fast
development of the combustion, despite of the fact that the self ignition
temperatures
are
similar.
The
thermodynamic
and
heat
transfer
characteristics of hydrogen allow high compression temperatures, resulting
in a better engine efficiency and lean mixture operation. The chemical
kinetics of hydrogen combustion are simple and well understood whereas
the chemical kinetics of hydrocarbon fuel oxidation, in particular complex
hydrocarbons, involve slower endothermic reactions that are associated
with fuel breakdown.
However, there are some disadvantages associated with the use of
hydrogen as a fuel in internal combustion engines. Hydrogen compressed at
200 bar pressure and room temperature only has approximately one third of
the energy density of methane under the same conditions. The mass flow of
intake air is reduced for any engine size because of the relatively high
stoichiometric hydrogen to air ratio. The high combustion rates of hydrogen
produce high peak pressures and temperatures in engines when operating
with near-stoichiometric mixtures. Further, the material used to construct
the engine must be selected carefully, as some materials can react with
hydrogen. Heat transfer losses can be high, unless special attention is paid
to engine heat transfer design.
2.2 Spark ignition hydrogen engine operation
Much of the information reported in the open literature about hydrogen
engines refer to spark ignited engines, and tend to highlight the positive
features of the hydrogen fuelled engines while de-emphasizing or even
ignoring the many limitations associated with such fields of application.
There is a need to focus equally well on these negative hydrogen features
that may need some research effort. Karim,(2003), has dedicated a great
deal of attention to the use of hydrogen as a fuel for reciprocating engines,
in his paper “Hydrogen as a spark ignition engine fuel” presented a realistic
11
survey of positive and negative features of the hydrogen spark ignited
engines, their characterization and the need for further research. A great
deal of research has been directed towards the use of hydrogen as a
combustion improver of natural gas fuelled spark ignited engines, but no
agreement has been obtained in terms of a generalised value for mixture
percentages to use, Yi H. et al., (2000), Karim et al., (1999), , Verhelst et
al., (2001). Aspects of combustion duration, performance and emissions of
hydrogen fuelled SI engines is well explained by Yamin J. et al.,(2000), that
concluded that one of the main parameters affecting engine performance
and emissions is the combustion duration which is controlled by adequate
timing. On this paper it is discussed how the combustion duration is
affected by engine operating parameters such as compression ratio,
equivalence ratio, spark plug location, spark timing and engine speed,
therefore being a reference work on hydrogen SI engines. Finally another
important area of research of hydrogen fuelled SI engines is dedicated to
the discussions around the most appropriate heat loss model and the
respective heat transfer coefficients, T.Shudo et al., (2002); Assanis D. et
al. (2004); however a consensus is far to be obtained, except on the non
adherence to the conventional heat transfer models, as there is almost no
radiation due to the absence of carbon, and therefore the heat transfer
during the hydrogen combustion needs to be adequately modelled.
2.3 Improvement of CI engines performance: bi-fuel
experience
Bi-fuel operation was and is still a way of using hydrogen in CI engines,
although the hydrogen usually burns via flame propagation in such engines.
Hydrogen was designated as such an “auxiliary fuel” by many authors for
quite a long time and well known as an improver of CI engine performance.
Auxiliary fuels were all the fuels that being induced or introduced into the
engine cylinder contributed positively for the improvement of the engine
performance. The simplest way of achieving this was by means of
“fumigation” of such a fuel into the engine inlet air manifolds. From this
12
single point architecture, other architectures were derived, for example
mechanical and electronic auxiliary fuel port injection systems. Ignition of
the auxiliary fuel was obtained by the in-cylinder direct injection of, for
example, diesel fuel.
During the compression stroke the auxiliary fuel has ample time to become
dispersed throughout the cylinder volume. On the basis of the work
reported by various investigators on pre-combustion reactions, it is
speculated that this fuel undergoes pre-flame oxidation but not ignition
when properly proportioned with the air charge. A favourable cylinder
environment is thus created which serves as a homogeneous propagation
environment for the flame when this is triggered by the diesel charge
ignition.
2.3.1 Early work on bi-fuel CI engine operation
The motivation that first led to fumigation operation of diesel engines was
to maintain the diesel engine fuel efficiency while keeping the smoke
within acceptable limits and to allow the used of fuels with poor ignition
quality. Alperstein et al. (1958) reported manifold introduction of fuel into
a compression ignition engine as early as 1941 at Pennsylvania
State
University. Vaporisation and carburetting were used for introducing
auxiliary fuels such as hexane, heptane, different diesel oils, white
gasoline, hydrogen peroxide, benzoyl peroxide, methyl alcohol, benzene,
cetane, and diethyl ether. Cetane, hexane and low boiling point diesel
fuels were found to be most effective, whereas acetone, benzene and ethyl
alcohol were the least effective when introduced into the intake manifold.
McLaughlin (1956) and others used manifold introduction of fuels, mostly
gasoline and liquid petroleum gas, to kill smoke and/or boost power. In
most cases, the main fuel was injected into the cylinder in the
conventional manner and a small amount of auxiliary fuel was introduced
into the intake manifold as an aid to combustion. Also tested was
carburetted alcohol as a main fuel injected and a small quantity of diesel
fuel fumigated into the air intake manifold to ignite the compressed air-
13
alcohol vapour mixture. Using this technique it was possible to burn alcohol
that otherwise could not be ignited and used as a main fuel of a
compression ignition engine.
Derry et al. (1953) reported that by fumigating an auxiliary fuel into the air
inlet manifold it is possible to reach a power increase in the order of 20%
above the maximum rated load, without producing more smoke than when
the engine is operated at full load under normal conditions.
Alperstein et al. (1958) introduced a portion of the fuel charge as a fine
mist into the manifold of a compression ignition engine of open and swirl
chamber types reporting smoke reductions in the order of 80%, an increase
in smoke limited power output of up to 18.5%, and a decrease in specific
fuel consumption of up to 9.8%. Further, shorter ignition lag, lower
maximum rate of pressure rise and smoother operation were found.
Alperstein reported that a diesel engine could operate satisfactorily on
substandard fuels down to a cetane number of zero when fumigation was
employed.
Arnold et al. (1957) carried out a systematic study of the types of
fumigation fuels and their impact on the engine under residual fuel
operation. They concluded that the benefits of bi-fuel operation are not
the same for all types of engines but found some favourable results in tests
using medium and high speed diesel engines. Fumigation allowed lowquality fuels to be effectively burned under conventional engine operating
conditions and with an acceptable exhaust gas smoke level. A reduction in
wear and deposits was found, to a level considerably below that generally
obtained with low quality fuels. Bi-fuel operation allowed smoke limited
power output of an engine to be increased, as well as allowing operation on
fuels normally not considered feasible for use in conventional diesel
engines.
2.3.2 Bi-fuel operation with hydrogen induction
Varde et al. (1983) studied fumigation of hydrogen into the air inlet
manifold of a diesel engine. The focus of this investigation was the
14
reduction of particulates levels in the exhaust by fumigation of small
quantities of gaseous hydrogen. Hydrogen flow rates equivalent to 10% of
the fuelling rate (on an energy basis) reduced the smoke emissions at part
load, however at full load the particulate reductions were more modest
most probably due to the reduced amounts of excess air available in the
cylinder. It was also found that very low hydrogen flow rates had adverse
effects on the engine thermal efficiency but that notable improvements in
efficiency were achieved by increasing the percentage of hydrogen supplied
to the engine. Figure 2.1 shows the effect of hydrogen addition on the
engine brake thermal efficiency for a fixed diesel fuel injection timing and
different hydrogen flow rates. A clear positive effect on the brake thermal
efficiency for certain hydrogen injection rates can be seen.
Figure 2.1: Engine brake thermal efficiency as a function of brake power
for different hydrogen flow rates (Varde et al., 1983).
Figure 2.2 shows the effect of hydrogen addition on the smoke levels in the
diesel engine exhaust. A clear trend of allowing the engine to run at higher
power without being limited by smoke emissions is seen.
15
Figure 2.2: Exhaust gas smoke levels as a function of engine load for
varying hydrogen injection rates (Varde et al., 1983).
A more recent study on bi-fuel operation of diesel engines with hydrogen as
a way to burn low quality residual fuels was carried out by Geisler et al.
(1993). This group performed a number of tests using a medium speed
engine to evaluate the possibilities of using hydrogen and natural gas as
fuels on LH2 carriers from Canada to Europe. In this study, the gaseous fuels
(hydrogen and natural gas) were introduced into the engine cylinder using a
pre-injection injector and a pilot injector. The quantity of hydrogen was
varied from 0% to 100% of the gaseous fuel mixture, allowing a smooth
transfer from natural gas to hydrogen, with the pilot fuel amounting to a
maximum of 7% of the engine heat input rate at full load. It was found that
pre-injection of gaseous fuels significantly reduces the ignition delay, the
cylinder pressure rise rate and the combustion noise, in particular under
part load operation. Also, it was concluded by the authors that using
hydrogen together with low quality fuels with poor ignition quality could be
employed in medium speed diesel engines with pre-injection systems to
decrease the smoke levels. It was identified by the authors that further
increase in the hydrogen fuelling rate will result in too high maximum
combustion pressures and thereby excessive mechanical and thermal loads.
Herbert et al. (1993) carried out work on the same test plant as Geisler et
al., but with port injection solenoid valves. The research aimed at the use
of gaseous hydrogen in a dual fuel stationary engine to minimise CO2
emissions. A number of other findings were presented regarding the
16
beneficial effects of hydrogen use as a fuel for four stroke medium speed
diesel engines, namely in what concerns the thermal efficiency and
reduced emissions of CO2, NOx, CO and unburnt hydrocarbons. The authors
tested experimentally the limits of engine operation with hydrogen by
successively increase of hydrogen content by varying the gas mixture
through the variation of the turbocharger power (an electrically driven
centrifugal compressor). The limit of minimum excess air (or rich mixture)
was identified by excessive exhaust gas temperatures whereas the
maximum excess air (or lean mixture) limit was identified through ignition
failures with corresponding emissions and erratic operation. Values of
excess air of λ >2.4 were reached whereas for natural gas excess air ratios
varied between 1.7< λ <1.8. It was concluded that concentration in the fuel
gas mixture of 60% (vol.) was achievable at full load, reducing the CO2
emissions down to 30%. A further increase of hydrogen concentration
results in a reduction of power of about 30% for pure hydrogen operation.
To counteract the engine power de-rating, compression ratio control should
be adopted in such a way that for higher engine loads the compression ratio
is decreased, allowing the combustion of a larger hydrogen quantity
without excessive combustion pressures and heavy knocking. Some
hydrogen slip into the exhaust gas was experienced in this research,
influenced by the timing of the port hydrogen injection, leading to direct
passage of hydrogen into the exhaust channel before the cylinder is closed.
This research work also suggested that the use of hydrogen direct injection
into the cylinder should be adopted to avoid flash-back and pre-ignition.
2.4 Compression ignition hydrogen engines
As discussed previously, the use of hydrogen as a fuel in spark ignition
engines was first investigated around 1820, however hydrogen use in
compression ignition engines was not seriously investigated until the 1990s.
The information available on direct injection of hydrogen (DIH2), is limited
to a few academic research reports. One project, carried out by the
technical university of Munich in cooperation with among others MAN under
the WE-NET Phase II programme, aimed at the development of a hydrogen-
17
fuelled single cylinder CI engine of 100 kW power output and with a
thermal efficiency above 40%.
The amount of research on the use hydrogen in homogeneous charge
compression ignition (HCCI) engines is also scarce; however, a number of
engine research groups around the world have recognised the potential
advantages of this mode of operation in terms of thermal efficiency and
emissions. The main difficulty with HCCI operation is the engine
controllability, and until now no commercial HCCI engine has been offered
on the market.
Figure 2.3 Illustrates the conceptual difference between port injected
systems, in which hydrogen gas is injected close to the intake port but
external to the cylinder, and direct injection, where hydrogen is injected
directly into the closed cylinder.
External cylinder
charge
Port injection
Internal cylinder
charge
Direct injection
Figure 2.3 Internal and external cylinder charge formation.
2.4.1 The DIH2 engine
The stratified charge compression ignition engine is characterised by the
direct or indirect injection of the hydrogen fuel into the cylinder during the
compression stroke and after the closure of both exhaust and intake valves.
The thermal energy required for the ignition of the injected fuel is
provided by the compression of the air contained inside the cylinder.
18
The temperature of the cylinder charge depends on a number of
operational variables, in particular on the air inlet temperature, engine
load, speed and residual fuel temperature. Despite this, it is possible to
ensure that the temperature of the charge at the time of fuel injection is
above the ignition temperature or very close. In this way the angle of
ignition can be controlled by the angle of injection. Direct injection
engines are further characterised by:

acceptable controllability, in particular when load variations are
present during operation;

high thermal efficiencies;

very lean combustion possible; and

low emissions, in particular NOx.
Among the most recent research on diesel engines fuelled with hydrogen
are the reports by Fukuma et al. (1986) and Welch et al. (1990). Fukuma
studied hydrogen direct injection where cylinder charge ignition was
achieved by using a glow plug as a hot surface and one injector nozzle with
only one hole. The conclusions from this research work was later revisited
as engine performance deterioration due to a slow flame propagation
through the heterogeneous mixture was identified, and better performance
was achieved with an eight-hole nozzle. The use of a hot surface to assist
ignition in a direct injection hydrogen engine was followed by Welch. This
group concluded that the hydrogen fuelled diesel engine with glow plug
ignition develops more power with lower emissions than the same engine
operated with diesel fuel. Indicated thermal efficiencies for lower brake
loads were above 50%. Welch et al. also identified an exponential
dependence between the rate of pressure rise and the engine load. This
increase can be explained by an increased thermal load of the cylinder and
a higher temperature at the end of compression due to a larger mass of hot
residual gases.
The work reported by Rottengruber et al. (2004) is considered a reference
work for a commercial CI engine because of the methodology and
19
consistency of results. The test engine used in this research was a MAN type
1L24/30 single cylinder engine with a bore of 240 mm and a stroke of 300
mm. The original diesel fuel injection system was kept in place and a
hydraulically actuated hydrogen injection system with electronic control
was installed, allowing control of the start and duration of the hydrogen
injection. Figure 2.4 illustrates the test rig, showing the separate fuel
systems and the injector hydraulic system. The engine was not
turbocharged; instead an air fan driven by an electric motor was used. An
air heater was installed to assist during starting and to study the effects of
the inlet air temperature on engine operation.
Figure 2.4: Hydrogen engine and sub-systems experimental setup
(Rottengruber et al., 2004).
Using a dynamic pressure transducer to measure the in-cylinder pressure
development it was identified that the DIH2 combustion is of very short
duration, commencing at the initiation of the injection and terminating
with the end of the injection. Figure 2.5 shows among other things the
measured pressure trace and calculated rate of heat release as a function
20
of the crank angle for a start of injection at 350º, injection duration 47º
and an engine load corresponding to bmep = 18.17 bar. Also, it is evident
from the heat release plot in Figure 2.5 that the start of combustion and its
end are almost coincident with the injector needle lift. This suggests that
the ignition delay is small and that good engine control can be achieved
through control of the hydrogen injection timing.
Figure 2.5: Engine performance under direct injection hydrogen operation
(Rottengruber et al., 2004).
Figure 2.6: Cycle-to-cycle variations in cylinder pressure under DIH2
operation (Rottengruber et al., 2004).
21
Rottengruber et al. also analysed the cycle-to-cycle variations in cylinder
pressure as shown in Figure 2.6. It was concluded that variations in the
pressure trace tended to be less for higher engine loads and more
pronounced for medium to light loads. It was also reported that the nozzle
geometry has little influence on the indicated thermal efficiency for engine
loads varying between indicated mean effective pressures from 8 to 12 bar.
Under these conditions, the indicated efficiency stayed around 50%, as
illustrated in Figure 2.7.
Figure 2.7: Indicated thermal efficiency as a function of brake mean
effective pressure for different injector nozzle designs (Rottengruber et
al., 2004).
In summary, Rottengruber et al. found that the use of hydrogen as a fuel in
CI engines has the following advantages in comparison to hydrogen fuelled
spark ignition engines: good engine control; no danger of combustion in the
intake manifolds; absence of knocking; increased power output; and
increased thermal efficiency.
22
Research work carried out by Naber et al. (1998) focussed on the
characterization of the combustion of hydrogen in a CI engine and
considered the variation of parameters such as injection pressure and
temperature, nozzle orifice diameter, and ambient gas pressure and
temperature. Results from the tests are illustrated in Figures 2.8-2.13.
From Figure 2.8 it can be seen that hydrogen ignition delay is dependent on
the temperature of the cylinder charge. The same type of dependency can
be seen in Figure 2.9 for the effect of cylinder charge density. It was
concluded that the ignition delay of hydrogen under direct injection
operation varies exponentially with temperature; the influence of other
parameters was not significant. Ignition delays of approximately 1.0 ms
were observed for injection at TDC.
Figure 2.8: Effect of ambient gas temperature on hydrogen ignition delay
(Naber et al., 1998).
23
Figure 2.9: Effect of the ambient gas density () on hydrogen ignition delay
(Naber et al., 1998).
The effect of oxygen concentration on the ignition delay was also studied
and ignition of the cylinder charge with concentration of oxygen as low as
5% (by volume) was achieved. This suggests that the rates of combustion
are insensitive to reduced oxygen concentrations and that there are good
burning characteristics in tight volumes of the cylinder, such as crevices,
where oxygen concentrations are usually low. In turn this means that there
is a potential for a reduction in emissions.
Figure 2.10: Effect of O2 concentration on ignition delay (Naber et al.,
1998).
24
Figure 2.11: Effect of fuel temperature on ignition delay (Naber et al.,
1998).
As illustrated in Figure 2.11, the ignition delay is also dependent on
hydrogen temperature at injection.
This is comparable with the
dependence of ignition delay on cylinder charge temperature, underlining
the strong effect of the charge temperature on that parameter. The effect
of cylinder charge oxygen concentration on the pressure rise is illustrated
in Figure 2.12, showing that for all concentrations of oxygen the rate of
pressure rise is the same.
25
Figure 2.12: Effect of ambient gas O2 concentration on H2 combustion and
pressure rise with ambient temperature 1000 K (Naber et al., 1998).
The effect of temperature on the rate of heat release is illustrated in
Figure 2.13, indicating that the oxygen concentration and temperature of
the cylinder charge has an negligible effect on the rate of combustion and
heat release.
Figure 2.13: Effect of ambient air temperature and O2 concentration on the
rate of heat release (Naber et al., 1998).
26
These findings are in agreement with those of Rottengruber, further
indicating that compression ignition of hydrogen is possible in a diesel
engine.
2.4.2 DIH2 engine specific problems
Due to its working principle, the DIH2 compression ignition engine can
provide good speed and load control, however control of the in-cylinder gas
pressure and pressure rise rates, and the consequent mechanical stress on
the piston rings and piston crank mechanism, needs to be addressed to
ensure reliable engine operation (Rotengrubber et al., 2004, Naber et al.,
1998). The DIH2 engine should therefore be equipped with an electronic
hydrogen injection system, to allow accurate control of the quantity of
hydrogen being injected as well as the manner of injection. This can be
achieved through pulsed injection, as described in Chapter 4 below; this
gives rise to a certain amount of recirculation of the combustion products
which is an effective way of controlling the rate of pressure rise.
In summary, the feasibility of the DIH2 engine concept has been
demonstrated, and it is currently at a stage where it is ready to undergo
operational tests that will give better insight into engine performance and
operational characteristics.
2.4.3 The HCCI hydrogen engine
A
hydrogen
homogeneous
charge
compression
ignition
engine
is
characterised by the induction or injection of hydrogen during the air
intake stroke, after the exhaust valve closure. As is the case for the
stratified charge compression ignition engine, the thermal energy required
for the ignition of the cylinder charge is provided by the compression of the
trapped volume of air contained inside the cylinder, provided by the
crankshaft/connecting rod mechanism. Unlike the stratified charge
compression ignition engine, the ignition of the cylinder charge will take
27
place at the crank angle at which the ignition temperature is reached. This
means that the angle of ignition can be somewhat erratic as the
temperature of the charge can vary between cycles, depending on a
number of operational variables, in particular the air inlet temperature,
engine load, speed and residual gas temperature. The quantity of hydrogen
that can be used per cycle is limited by the volume air displaced, resulting
in a power limitation compared with direct injection operation. Hydrogen
HCCI engines are also characterised by:

reduced controllability, in particular when load variations are
present during operation;

high thermal efficiencies due to fast combustion; and

homogeneous and very lean combustion possible, giving reduced
emissions, in particular NOx.
Pursuing the objective of achieving extremely low exhaust emissions and
high thermal efficiencies, researchers early identified the HCCI engine as a
strong candidate. A detailed study of four stroke HCCI natural gas fuelled
engines was carried by Fiveland et al. (2000), in which indicated thermal
efficiencies in the range of 55% were reported. This work further
demonstrated that a prediction of heat transfer losses in HCCI engines using
the Woschni method differs substantially from the observed heat transfer
rates. The main reason for this is due to the database from which the
Woschni coefficients are derived, and also due to the ignition angle shift
which can be as big as 10ºCA. Fiveland et al. conducted a number of tests
varying the compression ratio through active control of the inlet valve and
natural gas injection pressure, concluding that the lowest compression
ratio and the highest injection pressure resulted into the highest thermal
efficiency and power output. The most favourable value for the
compression ratio and gas injection pressure was found to be respectively
17:1 and the 3.0 bar.
HCCI hydrogen fuelled engines were also investigated by Stenlaas et al.
(2004). This research work focussed on the characterization of HCCI
hydrogen engines in terms of efficiency, combustion phasing and emissions
28
using a single cylinder engine modified to allow variation of the air fuel
ratio, speed, compression ratio and air intake temperature. Stenlaas
reported that engine operation with very lean air to fuel ratios, up to 6,
was possible. The hydrogen temperature was used to control the angle of
ignition but it was found that this was not ideal as control was very limited,
in particular for richer cylinder charges. It was also identified that the
power developed by the HCCI hydrogen engine was about half the power
developed by the same engine when fuelled with other fuels. The NOx
emissions decreased with the increase of the air fuel ratio, as expected.
2.4.4 HCCI hydrogen engine specific problems
As stated in the literature, the HCCI engine concept presents a great
potential for reduced emissions and high thermal efficiencies, but a
number of problems remain to be solved. This includes engine load and
speed control, operational stability and also the reduction in power output
compared with using direct injection. These problems become more
difficult to solve when using hydrogen fuel because the combustion is faster
and the quantity of hydrogen introduced in the cylinder per cycle is limited
by its specific volume. A number of methods have been proposed to control
the angle of ignition, for example the use of a secondary fuel with a wellestablished ignition angle (Derry et al., 1953), or the use of a spark plug
(Rottengruber et al., 2004). However, the first method would not be easy
to achieve in practice and the second method would mean that the engine
would be a spark ignition engine.
There is, however, not much doubt that HCCI and DIH2 hydrogen fuelled
engines will be a valid option in the future for automotive applications, as
stated by Rottengruber et al. (2004). Despite the fact that the research was
carried out using a spark plug to initiate ignition, the authors maintain that
the DIH2 (with spark ignition) concept has the potential to achieve better
performance than conventional gasoline fuelled engines. It was estimated
that the maximum engine power output can be increased by more than 20%
with an indicated thermal efficiency above 33% and that for lower loads an
29
external mixture formation (port injection) can be used to give thermal
efficiencies of more than 40%. Rottengruber et al.2004, therefore
recommended that an engine control strategy contemplating both modes of
hydrogen fuelled engine operation should be implemented.
2.5 Fundamental hydrogen engine specific properties
Comprehensive research results on hydrogen engine combustion have been
presented by Professor Tushio Shoudo from Hokkaido University, and some
of the main findings will be introduced here as they constitute the state of
the art in the area.
Shudo et al., (1999) recognised that reciprocating engines designed for
conventional hydrocarbon fuels have different heat balances when they are
operated with hydrogen. To understand the reasons for such differences, a
number of experiments were carried out using the same engine and
identical operation conditions with hydrogen and methane fuel. Although
the engine used for the experiments was of a spark ignited type, many of
the conclusions will be extendible to CI engines. Based on the relationship
between indicated thermal efficiency, ηi , and the cylinder cooling losses,
Shudo et al. (1999) defined the following characterisation of the heat
balance:
ηi = ηth η glh ηu 1  φw 
(2.1)
where ηth is theoretical thermal efficiency, ηglh is the degree of constant
volume combustion, ηu is combustion efficiency and φw is a cooling loss
ratio given as φw = QC / QB , where Qc and QB are the cylinder heat loss
and
the actual heat release calculated from the indicator diagram,
respectively.
The cooling loss ratio and the combustion efficiency, ηu , were estimated
using the pressure data and respective heat release in a cycle, Q , as well
as the heating value of the fuel supplied in the cycle, Q fuel , as follows:
30
Q  QC  = QB  QC  ηu = η 1  φ 
Q
= B
u
w
Q fuel
Q fuel
QB
(2.2)
Therefore, Q / Q fuel corresponds to a function of combustion efficiency
ηu and cooling loss ratio φw that can be evaluated with indicator diagrams
for quasi-constant combustion efficiency ηu .
2.5.1 Comparison of hydrogen versus methane combustion
Using the above defined variables and theoretical approach, it was possible
to compare in what respects the hydrogen and methane combustion
processes differ. Figure 2.14 shows the effect of ignition timing on incylinder pressure, rate of heat release (ROHR), cumulative heat release, incylinder gas temperature and cylinder wall temperature (under constant
engine operating conditions).
(a) Hydrogen
(b) Methane
Figure 2.14: Influence of ignition timing on hydrogen and methane
combustion (Shudo et al., 1999).
31
Figure 2.15: Influence of ignition timing on components of thermal
efficiency (Shudo et al., 1999).
Figure 2.15 summarises graphically the effect of ignition timing on the
various components of thermal efficiency. From Figure 2.15 it can be
concluded that the degree of constant volume combustion, η glh , and
cooling loss to the combustion chamber walls control the thermal efficiency
in hydrogen stoichiometric combustion, and that compared with methane
combustion, hydrogen has a higher amount of cooling losses for any ignition
timing. This is though to be due to a thinner temperature boundary layer
because of the shorter quenching distance of hydrogen, but also increased
forced convection due to higher combustion velocity may promote the heat
transfer to the combustion chamber walls. However, in particular for rich
mixtures and high pressures, the influence of the higher values of specific
heats for hydrogen (at 200 C and 1 bar, hydrogen specific heats are: Cp =
14.32 kJ/kgK and Cv = 10.16 kJ/kgK, whereas those for methane are: Cp =
2.22 kJ/kgK and Cv = 1.70 kJ/kgK) compared with those of methane should
32
be the main reason of the differences in cooling losses between these two
gases.
Figure 2.16 shows a comparison between stoichiometric and lean
combustion with an excess air ratio of λ = 1.5 . The results show slower
combustion due to decreased combustion velocity, therefore decreasing
the gas temperature. Also, the combustion chamber wall temperatures are
decreased and the reduction in thermal losses increase the engine thermal
efficiency. As a result, the cooling loss, defined as ηu 1  φw  , decreases
with the increase in excess air ratio therefore being a possible and
plausible reason for the high thermal efficiencies of the hydrogen fuelled
engines.
Figure 2.16: Influence of excess air ratio on the components of thermal
efficiency (Shudo et al., 1999).
Based on this research work, Shudo et al. concluded that the increase of
excess air is an effective way of increasing the engine thermal efficiency by
decreasing the heat losses, and that hydrogen has a higher amount of
cooling losses than methane. Also, it was found that there is a relationship
between the degree of constant volume combustion and thermal losses, the
two factors that dominate the thermal efficiency.
33
2.5.2 Heat transfer in hydrogen fuelled engines
As a consequence of the identified heat loss characteristics of hydrogen
fuelled engines, Shudo et al., (2002) concluded that current heat transfer
models developed for hydrocarbon fuel operated engines do not apply
directly to hydrogen fuelled engines. Comparing the heat losses predicted
using a number of existing models applicable to hydrocarbon fuelled
engines with the measured heat losses from an experimental engine, Shudo
et al. concluded that those models underestimate heat transfer losses, and
that the use of correction coefficients doesn’t accurately define the actual
heat transfer rate. A new heat transfer model for hydrogen fuelled engines
had to be derived to better predict the real process. According to the
authors, the existing models, based on turbulent heat transfer in tubes and
correlating mean cylinder gas temperatures and mean in-cylinder pressure
to determine average heat transfer coefficients, give rise to errors in the
determination of heat losses in the order of 4 compared with experimental
data. As a consequence, the following equation for the calculation of heat
transfer to the combustion chamber wall as a function of crank angle was
proposed:
dQw
1
= Sα Tg  Tw 
6n
dθ
(2.3)
Here, S is the total surface area of the combustion chamber (m2), n is the
engine speed (rpm), α is the heat transfer coefficient (W/m2/K) and Tg
and Tw are the mean in-cylinder gas temperature and the mean combustion
chamber wall temperature, respectively.
The authors used standard techniques when calculating the rate of
apparent heat release from the test data. However, an important finding
was the variation in the cylinder charge specific heat has high influence in
hydrogen engines. For hydrocarbon combustion, the change in specific heat
ratio γ over the cycle is relatively small and can therefore be neglected,
this cannot be done for hydrogen engines. When subjected to in-cylinder
pressures and temperatures, a hydrogen-air mixture experiences high
variations of C p and Cv , therefore also giving large variations of the
34
specific heat ratio γ . For richer cylinder charge, this effect becomes
increasingly important.
Hydrogen heat losses depend on the amount of constant pressure
combustion in the cycle and can analysed using the methodology outlined
by Shudo et al. (1999). Shudo et al., (2002) used two fuels, methane and
hydrogen, to establish a comparison for the same engine under identical
operating conditions, and the differences are presented in Figure 2.17.
Figure 2.17: Comparison of hydrogen and methane combustion.
Figure 2.17 shows a comparison of stoichiometric combustion of hydrogen
and methane under various ignition timing conditions. It was found that the
heat release of hydrogen combustion is completed in a shorter period of
time than that of methane combustion because of the higher burning
velocity. A negative apparent heat release after the end of combustion is
significant in hydrogen combustion. Changes in the temperature of the
combustion chamber wall during the combustion period are also significant.
These results suggest higher heat transfer losses in hydrogen combustion.
Having demonstrated the limitations of existing heat transfer models,
including well-established models such as those of Van Tyen, Nusselt,
Eichelberg, Woschni and Briling, Shudo et al., (2002) presented a new heat
35
transfer equation applicable to hydrogen fuelled engines, taking into
account factors such as the higher flame speed of hydrogen and the
differences in combustion chamber heat transfer when compared with
hydrocarbon fuels. It was suggested that the heat transfer coefficient, α ,
should be calculated as:
α = C1 D 0.2 P 0.8Tg0.53 w 0.8
w = Cm + C2
dQ Tr
dt PrVr 
(2.4)
(2.5)
Cm is the mean piston speed and dQ / dt is evaluated as defined above. The
coefficients C1 and C2 in the new heat transfer equation for hydrogen
combustion with different ignition timings and excess air ratios are shown
in Figure 2.18. It can be seen by inspection of the figure that the influence
of the excess air ratio on the coefficients is larger than that of the ignition
timing. Both coefficients increase with a decrease in the excess air ratio,
hence for high excess air ratios the heat losses through the combustion
chamber walls is decreased. In future research, expressing the coefficients
C1 and C2 as a function of the various wide operation conditions and fuel
properties could increase the universality of the proposed heat transfer
equations.
Figure 2.18: Coefficients in the new proposed heat transfer equations for
different operating conditions (Shudo et al., 2002).
36
2.5.3 Exhaust heat losses in hydrogen fuelled engines
Following their line of research on hydrogen fuelled engines and the
associated heat losses, Shudo et al., (2007) carried out a number of tests
with direct injection of hydrogen, i.e. stratified charge, in order to reduce
heat losses and improve engine efficiency. A reduction in cooling losses
does not always translate directly into an improvement in indicated
thermal efficiency, however it can do so if simultaneous exhaust losses do
not increase as well.
Based on the previously presented cooling loss ratio theory, the indicated
thermal efficiency can be defined as
ηi = ηth η glh ηu 1  φw 
(2.6)
By defining the cooling loss fraction φ w as the fraction of the cumulative
cooling loss heat QC to the cumulative real heat release QB , the apparent
heat release fraction Q / Q fuel corresponds to a function of the combustion
efficiency and the cooling loss fraction ηu 1  φw  (Shudo et al., 1999).
Therefore:
Q / Q fuel = ηu 1  φu 
(2.7)
This can be used to express the cooling losses as a function of the heat
supplied and the combustion efficiency:
φw = 1  Q / ηu Q fuel 
(2.8)
Similarly, the exhaust gas loss ratio can be defined as follows: Exhaust heat
loss Qex , is the heat carried away by the exhaust gases per cycle, which can
be defined as a function of the heat supplied per cycle and the indicated
work Wi produced per cycle as:
Qex = Q  Wi
where Q
(2.9)
is the cumulative apparent heat release calculated from the
measured pressure data and the amount of heat supplied per cycle, Q fuel .
The cumulative heat release is
Q = ηu 1  φw Q fuel
(2.10)
From the definition of indicated thermal efficiency, the indicated work can
be defined as
37
Wi = ηth ηglh ηu 1  φw Q fuel = ηth ηglh Q
(2.11)
Therefore the exhaust heat loss, Qex , can be described as
Qex = ( 1  ηth η glh )ηu 1  φw Q fuel
(2.12)
Combining these, an exhaust loss fraction φex , i.e. the fraction of exhaust
gas heat Qex to the supplied fuel heat Q fuel , can be defined as follows:
φex = Qex / Q fuel
(2.13)
The exhaust loss fraction can then be expressed as a function of the heat
loss fraction as follows:
φex = 1  ηth ηglh ηu 1  φw 
(2.14)
From the above equations it is expected that there will be an optimum
point of operation, at which the losses of heat through the exhaust gases
and the heat losses to the combustion chamber walls give the highest
possible engine efficiency. According to Shudo et al., (2007), this depends
on the degree of constant volume combustion ηglh , therefore a method that
doesn’t reduce ηglh can be the stratified direct injection of hydrogen.
When the degree of constant volume combustion is lower than a critical
value (depending on the actual engine characteristics), the exhaust loss
fraction increases and leads to a decrease in the theoretical thermal
efficiency. Therefore, engines with lower compression ratios require higher
degrees of constant volume combustion to improve the thermal efficiency,
due to the cooling loss reduction. The increase in the apparent heat release
fraction effectively leads to improvements in thermal efficiency, because
the stratified charge reduces the cooling loss without lowering the degree
of constant volume combustion. The authors reported that for one engine
with a compression ratio of 14 and a degree of constant volume combustion
of 0.95, a 15 % cooling loss fraction achieves a very high value of indicated
thermal efficiency, over 50%. Therefore, according to the authors, direct
injection stratified charge is an effective technique to improve the thermal
efficiency of hydrogen combustion engines.
38
2.6 Hydrogen engine safety
Hydrogen fuelled engines require particular safety measures in the design
of the engine, auxiliary equipment and operating environment. These
safety measures can be grouped into three main groups:
1.
Prevention of explosive atmosphere.
2.
Removal of ignition sources outside the combustion chamber.
3.
Protection against explosion in engine components.
The measures required for hydrogen fuelled engine operation can typically
include:

installation of a hydrogen leakage sensor on the supply lines;

sufficient ventilation of the engine room (at least 20 air
volume renewals per hour);

monitoring of the engine room atmosphere for hydrogen
leakage;

crank case ventilation with fresh air, or connection to the
engine air inlet;

glow plugs fitted at the exhaust pipe flange to ensure that
any hydrogen is combusted at the engine outlet; and

dilution of the exhaust gas on the stack by installing a forced
draft ventilator to ensure that LEL (Lower Explosion Limit)
concentrations are never reached.
Avoidance of ignition outside the combustion chamber is achieved by using
equipment designed for explosive atmospheres, for example with ATEX
certification. Protection against explosion is achieved by installing rupture
discs in the exhaust system, to ensure that the pressure in the exhaust
system does not rise above its design limit. Safety measures for hydrogen
engine operation will be discussed in more detail in Chapter 3.
39
2.7 Conclusions
From the limited research work available on HCCI and DIH2 engines, the
large potential of these modes of operation is clear, in particular when a
comparison of thermal efficiencies with other fuels and modes of operation
is made. However, significant research and development efforts in a
number of areas related to engine operation, safety and fuel supply and
storage are required before hydrogen fuelled engines will present a
realistic commercial alternative to conventional engines.
An important conclusion that can be drawn from this research review is
that the very distinct physical properties of hydrogen make the use of
existing modelling approaches developed for hydrocarbon fuel combustion
questionable, and care should be taken if such models are used. Possibly,
new approaches need to be developed, tested and validated, in particular
for compression ignition hydrogen engines. Improving the understanding of
hydrogen engine combustion and operation is fundamental for achieving
high thermal efficiencies.
The design of HCCI or DIH2 engines can in principle be derived from existing
commercial models, however development work in the control of the rate
of pressure rise for both modes of operation and on the control of the
ignition timing for the HCCI mode will be required. One option to develop
this technology and allow it to mature before the use of hydrogen as fuel is
to operate such engines with natural gas, which has more favourable
properties with respect to storage, availability and safety.
High performance HCCI and DIH2 hydrogen engines will be suitable for use
in a range of applications. In addition to automotive engines, which much
of the recent research focuses on, such engines should be suitable for
marine and rail propulsion, as well as stationary power generation such as
combined heat and power systems. Considering the challenges associated
40
with storage and transport of hydrogen, stationary systems are perhaps the
most realistic application for hydrogen engines in the near future.
41
42
Chapter 3
Engine performance analysis through
experimentation
« There is no such thing as a failed experiment, only experiments with
unexpected outcomes. »
Buckminster Fuller
This chapter presents the engine and monitoring system which was
designed and constructed to conduct experiments on HCCI and DIH2
hydrogen fuelled operation. Both hydrogen injection system designs and
their characterization, control and engine monitoring are discussed.
3.1
Engine experimental setup
The compression ignition engine test rig was designed to allow the
experimentation of various modes of operation covered by this research,
namely, homogeneous charge compression ignition and direct injection of
hydrogen operating with a compression ignition engine.
3.1.1 Compression ignition engine
With the objective of adhering as much as possible to an implementation of
the most common industrial type of engine, the main experimental part of
the research work was carried out using a production four stroke, single
cylinder, direct injection, naturally aspirated, air cooled compression
ignition engine. The engine is illustrated in Figure 3.1.
43
Figure 3.1: Cross section of the test diesel engine. (In this case the twocylinder version of the Deutz F1L 511).
Engine components illustrated in Figure 3.1 include:
1: Air cooling fan
13: Oil level plug
2: Fuel injector pipes
14: Crankcase
3: Inlet valve
15: Crank
4: Exhaust valve
16: Counterweight
5: Cylinder head
17: Lub oil suction pipe
6: Rocker arm
18: Connecting rod
7: Finned cylinder liners
19: Lub oil pump
8: Piston
20:Rack lever
10: Push rod
21: Main distribution pulley
11: Camshaft
22: Driving belts
12: Flywheel cover
23: Fuel injection pump
44
The specification of the engine used for experimentation was:
Manufacturer:
Deutz
Model:
F1L511
Bore
100 mm
Stroke
105 mm
Swept volume
825 cm3
Compression ratio
17:1
Maximum engine speed
3000 rpm
Intake air valve opens
32º BTDC
Intake air valve closes
59º ATDC
Intake air valve open duration
91º
Intake air valve closed duration
269º
Exhaust valve opens
71º BTDC
Exhaust valve closes
32º ATDC
Injection angle (diesel oil)
24º BTDC
Exhaust valve open duration
103º
Exhaust valve closed duration
257º
Valve overlap duration
27º
The engine was directly coupled to a constant displacement hydraulic pump
via two flexible couplings. The engine load using the hydraulic pump was
changed through a restriction valve, by varying the discharge pressure. The
hydraulic system arrangement of the test rig allowed a stable load to be
applied under all the engine operating conditions. The hydraulic pump
compression line was instrumented with a pressure sensor which, in
conjunction with the speed signal from an encoder fitted to the engine
camshaft end, allowed the calculation of the engine shaft power. This
calculation was performed based on the pump displacement, hydraulic
pressure and pump efficiency. The hydraulic pump efficiency was
45
considered over the entire operating range and incorporated into the power
calculation.
Figure 3.2 shows the engine test rig and the hydraulic brake specification is
given in Table 3.1.
6
1
2
7
8
3
4
5
9
1 - H2 flow meter; 2 - diesel fuel burette; 3 - Air inlet manifold; 4 – Engine;
5 - Data acquisition system cables; 6 - 200 bar H2 cylinder; 7 - Exhaust pipe;
8 - Hydraulic oil tank, brake control panel and pressure transmitter; 9 Hydraulic brake pump
Figure 3.2: Engine test rig.
46
Table 3.1: Specification of the hydraulic engine pump.
Hydraulic pump characteristics:
Hydraulic pump maker:
Voith
Hydraulic pump model:
IPH 3-10
Pump displacement:
5.2cm3/turn
Maximum speed:
3000 RPM
Maximum discharge pressure:
300 bar
Efficiency (100 – 200 bar range): 0.85 – 0.90
Hydraulic tank capacity:
200 dm3
3.1.2 Air supply system modifications
The engine air supply system, as shown in Figure 3.3, was modified to
accommodate a 3.5 kW air heater, capable of raising the air inlet
temperature up to approximately 120ºC, with an ambient temperature of
10ºC. Also, an air mass flow meter was installed in the air inlet manifold
and the air filter was removed, in order to reduce the pressure loss across
supply system.
A i r he a ter
r esis tance
Air mass
flowmeter
Figure 3.3: Inlet manifold heater and mass flow meter.
3.1.3 Fuel system
The diesel fuel system, as shown in Figure 3.4, allowed diesel consumption
measurement through an electric pulse type flow meter installed in the
diesel fuel pipe. This signal was acquired by a data acquisition system. As
an additional measurement of the fuel flow rate, a 1.5 dm3 graduated
47
burette, placed 1.5 m above the injection pump was used and consumption
measured manually.
D ie se l
Flow meter
D ie se l F i l t e r
Figure 3.4: Diesel oil flow meter.
Hydrogen was supplied to the engine, from an 80 dm3 hydrogen cylinder at
200 bar. The engine hydrogen fuel system comprised two different
arrangements, depending on the injection pressure in use. For hydrogen
low pressure injection (< 8 bar) the system was operated with two pressure
regulators, one that reduced the pressure from 200 bar to 15 bar, in series
with a second regulator with a fine regulating adjustment valve and
respective pressure indicator allowing a constant pressure regulation
between 15 bar to 0.3 bar. For hydrogen high pressure injection (>120 bar)
the system was operated with only one pressure regulator, capable of
supply constant pressure between 45 bar and 200 bar. A needle isolating
valve and a flame trap were placed before the hydraulic injector in both
supply systems.
3.1.4 Exhaust gas system
In order to allow the engine exhaust gas to be analysed, a ball valve was
fitted after the silencer, to allow the exhaust gas analyser probe to be
inserted inside the exhaust pipe.
48
3.1.5 Test rig instrumentation and data acquisition system
To collect the data from engine test rig, a high speed data acquisition
system was developed based on National Instruments hardware. Software
was also developed to control the engine using a purpose developed
injector controller. The instrumentation used on the engine test rig, for
DIH2 and HCCI operation is presented in Figures 3.5 and 3.6. All the sensors
shown were interfaced with the data acquisition system described below.
Exhaust
Exhaust Gas
Analyzer
+
FID
T U
m
Inlet Air
Heater
2800w
U
P U
T U

U
P U
Air filter
T U
Diesel
Engine
m
P U
U

N
U
U
M
H2
200
bar
Figure 3.5: Hydrogen direct injection engine test rig schematic diagram.
49
Hydraulic Oil
Tank
Figure 3.6: Hydrogen Homogenous Charge Compression Ignition engine test
rig schematic diagram.
3.1.6 Data acquisition system hardware
A real time data acquisition system, based on LabView software and
National Instruments hardware was developed for data logging and
monitoring. The engine speed and crank angle was measured using an
absolute encoder directly fitted on the engine camshaft, allowing the
cylinder pressure data to be referenced to the piston top dead centre
(TDC). The sampling speed and resolution of the system was 100 kHz and
16-bit respectively.
The data acquisition system was made up of five National Instruments
boards and its block diagram is presented in Figure 3.7. One NI PCI-7830R (a
communication board using the digital input output (DIO) connection with
40 I/O digital lines); two cRIO-9423, which has 8 input voltage channels up
to 30 volt, one cRIO-9211 (four thermocouple analogue input board); CB68-
50
LP (a terminal block board which allowed four analogue inputs, four digital
outputs and sixteen additional high speed digital lines TTL) and the
expansion chassis RIO NI cRIO-9151, into which all the boards were
inserted. The system shared the engine crank angle and speed signal from
the encoder.
51
52
Terminal block CB68-LP
AI0
AI1
AI2
AI3
PC
MIO
Cylinder Pressure sensor
Hydraulic oil brake pressure sensor
Air mass flow meter
Hydrogen mass flow meter GFM
SH68-C68-S
Board NI PCI-7830R
DIO
SH68-C68-S
Comptact expansion chassis RIO NI cRIOcRIO-9423
cRIO-9423
cRIO-9211
8 digital
input
24VDC
8 digital
input
24VDC
4 analog
input
Termocoupl
Available
slot
Thermocouple 1
Thermocouple 2
Encoder Signal
To the PIC injection controller
Figure 3.7 Data acquisition block diagram
3.1.7 Data acquisition system software
A high speed data acquisition program was developed to allow the engine
variables acquired, monitored and analysed on and off line. This software
was developed using LabView, and was designed to have four interface
displays: Visualization, Analysis, Configuration and Diagnosis.
The Visualization Display (Figure 3.8) was designed to allow the
simultaneous acquisition and monitoring of the following variables: engine
power (kW); in cylinder pressure (bar); hydraulic load pressure (bar);
hydrogen flow rate (l/min and kg/h); diesel fuel oil flow rate (kg/h); air
mass flow rate (kg/h); equivalence ratio; air intake temperature (ºC);
exhaust gas temperature (ºC); crank angle (º); engine speed (rpm); and
engine thermal efficiency (%).
Figure 3.8: Visualization Display menu.
The Analysis display menu is shown in Figure 3.9 and was designed to allow
the simultaneous analysis of selected variables and provided a series of
analysis tools, such as:
53

Plots with zoom and graph scaling functions, indicating the
variable magnitude against crank angle or time.

Derivatives in the time domain or crank angle domain. This
analysis tool can be used to analyse the signals, for
example rate of change in the cylinder pressure and the
determination of ignition angle.

Fast Fourier Transform (FFT) of any signal. FFT analysis can
be used to determine the existence of resonance
phenomena during the measurement of the in-cylinder
pressure.

Forward 720º and rewind 720º to analyse cycle by cycle
graphically.

Averaging of cycles: the average from a number of cycles
can be derived. This is a basic parameter to overcome the
noise effect present in the signals.

Median filter. As the derivative tends to amplify the signals
noise, a median filter was included to mitigate such effect.
The Analysis display menu allowed the selection of any of the available
variables and also the export and import of data in ”CSV”, “DAT” and
“BMP” formats.
54
m
Figure 3.9: Analysis Display menu.
Figure 3.10 shows an example of a cylinder pressure diagram (against crank
angle) produced by selecting the in-cylinder pressure option in the
software.
Figure 3.10: Cylinder pressure against crank angle analysis display.
55
The Configuration display menu is shown in Figure 3.11 and was designed
to allow calibration of the various interfaced transducers and also the
introduction of all the constants required for the program algorithm
calculations.
Figure 3.11: Configuration display menu.
The Diagnosis display menu is shown in Figure 3.12 and was designed to
allow the setting of sampling rate to be used, memory space management,
and monitoring of the channels data.
56
Figure 3.12: Diagnosis display menu.
3.1.8 Engine speed and crank angle measurement
The engine speed and crank angle measurements were made using an
absolute encoder fitted at the engine camshaft end, as shown in Figure
3.13. This encoder fed its signal to the injection control system and the
high speed data acquisition system through a specially made junction box.
Figure 3.13: Absolute encoder fitted at the camshaft end of the test
engine.
57
3.1.9 Cylinder pressure transducer
A autoPSI-S cylinder pressure transducer was utilised to measure in-cylinder
pressure variation over the engine cycle. The accuracy of this transducer is
within  1.0 % (linearity and hysteresis combined). Special care was taken
in the selection of the sensor in respect of its resonance frequency, since
masking of cylinder pressure transients could occur if the sensor resonance
frequency coincides with the pressure transient frequency.
3.1.10 Air mass flow transducer and measurement
Air mass flow to the engine was measured using a special hot film mass
flow meter (Bosch HFM5), shown in Figure 3.14. By processing the sensor
data it was possible to detect when return flow takes place during air flow
pulsation. The air mass flow was compensated in terms of density and
temperature and had an extensive measuring range.
Figure 3.14: Air mass flow meter Bosch HFM5.
The transducer had a measuring range of 8 to 800 kg/h, accuracy better
than 3 %, a fast response of 15 ms, and was able to withstand vibration
accelerations up to 150 m/s2.
58
3.1.11 Hydrogen mass flow measurement
Hydrogen mass flow was measured using a Dwyer GFM-1107 mass flow
meter with a totalizer. It was based on a straight tube sensor with a
restrictor flow element to provide high accuracy (±1.5% of full scale) and
repeatability (± 0.5% of full scale).
3.1.12 Lambda oxygen transducer and measurement
A lambda oxygen sensor (Bosch LSM 11) was installed just after the exhaust
port, to monitor λ , the O2 concentration in the exhaust gas from the
engine.
The response time for lean mixtures was approximately 2 s and the relative
sensitivity ΔU s / Δλ at a -value of 1.3 was 0.65 mV / 0.01, where the ΔU s
is the sensor supply voltage.
3.2 Hydrogen fuel injection systems
As described above, hydrogen injection systems for both HCCI and DIH2
operation was developed for this project. This section describes design
considerations and the solutions chosen in this work.
3.2.1 Material considerations when using hydrogen
A difficult problem with the use of hydrogen is associated with a form of
corrosion caused by the depletion of oxygen from protective metal oxide
coatings. Without the protecting oxide film most metals are highly reactive
and corrosion results. Electrons pass through the metal to areas where
oxygen availability is greatest and this leaves metal ions in the exposed
area free to enter into solution if an electrolyte is present. The electrolyte
may be formed by a film of water that is acidic due to dissolved hydrogen.
Damage to stainless steel in strained areas where sufficient oxygen is
available is characterised by the following stress corrosion process. The
59
reaction provides hydrogen ions to make the water electrolytic and to
enter the metal lattice:
4Fe++ + O2+10H2O -> 4Fe(OH)3+8H+
Stainless steel sealing surfaces are often designed to be deformed by
mechanical stress until a line of contact produces the desired sealing
effect. Near the line of contact capillaries are formed which trap moisture.
As a result of this mechanism stainless steel seals near the line of contact
and under gaskets experience deterioration in their physical properties.
Discolouration and pitting may result from depletion of the protective
chromium oxide coatings. Dark red, brown, or black stains reveal the
migration of iron from the sealing surfaces where the chromium oxide is
depleted or damaged and cannot be repaired because of oxygen deficiency.
Exposure to air during inspection causes these iron ions to rapidly oxidise.
Leaks develop because corrosive movement of metal ions leaves pathways
through seal surfaces through which hydrogen can pass. G.Santhana et al
1988.
Rapid embrittlement of steel and other metals, such as nickel and copper
nickel alloys, will occur at ambient temperatures when exposed to
hydrogen gas at high pressure. The embrittlement is produced by the
effects of solution and diffusion of hydrogen on the crystal structures of
these metals. At temperatures above approximately 300ºC, in addition to
embrittlement, hydrogen attack occurs where it reacts with the carbon in
the steel to produce gaseous methane:
Fe3C + 4H -> CH4 + 3Fe
Methane molecules are much larger than hydrogen and carbon in solid
solution and produces internal pressure. Accumulation of methane
molecules greatly increases internal stress. The disappearance of the
carbides and the formation of methane can seriously weaken the alloy in
question. However, alloy steels containing such elements as chromium,
tungsten,
titanium,
and
vanadium,
which
form
chemically
stable
intermetallic carbides with the carbon within the steel, are resistant to
60
high-temperature hydrogen attack and allow them to retain their strength
up to about 400ºC. G.Santhana et al 1988.
3.2.2 Hydrogen HCCI injection system
To operate the engine in HCCI mode, hydrogen is injected in the proximity
of the air intake manifold, through a fast injection solenoid valve
controlled by a microprocessor. The frequency of injection was set by the
engine speed, whilst injection duration was determined by a pulse width
modulation signal (PWM) control. Figure 3.15 shows the injection valve
fitted on the engine air inlet manifold.
Digital
pressure
gauge
Injection
valve
Hydrogen
supply
Figure 3.15: Fast-acting solenoid injection valve and hydrogen pressure
gauge, fitted on the engine
The injector system for the HCCI mode of operation was relatively simple
as the pressure and temperature of operation was low and the time
available for injection was over a relatively large crank angle. A pulse
width modulated (PWM), two way, normally closed solenoid activated ball
valve injector was used in this mode of operation. The valve is shown in
Figure 3.16 and a simplified cross section is presented in Figure 3.17.
Designed with hydrogen embrittlement resistant materials, this was used to
inject hydrogen directly into the engine intake air manifold in a controlled
61
manner. For adequate control of hydrogen flow rate, the pulse width
modulation was used to meter fuel quantity at injection frequencies up to
200Hz and with timing precision of  25 microseconds. Under these
conditions, sonic flow across the valve inlet occurs when fully open and
mass flow is approximately proportional to the supply pressure (Barkhimer
et al., 1995).
Figure 3.16: HCCI Hydrogen injection Figure 3.17: Simplified diagram of the
solenoid activated ball valve.
HCCI hydrogen injection solenoid
activated ball valve.
With the solenoid de-energised, the supply pressure, assisted by a spring,
forces the solenoid ball poppet on its seat, barring gas flow. When the
solenoid is energised, the ball poppet is lifted off the seat and held against
the stop. Gas then passes through the valve seat and outlet port of the
injector. The solenoid has a low impedance coil designed for fast response.
It is typically actuated by a current of 4 amperes, which is reduced and
held at 1 ampere to conserve energy for the duration of the energised time
(pulse width) (Barkhimer et al., 1995).
3.2.3 Hydrogen direct injection system
There are no hydrogen injectors available on the market, therefore it was
necessary to design, construct and calibrate a DIH2 injector operating at
62
high pressure. Flow, dynamic response and leak tests were performed to
characterise the injector.
The DIH2 fuel system comprised a high pressure hydrogen cylinder (200
bar), a regulator to reduce the hydrogen pressure before the hydraulic
injector (from 200 bar to 90 bar), a flame trap valve, and a solenoid valve
controlled hydraulic injector controller. (See Figure 3.5 for a schematic
diagram of the setup.) The frequency of injection was set by the engine
speed whilst the injection duration was determined by the engine load. The
hydraulic actuation oil supply was provided by a hydraulic pack which was
continuously flushed with nitrogen to avoid an explosive atmosphere inside
the hydraulic oil tank in case any hydrogen leakage occurred.
The design of the DIH2 injector was carried out using Solid Works, and the
injector cross section is shown in Figure 3.18. Figure 3.19 shows the
manufactured solenoid controlled DIH2 injector and Figure 3.20 the
installation of the injector in the engine.
Figure 3.19: Solenoid
Figure 3.18: Solenoid
controlled hydraulic DIH2
controlled hydraulic DIH2 cross
injector.
section.
63
H yd r a u l ic
actuator oil
s up p l y
H2
s up p l y
C yl i n d e r
Head
D IH 2
Injector
Figure 3.20: Solenoid controlled hydraulic DIH2 injector installation.
Important design considerations for the hydrogen injector include
clearances and tolerances, weight, rubbing and lubrication of the moving
parts, and material selection for resistance to hydrogen embrittlement.
Clearances are critical to eliminate hydrogen leakage into the hydraulic
actuation oil system or the atmosphere. The injector developed and used in
the experimental work exhibited a considerable amount of leakage of
hydrogen into the hydraulic actuation system, caused by excessive
clearances found between the needle valve and the injector body. The
clearance was due to inappropriate tolerances in the machining process,
and this needs to be rectified in case of a non-experimental injector
production.
The injector body was manufactured from steel, and the needle valve and
nozzle were manufactured from highly tempered steel. Since hydrogen has
a “washing effect” over the lubricating oil, lowering the oil viscosity and
therefore the oil adherence to the moving parts, the injector was designed
with a Teflon liner to accommodate the actuator rod, which is solidly
attached to the needle valve. Hydrogen embrittlement was observed
mainly on the needle valve tip, since it operates at around 300ºC and the
hydrogen flow speed is high. Nitrile O-rings were used to seal injector
components as shown in Figure 3.18. No detrimental effects of using
64
hydrogen was found, however it is known that cracking can occur when in
contact with certain rubber materials. Therefore, the use of special O-ring
materials such as Viton may need to be considered.
The dynamic performance of the injector was determined by the mass of
the moving parts, the pre-load of the needle valve spring, and the
hydraulic actuation pressure. The spring stiffness and hydraulic pressure
could be adjusted to optimise the dynamic response. Lighter materials can
also be employed to reduce the inertia of the moving parts and therefore
providing a better injector dynamic response. For example, Titanium could
be used for the actuator and needle valve which were found not to
experience embrittlement problems.
3.2.4 Injector hydraulic power pack
The DIH2 injector was designed to be hydraulically actuated and
electronically controlled using a solenoid valve. Therefore, a hydraulic
power pack was used to generate the hydraulic power required by the
injector actuator. Figures 3.21 and 3.22 show a picture of the hydraulic
power pack and injector and a schematic diagram of the system
respectively. The hydraulic power pack has the capability of regulating the
actuation delivery pressure and also the flow rate through the injector. The
power pack system consists of an electro-hydraulic pump, a pressure safety
valve, a pressure control valve, a flow control valve, a thermo-static valve
and a plate heat exchanger.
65
Figure 3.21: Hydraulic power pack and DIH2 injector.
3 to
Purge line to
atmosphere
Hydraulic oil IN
Hydrogen pressure
transducer
0.1MPa
0.3M
Hydrogen 100 Bars
Hydraulic oil return +
Hydrogen
N2
Hydraulic oil tank
200
bar
N2 inert space
Figure 3.22: DIH2 hydraulic actuating and inert gas (nitrogen) purging
systems
A fast actuating hydraulic three way valve was incorporated to control the
direction and duration of the hydraulic oil pulses, based on a PWM signal
from the injection controller. Once the solenoid valve allows high pressure
66
oil to enter the actuator lower chamber, an upward force is exerted on the
injector needle valve, causing the injector to open and allowing hydrogen
to flow into the engine cylinder. With the removal of the PWM signal pulse,
the oil in the lower chamber is forced back through the control valve into
the hydraulic oil tank. This forces the piston and needle valve downwards,
closing the injector again.
3.2.5 Alternative DIH2 injector design
A second injection system was designed and developed, but not
implemented as it was not sufficiently flexible for experimental purposes.
The injection system, illustrated in Figures 3.23 and 3.24, provides rapid
hydrogen injection without requiring any electronic control system. The
system utilises the camshaft actuated diesel fuel injection pump to
pressurise the working fluid, a blend of 50% diesel oil and 50% lubricating
oil (10W-30). It therefore eliminates the complexity associated with the
external hydraulic actuation system. The stroke of the injection pump
would set the duration of injection, and the timing of the injection would
be set by steel shims placed under the injection pump body after the cam
rise. At the end of the injection pump stroke, the compressed spring would
provide the energy required to force the injector needle on to its seat, as
the working fluid flows through the bleed valve.
67
Figure 3.23: Hydraulically controlled and actuated hydrogen injector.
The duration of hydrogen injection would be determined by the rack
setting of the injection pump, which is controlled by the existing engine
governor. By adjusting the bleed valve, the ideal compromise between the
opening and closing response time can be achieved.
68
injection
p u mp
s h i ms
b l ee d
va l ve
H2
1 20
b ar
c a msh af t
5 0 %d i e s e l
5 0 % l u b o il
t ank
Figure 3.24: Schematic diagram of the hydraulically controlled and
actuated injection system
As mentioned this injector system was built but not fully implemented for
hydrogen injection on the engine. Although this system is simple and
functional, the control of the start of injection (an important engine
control parameter) was not possible as it was set by the shims height.
Therefore a solenoid controlled injector was developed to overcome this
problem. Therefore the solenoid controlled injector with external hydraulic
actuation, which allowed the use of electronic controller circuit to change
the start of injection and duration as a function of engine load, was used
69
for the experiments. For a commercial engine, an injection system such as
this may however be investigated further.
3.3 Hydrogen injector test rig
It was necessary to characterise the hydrogen injection system with respect
to its flow rate and the PWM control settings, so that the correct amount of
fuel is injected per engine cycle. Therefore, a test rig was designed and
built to carry out the static and dynamic characterisation process for both
the HCCI and DIH2 injectors.
3.3.1 Pressure vessel for injector testing
Due to the hydrogen pressures involved, it was decided to construct a
pressure vessel to test the injectors, shown in Figures 3.25 and 3.26.
I n jec t or h yd r au l ic
so le no id e va lve
In jec tor tie b ar
T her moc oup le
H 2 Injec tor
D ig i ta l
pressur e
g age
De press uriz ing
valve
Known
volume
p resur e
vess el
Figure 3.25: Injector test vessel.
70
Supply Line Pressure
Transducer
H yd r o g en
s up p l y
To purge line
Injector
PWM
generator
1. 937 x
-3
3
10 (m )
H 2 F l ow m e t e r
Digital
pressure gage
Pr ess ur e
ve s s e l
Thermocouple
Double Trace Digital
Storage Osciloscope
Figure 3.26: Injector testing rig.
The injector test vessel was designed to be capable of supporting a gas
pressure up to 90 bar, and equipped with a temperature measuring system
(a K type thermocouple), a digital pressure gauge, an injector pocket and
a needle valve. According with the perfect gas law, the mass of hydrogen
contained within a volume is given by the following expression:
71
m=
p× v
R× T ,
(3.1)
Therefore, knowing the pressure and the temperature inside the vessel, it
was possible to determine the mass of hydrogen injected into the vessel
during testing. Knowing the frequency of injection, pulse width, and the
time of observation, it is possible to determine the quantity of hydrogen
injected per injection at constant injection frequency and pulse width.
Great care was required to make sure all measurements were taken at a
stabilised temperature, with the added difficulty of reducing the heat
losses through the vessel walls as close as possible to zero.
In addition to this method, the hydrogen injected and contained inside the
vessel was released through a hydrogen flow meter, which produced
another reading of the hydrogen quantity contained in the vessel. In this
way two independent methods of measuring the mass of injected hydrogen
were used to characterise the injector being tested. Figure 3.27 shows the
hydraulic injector testing apparatus; the oscilloscope was used to set the
frequency of injection and the power supply was used to feed the PWM
electronic circuit.
72
H2
P r ess ur e Ve ss
D IH 2 In je ct o r
DIH 2 injector
D ig ita l Pr es su re
Purge
D IH 2
T he r mco up l
H 2 l in e
Pow er
O sc i lo sc op
PWM Circuit
B rea d Bo ar d
Figure 3.27: Photograph of the DIH2 injector under test.
3.3.2 Pulse width modulation control circuit
To generate a pulse width modulated signal to control the injector solenoid
valve, a circuit was developed based on the oscillator LM 555 and the
comparator LM 393. The control circuit interfaced with the power circuit
(12 V battery, 55 Ah) through a Zener diode and an opto-isolated Silicon
73
Controlled Rectifier (SCR). The Zener diode was used to limit and therefore
to protect the circuit from a current surge developed by the 2 ohm
impedance coil of the solenoid valve when it is activated.
The circuit, shown in Figure 3.28, operates at a frequency determined by
R1, R2 and C1 and has a pulse width range of 0 to 100 percent.
Figure 3.28: Pulse width modulation control circuit.
Figure 3.29 shows the pulse width modulation output of the basic circuit for
a given control voltage input. All measurements were made with a
calibrated multimeter. Figure 3.30 shows a modified circuit which uses a
second LM 555 oscillator timer to provide a power output stage for the
basic PWM circuit. The PWM circuit was constructed using a “bread board”
as shown in Figure 3.31.
74
Figure 3.29: Basic pulse width modulation control circuit characteristic
curve.
Figure 3.30: Variable pulse width modulation control circuit.
75
Figure 3.31: Variable pulse width modulation control circuit bread board.
3.3.3 Static and dynamic characterisation of HCCI and DIH2 injectors
The injectors are used to inject the right quantity of hydrogen at the right
time. To perform these tasks, the injectors need to be characterised in
terms of their static and dynamic performance, which depend on variables
such as supply pressure, back pressure, frequency, and injection frequency.
The static performance test objective was to determine the injector full
open flow rate. The high speed at which the injector operates, and the
small flow rate for each opening, meant that it was impossible to measure
the flow rate per injection directly. Therefore, to determine the mass flow
rate of hydrogen per injection the following experimentation procedure
was implemented.
Once the injector was installed in the pressure vessel, sealed with copper
washers, and hydrogen, hydraulic oil, and electric connections done, a
PWM injection frequency equal to 16.66 Hz (corresponding to an engine
speed of 2000 RPM) and a series of duty cycles from 6.67% to 26.71% were
set on the PWM generator.
76
After a time interval of injection into the closed pressure vessel, the
hydrogen was slowly released and measured using the calibrated flow
meter. In parallel, using the thermocouple and pressure gage readings, a
calculation of the hydrogen injected into the cylinder was carried out using
the ideal gas law. However, this gave results which deviated when the
pressure inside the vessel was above 20 bar and therefore it could not be
used to check the DIH2 injector. This deviation was observed using thermal
isolation of the pressure vessel, as the vessel increased its temperature
during injection and decreased its temperature during expansion. Another
factor contributing to this deviation was the ambient temperature of 11ºC
at the time of the tests. Even though this calculation of the injected flow
produced inaccurate results at higher pressures, it was valuable to check
the calibration of the flow meter for low pressures. Also, it was possible to
conclude that the use of the ideal gas law it is not sufficiently accurate
when temperature can not be maintained constant.
The hydrogen injection test procedure was repeated four times for each
selected pulse width for both injectors (DIH2 and HCCI injectors).
3.3.3.1 Static performance test results for the HCCI injector
The static performance test results for the HCCI injector using the test rig
are presented in Tables 3.2 to 3.5. These results are plotted in Figure 3.32
and show the hydrogen flow rate as a function of pressure and pulse width.
Knowing the test duration and the set injection frequency, the number of
injections was determined. The total mass flow rate measured was divided
by the number of injections to obtain the mass of hydrogen per injection.
77
Table 3.2: HCCI Injector flow rate (mg/injection) data for an average
supply pressure of 2.47 bar.
Pulse
Test
Test
Test
Test
AVG
ST
width (ms)
1
2
3
4
Q
Dev.
4
1.56
1.55
1.54
1.54
1.55
0.0094
6
2.21
2.21
2.20
2.20
2.21
0.0031
8
2.79
2.79
2.78
2.79
2.79
0.0058
10
3.36
3.37
3.36
3.37
3.37
0.0084
12
4.01
4.01
4.02
4.01
4.01
0.0083
14
4.62
4.63
4.63
4.63
4.62
0.0061
16
5.24
5.25
5.26
5.26
5.25
0.0080
Table 3.3: HCCI Injector flow rate (mg/injection) data for an average
supply pressure of 7.58 bar.
Pulse
Test
Test
Test
Test
AVG
ST
width (ms)
1
2
3
4
Q
Dev.
4
2.91
2.91
2.89
2.88
2.90
0.0152
6
3.50
3.52
3.53
3.51
3.52
0.0100
8
4.68
4.66
4.68
4.67
4.67
0.0076
10
5.77
5.78
5.78
5.78
5.78
0.0041
12
6.85
6.86
6.86
6.86
6.86
0.0041
14
7.95
7.99
7.99
7.99
7.98
0.0227
16
9.12
9.12
9.14
9.14
9.13
0.0127
78
Table 3.4: HCCI Injector flow rate (mg/injection) data for an average
supply pressure of 10.96 bar.
Pulse
Test
Test
Test
Test
AVG
ST
width (ms)
1
2
3
4
Q
Dev.
4
3.46
3.45
3.46
3.48
3.46
0.0137
6
4.69
4.70
4.76
4.70
4.71
0.0314
8
6.20
6.23
6.26
6.26
6.24
0.0248
10
7.81
7.83
7.84
7.84
7.83
0.0157
12
9.47
9.54
9.53
9.53
9.51
0.0396
14
11.02 11.06 11.09 11.07 11.06 0.0296
16
12.78 12.79 12.81 12.83 12.80 0.0216
Table 3.5: HCCI Injector flow rate (mg/injection) data for an average
supply pressure of 14.5 bar.
Pulse
Test
Test
Test
Test
AVG
ST
width (ms)
1
2
3
4
Q
Dev.
4
4.49
4.45
4.44
4.48
4.46
0.0247
6
5.88
5.83
5.79
5.83
5.83
0.0383
8
7.74
7.76
7.76
7.75
7.75
0.0094
10
9.88
9.87
9.89
9.87
9.88
0.0120
12
12.14 12.13 12.16 12.13 12.14 0.0145
14
14.33 14.34 14.36 14.34 14.34 0.0127
16
16.46 16.47 16.44 16.46 16.46 0.0108
79
18
16
14
Flow Rate (mg/inject.)
12
P = 4,27 bar
10
P = 7,584 bar
P = 10,962 bar
8
P = 10,548 bar
6
4
2
0
2
4
6
8
10
12
14
16
18
PWM (4-16 ms)
Figure 3.32: Flow rate (mg/injection) as a function of the PWM and supply
pressure for the HCCI injector.
As can be seen from Figure 3.32, above an 8 ms pulse the flow rate with
respect to PWM is approximately linear. Below an 8 ms pulse, the control
characteristic of the valve flow becomes non-linear. The results presented
in Figure 3.32 provide the characterisation of the HCCI injector and allowed
the initial setting up of the injection system on the engine to be carried
out. The small standard deviation observed for each pulse width sets of
tests was indicative of a good metering behaviour of the injector, therefore
allowing a precise control of the fuel quantity injected per cycle. It was
possible to conclude that the injector dynamics is playing an important role
in terms of its response for smaller pulse widths, therefore limiting the
engine operation at very low loads.
3.3.3.2 Static performance tests of the DIH 2 injector
The static performance test results for the DIH2 injector using the test rig
are presented in Tables 3.6 to 3.8. These results are plotted in Figure 3.33
and show the hydrogen flow rate as a function of average supply pressure
and pulse width.
80
Table 3.6: DIH2 injector flow rate data for an average supply pressure of
60 bar.
Pulse
test
Test
Test
Test
ST
width (ms)
1
2
3
4
AVG Q
Dev.
4
3,99
3,97
3,97
3,94
3,97
0,0233
6
5,07
5,03
5,03
5,03
5,04
0,0189
8
6,81
6,8
6,8
6,79
6,8
0,0098
10
8,44
8,44
8,44
8,42
8,42
0,0066
12
10,2 10,21 10,21 10,19
10,2
0,0089
14
12,15 12,2 12,19 12,22 12,19
0,0288
16
13,82 13,9 13,93 13,96
0,0606
13,9
Table 3.7: DIH2 injection flow rate data for an average supply pressure of
70 bar.
Pulse
Test
Test
Test
Test
ST
width (ms)
1
2
3
4
AVG Q
Dev,
4
4,1
4,04
3,99
3,98
4,02
0,0549
6
5,75
5,78
5,77
5,77
5,77
0,0138
8
7,5
7,44
7,42
7,42
7,45
0,0366
10
9,41
9,41
9,4
9,38
9,4
0,0129
12
11,72 11,72 11,71 11,71 11,71
0,0054
14
13,8 13,79 13,76 13,77 13,78
0,0197
16
15,77 15,68 15,55 15,96 15,74
0,1748
81
Table 3.8: DIH2 injection flow rate data for an average supply pressure of
80 bar.
Pulse
Test
Test
Test
Test
ST
width (ms)
1
2
3
4
AVG Q
Dev,
4
4,28
4,31
4,32
4,34
4,31
0,0219
6
5,9
5,91
5,89
5,88
5,89
0,0092
8
8,26
8,28
8,28
8,22
8,26
0,027
10
10,35 10,34 10,39 10,35 10,36
0,0202
12
12,99 12,97 12,96 12,93 12,96
0,0224
14
15,31 15,28 15,23 15,23 15,27
0,0396
16
17,6 17,62 17,61 17,61 17,61
0,0104
Hydrogen Injector valve for DI operation mode
20
18
16
Flow rate (mg/injection)
14
12
P = 60 bar
P = 70 bar
P = 80 bar
10
8
6
4
2
0
2
4
6
8
10
12
14
16
18
Pulse With (ms)
Figure 3.33: Flow rate (mg/injection) as a function of the PWM and average
supply pressure for the DIH2 injection.
As can be seen from Figure 3.33, for the DIH2 injector above a 10 ms pulse
the flow rate with respect to PWM is approximately linear. Below a 10 ms
pulse, the control characteristic of the valve flow becomes non-linear.
It was noticed that the DIH2 injector had substantial hydrogen leakage into
the hydraulic oil return pipe. However, since the flow meter was used to
82
measure the amount of hydrogen injected into the vessel and then allowed
to flow out, this corresponded to the actual injected mass flow of hydrogen
which would be injected into the engine cylinder. Also, the injector
preload, the hydraulic pressure and PWM setting determine the injector
performance. It was possible to conclude that the injector dynamics is
playing an important role in terms of its response for smaller pulse widths,
therefore limiting the engine operation at very low loads.
3.3.3.3 HCCI dynamic injector response
For the HCCI injector, the time delay between the injector receiving the
opening signal and commencing fuel injection is mainly a function of the
solenoid time response and this introduced a 3.0 ms delay when opening
and a 2.0 ms delay when closing. It was assumed that all other factors
contributing to the dynamic response of the HCCI injector were lumped
together and taken in consideration on the considered delays, therefore no
additional dynamic testing was performed.
3.3.3.4 DIH2 dynamic injector response
The speed of response is an extremely important parameter for the DIH2
injector and the operation delay must be included in the programming of
the injection controller. As DIH2 engines control is dependent on the
injection timing, and hydrogen ignition delay is short, the ignition angle
and combustion control require an accurate injection. If the delays are not
taken into account in the controlled operation then a long delay in opening
time would lead to late injection, and hydrogen being wasted if the
injector closing has a large delay. The ideal injector should have a fast
dynamic response giving the smallest possible time delay to open and close.
This inherent delay must be compensated for in the controller.
It was found using a dynamic simulation model of the injector that opening
and closing times were not significantly determined by the inertia of its
moving parts but dominated by the hydraulic and injector spring preload
83
forces. Therefore, the injector dynamic response was predominately found
to be a function of the actuating piston effective area, hydraulic pressure,
and the spring elastic constant and preload. An additional factor which
adversely affected the injector dynamic response was the frictional force
introduced by the O-rings used to seal the piston within the hydraulic
actuator cylinder. The effect of this and the other variables mentioned
which determine the injector dynamic performance was studied in more
detail using dynamic simulation and this is presented in Chapter 5. The
dynamic model was used to simulate the engine test conditions
contributing to know in advance the values of hydraulic oil pressure and
piston pre load, but also to understand the effect of increasing the
frequency of injection and the respective duty cycle over the pattern of
injection. To determine the injector time delay in opening and closing
experimentally, a pressure transducer (dynamic response type) was
installed on the hydrogen supply line. The hydrogen pressure signal was fed
into a dual beam oscilloscope, together with the PWM signal. In this way,
the time difference between the edge of each PWM signal and the distinct
change in hydrogen pressure was used to determine the opening and closing
time delay. Both these time delay values were used in the injection
controller to ensure correct injection timing. The time delay test results
from the oscilloscope are presented in Figure 3.34 and show that there was
a 21 ms delay in opening and a 16 ms delay in closing for the DIH2 injector.
84
12
120
10
100
8
80
6
60
4
40
Secondary injection
2
Supply Pressure Signal
Solenoid voltage (V)
PWM signal (12V DC)
Supply pressure
20
0
0
0.05
0.1
0.15
0.2
0.25
0.3
0.35
0
0.4
Time (s)
Figure 3.34: Oscilloscope traces showing time delay measurements for the
DIH2 injector. Supply pressure signal (0-100%).
(Delay: 21 ms to open 16 ms to close)
3.4
Hydrogen injection engine control system
The hydrogen injection engine control system was designed to operate with
a high and low level control loop. The low level hydrogen injection control
system was used to monitor the crank angle of the engine via the camshaft
encoder and provide the control signal to actuate either the HCCI or the
DIH2 fuel injector.
The high level control loop consisted of software which ran on a PC
communicating with the low level control loop via the PC serial port. The
high level control loop provided a user interface to program parameters
such as the desired start of injection and its duration.
85
3.4.1 Low level hydrogen injection control loop
The main component of the low level hydrogen injection control loop was a
Microchip PIC16F677A 8-bit microcontroller which has a built-in serial port
and analogue to digital converter. A custom-designed printed circuit board
for this microcontroller and supporting components is shown Figure 3.35.
This circuit board was mounted in a rugged plastic enclosure to provide the
low level hydrogen injection control loop hardware shown in Figure 3.36.
Figure 3.35: Low level hydrogen control loop circuit board.
Figure 3.36: Low level hydrogen injector control system hardware.
The microcontroller firmware was largely interrupt-driven, and the main
programming loop monitored the camshaft encoder output and activated
the injector control signal at the appropriate time. The activation of the
injector started a timer and the injector was closed once the timer
expired.
86
All communication between the low level hydrogen injection control loop
hardware and the PC was via a microcontroller USART module, set up as an
asynchronous bi-directional serial port with a baud rate of 19.2 kbps. A
MAX232CPE RS232 voltage level converter device was used to convert the
signal voltage from the 5V used by the microcontroller to the ±12V required
by the host PC.
The main microcontroller program structure is shown in Figure 3.37. The
programme was designed to run in a loop, continually checking the value of
the camshaft encoder and comparing it with the encoder value at the start
of injection. When a match has been detected, the “Trigger Injector”
subroutine is called and the duration of injection period is programmed
into the TMR1 timer. The timer is then activated and the processor returns
to the main program structure, running the “Check Encoder” loop. The
TMR1 timer runs in parallel with the main program structure and triggers an
interrupt when it reaches zero.
Main Program
Re ad Sh a ft Enc oder
Do es enc od er va lu e =
s tar t o f injec tio n?
T rigge r In jec tor :
Yes
O pen in jector
T urn in dic ator L ED on
S t ar t dura t io n ti mer a t
“ In j ec tion Du ra tion” , co un ting
d ow n .
En ab le T MR1 o ver flow
interrupt
No
Figure 3.37: Main microcontroller program structure.
87
The interrupt service routine programming is shown in Figure 3.38 and
when called it determines the cause of the interrupt. If TMR1 caused the
interrupt, the “Close Injector” routine is called, switching off the injector
and the indicator LED, resetting the timer and disabling TMR1 as a source
of further interrupts. If a “serial port receive” event triggered the
interrupt, the received data is analysed and then the required action is
performed. This may be to enable or disable the injector in order to start
or stop the engine, or to trigger the “Log Data” process to record analogue
sensor values from the engine. By controlling the injector in this way, the
open duration of the injector was entirely independent of the engine
rotational speed, allowing more flexible control of the engine.
88
Interrupt Service Routine
T MR1 O verflow
In terru p t?
Yes
Ca l l “C lose In jec tor”
S u brou tin e
No
Exit
S e r ia l Po r t R e c e i ve
Interrupt?
No
Exit
Yes
Ca l l “ Ser ia l R X”
Su brou tin e:
If R X da ta = “ S” , u pdate
i n jec t ion c o n tr o l
p ara me te rs w i th n ew
va lu es an d se nd s ta tus
r epor t.
I f R X d a t a = “ P” , s e n d
s ta t us r ep or t b ac k to th e
PC.
I f R X d a t a = “ L ” , tri g g e r
“ Log D a ta” su bro u tine . T his
logs the data on the
a na log ue ch ann els and
se nds it back to the PC .
I f R X d a t a = “ E” , tr i g g e r
“R ead Enco der” s ubro u tine
t o r ea d a va lu e fro m th e
sh a ft enc od er . T his is us ed
d u r in g t h e i n i ti a l s e tup o f
th e s ys tem.
I f R X d a t a = “ X” , tr i g g e r
“R un /Stop ” su bro u tine to
s tar t or s top the e ng ine .
Otherw ise, reply with error
message.
Exit
Figure 3.38: Interrupt service routine program.
89
The microcontroller has several embedded hardware functions which
enable interaction with other systems and devices. The functions used in
this application are shown in Table 3.9.
3.4.2 High level hydrogen injection control loop
The high level hydrogen injection control loop software was written in
Matlab and ran on a PC to send and receive control data from low level
hydrogen injector control loop via the PC’s serial port.
The Matlab software allowed the timing and duration of fuel injection to be
set and allowed analogue engine sensor data to be sampled. This data was
plotted within the PC user interface and could be saved in a variety of
formats, including MS Excel and Matlab .mat files.
The user interface for the high level hydrogen injection control loop is
shown in Figure 3.39. The “Injector Control” box (top left) was used to
specify the crank angle at which the injector should open and the time it
should remain open in milliseconds. Clicking the “Set” button sent these
settings to the microcontroller which then used these values every cycle
until instructed otherwise.
The top-right box contains the data logging and saving controls. Figure 3.39
shows the result of clicking the “Log Data” button with the data shown in
the plot window logged by the low level hydrogen injection control loop
and then transmitted back to the PC. Clicking the “Save Data” button saved
this data with the file name and path specified in the boxes below the
buttons. The data shown in the plot window in Figure 3.39 was generated
by a signal generator connected to the analogue input port of the low level
hydrogen injection control loop to test its operation.
90
Table 3.9 Low level hydrogen injector control loop microcontroller hardware functions.
Hardware
Configuration Setting
Connection
Asynchronous bi-directional serial port.
Connected to PC
Allows communication between PC
Baud Rate 19.2kbps, 8 data bits, 1 stop
serial port via
and hardware module.
bit, no parity, no handshaking.
MAX232CPE
Purpose
Function
USART
voltage level
converter.
A/D
Voltage reference on Pin 5. High-speed
5V Zener diode
Allows data collection from engine
voltage
sensors. High-speed channel is
with 8-bit resolution. Low speed data
reference on Pin
intended for use with cylinder
acquisition on Channel 1, Channel 2 and
5. All data
pressure sensor to record pressure
Channel 4, with 10-bit resolution.
channels
profile over full cycle. Low speed
These channel numbers are those on the
connected to
channels can be used to collect
microcontroller only. The channels have
screw terminals
temperature data, for example.
been renumbered on the assembled
on enclosure.
Converter data acquisition on Channel 0 (Pin 2)
module to make connecting inputs more
intuitive.
91
Figure 3.39: High level hydrogen injection control loop user interface.
To set up the camshaft encoder, the engine was held at top dead centre at
the end of the exhaust stroke. The encoder value was then read and the
resulting value stored as TDC. All injection timing was then calculated from
this datum value.
The software included the facility to scale the analogue data before saving
or displaying it, so that the data could be plotted as actual values, e.g. 0100 bar, rather than a simple voltage value.
The analogue inputs from the engine sensors were required to be in the
range of 0-5 V sharing a common ground connection. The low level
microcontroller could log a maximum of 256 data points with single-byte
(8-bit) precision at a maximum of 20 kHz. This allowed a cylinder pressure
trace, for example, to be taken over a whole cycle, sampling every 2.8
crank angle degrees.
92
It was found that to fully analyse the engine cylinder pressure data to
determine performance through identifying parameters such as maximum
cylinder pressure, crank angle at maximum pressure, rate of pressure rise,
pressure at ignition point and ignition angle, it was necessary to have a
sampling resolution of 1 crank angle degree or less. This was important also
to provide data with sufficient accuracy to validate simulation models used
to explore a wider operational range than that possible through
experimentation. Therefore, it was decided not to use the data sampling
facility provided with the developed high level hydrogen injection control
loop but to implement an additional high speed data acquisition loop and
analysis system. This is shown in schematic form in Figure 3.40 which
indicates the use of the separate data acquisition and analysis system
which was used to sample engine sensor data with 0.5 crank angle degree
resolution.
93
PWM
controlled
Injector
PWM
Signal
Encoder
Signal
Junction
Box
Low Level
Hydrogen Injector
Control System
Injection
Controller
Test
Engine
Encoder
DAQ 12
Channel
16 bit
100 kHz
Engine
Sensors
Signals
High Level
Hydrogen Injector
Control System
For Data Logging
& Analysis
For Setting the
Injection
Parameters
Figure 3.40: Integration of high and low level hydrogen injector
control loop with the data acquisition system.
94
3.5 Conclusion
This chapter has described the engine and monitoring system which was
designed and constructed to conduct experiments on HCCI and DIH2
hydrogen fuelled operation. The engine test rig comprised a single cylinder
compression ignition engine capable of running on these hydrogen
operational modes. The instrumentation and acquisition of engine sensor
data was an important element of the test rig setup so that accurate data
could provide a detailed understanding of the engine performance during
these modes. A high speed system was developed, capable of monitoring all
relevant engine variables at 0.5 crank angle degree resolution and thereby
enabling detailed analysis to be carried out.
Two main hydrogen fuel injection systems were designed and implemented
with the engine test rig. These systems were prototypes for research
purposes and would require further re-engineering for example in
component tolerances and material selection before they would be suitable
for commercial application. A HCCI injection system operating at a
hydrogen pressure of 6 bar was designed and implemented using a fast
acting solenoid injection valve in the engine air inlet manifold. Two DIH2
injectors were designed, one which was a solenoid controlled hydraulic
system and the other a hydraulically controlled and actuated system. The
solenoid controlled hydraulic injector was found to be the most flexible
design and therefore was developed further and used on the engine test
rig.
This
DIH2
injector
had
significant
leakage
problems
due
to
manufacturing limitations, but this did not impact on the characterisation
of the injector and the determination of the quantity of fuel injected into
the engine. However, it prevented the originally planned wider range of
operational testing to be conducted because of potential safety concerns.
It was important to characterise the properties of and amount of hydrogen
delivered by HCCI and DIH2 injection systems and it was therefore
necessary to design and build an injector test rig and to conduct the
characterisation process. The static and dynamic delivery performance of
both injection systems were measured using this test rig, which at the start
95
of hydrogen injection had atmospheric pressure and temperature. The
results produced were invaluable in the initial setting up of the injection
control system on the engine since it is critical to know the relationship
between fuel flow and pulse width. Since HCCI injection takes place within
the air inlet manifold the conditions within the injector test rig were close
to those found in engine operation. However, the injector test rig did not
emulate in-cylinder conditions in terms of pressure and temperature for
DIH2 injection, since this would take place within the engine cylinder at
approximately 27 bar and 260oC. It was still found to be useful in order to
get an initial insight into the characteristics of the DIH2 injector to aid the
setup and engine calibration.
Finally, a hydrogen injection control system capable of being implemented
with both HCCI and DIH2 injection systems was designed and implemented.
It was found that the use of an 8-bit microcontroller within the low level
control module was sufficient for engine injection control requirements.
However, a higher resolution of data was necessary to monitoring engine
operation and to retrieve data of use for detailed analysis.
96
Chapter 4
Experimental testing of hydrogen engine
operation: results and analysis
“Think wrongly if you please, but in all cases think for yourself“
Doris Lessing
This chapter presents the methodologies, procedures and results obtained
from the testing carried out with the hydrogen fuelled engine test rig under
HCCI and DIH2 modes of operation. Experimental results of dual fuel
operation (hydrogen + diesel oil) are also presented as this mode of
operation has a great potential for immediate use in the industry, namely
on board ships.
4.1 Objectives of engine testing and methodology
The objectives of the test were to:
(a) prove that hydrogen can be used as a fuel for compression ignition
engines operating under HCCI and DIH2 modes;
(b) ascertain how the hydrogen impacted on the engine combustion in
terms of cycle pressure development and thus on the component
design; understand the limitations to the use of hydrogen use as a
fuel in commercial CI engines;
(c) compare the HCCI and DIH2 modes of operation in terms of thermal
efficiency; and
(d) collect data to validate the simulation model developed to study
operational conditions that were not practically feasible with the
test engine.
4.1.1 Testing procedures
The hydrogen fuelled engine was tested using the engine test rig described
in Chapter 3. Taking advantage of the installed instrumentation and high
97
speed data acquisition system it was possible to analyse test data in real
time.
For each mode of operation the engine was tested systematically at
different loads at constant speed, allowing engine cycle analysis.
A second set of tests were performed at constant load whilst varying the
inlet air temperature, thus investigating the effect of the air temperature
on the engine performance.
In order to have a reference set of values to compare with the HCCI and
DIH2 modes of operation tested, the engine was first run on diesel fuel
using its conventional injection system. The DIH2 mode that was tested only
at one load due to safety problems related to the injector hydrogen
leakage.
4.1.2 Engine operation and safety
For the testing, the engine was started with diesel fuel and run until the
engine components were sufficiently hot. For switching to the HCCI mode
of operation, hydrogen port injection was initiated and the diesel injection
jerk pump rack was simply pulled out, thereby reducing diesel oil injection
to zero. For the DIH2 operation mode, the engine had to be stopped and the
diesel injector was exchanged rapidly for the hydrogen injector.
Since the test engine had a modest compression ratio (17:1), and because
of the high self-ignition temperature of hydrogen, pre-heating of the intake
air was necessary to obtain stable running. An air heater with a simple PID
controller was installed in the engine inlet manifold, as described in
Chapter 3. This was used to maintain the air inlet temperature at
approximately 90ºC for the hydrogen fuelled tests.
In this way, sufficiently high temperatures were reached inside the engine
cylinder for the air-hydrogen cylinder charge to self-ignite.
4.1.3 Instrumentation set up and operation
Prior to testing, all the monitoring equipment was calibrated and checked
to avoid erroneous readings. The data logging and monitoring software was
98
checked and the values of fuel densities and net calorific values were set.
Special care had to be taken with the brake hydraulic oil temperature, as
continuous testing over long periods produced an increase of the hydraulic
oil temperature affecting the oil viscosity and therefore causing an engine
load drift. Therefore all the measurements were taken within a time frame
sufficiently short to overcome this undesirable effect.
4.2 Data logging and treatment
The raw data from the various transducers installed on the engine was
logged in a .DAT file format. This allowed the data of each cycle to be
available for further analysis after testing. The analysis of the logged raw
data was performed using the software described in Chapter 3 in analysis
mode, so that the raw data could be treated with the various analysis
tools. Filters, averaging, FFT, and zoom features were available in the
software. After the identification of the sets of data for further analysis,
these sets were exported as .CSV files. The common averaging of data gave
results that allow a reduction of noise as well as a representative data of
the cylinder pressure history. The online data acquisition software included
the capability of producing this averaging on a crank angle or time basis,
with up to fourteen averages.
4.2.1 Cylinder pressure measurement data
There are a number of factors that can affect the quality of the acquired
cylinder pressure data. Resonances, hysteresis, frequency response and
thermal stability are some of those factors.
The cylinder pressure sensor was connected to the combustion chamber by
a small channel of 2 mm diameter and 60 mm length. The resonance
frequency of gas through the channel was determined in order to exclude
such frequencies from the measured data. The resonance frequency of the
gas in the channel is inversely proportional to its length and the order of
the harmonic and is given by:
υ=
2K  1C
(Hz)
4L
99
(4.1)
C is the speed of sound of hydrogen (468 m/s to 520 m/s);
K is the harmonic index (1, 2, …)
L is the length of the channel.
Therefore the 1st harmonic the resonance frequency the first harmonic
would be 7.8 kHz. A FFT analysis of the pressure trace was made to verify if
a harmonic component with this frequency was present which could mask
the pressure trace signal. No resonance frequencies were identified and the
pressure transducer diaphragm resonant frequency was 120 kHz, which was
well above the resulting resonance frequency of the gas pressure waves
through the channel. Also the frequency response of the pressure
transducer covers a range between 0.1 Hz and 25 KHz, allowing the
discrimination of phenomena with a wide frequency range.
Investigation of the other sources of error was not performed as they were
stated by the pressure transducer manufacturer.
4.2.2 Cylinder pressure sampling rate
There was the need to determine the maximum frequency of interest, fm,
because the signal was to be recorded at discrete values of time. Therefore
there is the potential for the generation of false readings, or aliases.
The sampling rate theorem states that does not state that to avoid aliasing
and provide an accurate representation of the original waveform, the
sampling rate, fs, must be greater than twice the maximum frequency for
the signal, fm:
fs > 2 x f m
(4.2)
If the sampling rate restriction is met, then the original wave form can be
recovered using the series (Marks, 1991):
f (t ) 
1

sin  t / T  n 
 f nT  t / T  n

n  
100
,
(4.3)
f (t ) is the reconstructed function, f nT  is the discretely Sampled
values of the function, N is an integer corresponding to each sample and
ΔT is the sampling period (1/fs).
One important characteristic of this equation is that it assumes an infinite
set of sampled data and is hence an infinite series, whereas the actual sets
of sampled data are finite. However, because the set sampling rate is high
(the data points per sample at 2500 rpm were in excess of 3000) the series
is convergent, allowing the reconstruction of the original signal with a
finite number of samples.
In any case, since the data acquisition system was able to sample data up
to 100 kS/s, to make sure that no aliasing phenomena was present the
sampling frequency was set to 50kS/s. Once the function was decomposed
by FFT analysis into its harmonics values, it was reconstructed again to
check if all the values lay on top of the data points that constitute the
original function. If a point calculated by the reconstructed function does
not coincide with the original data, then an harmonic would be producing
some distortion. Using the above procedure, it was checked that no aliasing
phenomena or resonance was present, and therefore that no data
distortion occurred.
An FFT of the pressure–volume sampled data was performed to determine
the maximum frequency of interest for the signal. With 2000Hz, this was
well below the resonance frequency of the pressure transducer.
As discussed by Den Hartog (1956), Churchill (1987) and Kamen (1987), any
periodic function f(t) can be represented by the sum of a series of sine and
cosine waves multiplied by a constant value which, in symbolic form, this
can be written as:
f (t )  a1 sin  0 t  a 2 sin 2 0 t  ...  a n sin n 0 t  b0  b1 cos  0 t  b2 cos 2 0 t  ...  bn cos n 0 t
(4.4)
101
T
2
an   f (t ) sin( n0t )dt
T 0
(4.5)
T
bn 
2
f (t ) cos(n0t )dt
T 0
(4.6)
and T is the period, f(t) is the function of time, ω0 is the angular velocity,
t is time and n = 0, 1, 2, …
∞
Since the pressure versus volume function is odd, it can be represented
entirely with a series of sine terms, which is the Fourier sine series.
Having done the FFT analysis for the pressure versus volume function which
the frequencies and respective amplitudes are listed on table 4.1, and
substituting the amplitudes and angular frequencies, and then substituting
various values of time, it was found that the values encountered by this
substitution fall on top of the pressure versus volume function. Such a
result is indicative of good quality data which has not been distorted by the
aliasing phenomena.
Table 4.1: Cylinder pressure trace main harmonic components their
frequencies and amplitudes.
Harmonic Frequency (Hz)
Harmonic Amplitude (bar)
6.1035
7.0740
24,4141
6.9689
42,7246
5.7053
54.9316
4.2352
73.3420
3.9820
91,5527
3.1997
109.8633
2.2380
122.0703
1.9773
102
4.3 Methodology of engine testing
The in-cylinder pressure variation was examined to compare the
combustion process for the hydrogen fuelled HCCI and DIH2 modes with the
conventional diesel fuelled engine operation. Figure 4.2 shows full cycle
cylinder pressure plots for the three operational modes with the engine at
the same speed and load. Comparing the pressure traces, a significant
difference in the combustion process between the three modes of
operation can be seen.
75
2200 rpm 5.0kW Die sel
70
In-Cylinder. Pressure (bar)
65
60
55
50
45
40
35
30
25
20
15
10
5
0
0
50
100 150 200 250 300 350 400 450 500 550 600 650 700 750
Crank Angle (º)
Diesel oil operation
110
2200 rpm, 5.0kW, H2
100
In-Cylinder. Pressure (bar)
90
80
70
60
50
40
30
20
10
0
0
50
100 150 200 250 300 350 400 450 500 550 600 650 700 750
Crank Angle (º)
HCCI hydrogen operation
55
50
45
In-Cylinder Pressure(bar)
40
35
30
25
20
15
10
5
0
0
50
100
150
200
250
300
350
400
450
500
550
600
650
700
750
Crank Angle (º)
DIH2 hydrogen operation
Figure 4.1: In-cylinder pressure traces for Diesel, HCCI and DIH2 operation
at 2200 rpm and 5 kW.
103
The implemented engine testing methodology was the same for all the
operation modes (Diesel, HCCI and DIH2), to allow an easier comparison of
the results. The engine speed was set to be constant and equal to 2200 rpm
for the majority of the testing, however some other speeds were tested to
determine the effect of engine speed on the engine performance. While
maintaining a constant speed, the engine load was varied and other
operational variables such as maximum combustion pressure, angle of
maximum combustion pressure, and exhaust gas temperature were
recorded, and thermal efficiency and IMEP calculated.
4.4 Diesel fuel operation characterisation
The engine was tested in conventional Diesel fuel mode of operation and
variables such as the magnitude of the maximum combustion pressure,
angle of maximum pressure, IMEP, exhaust gas temperature and rate of
pressure rise were recorded as a reference condition. It was decided that
the control of such variables while testing the engine operating under HCCI
and DIH2 mode should be observed in order to maintain the engine
component´s mechanical and thermal integrity.
Table 4.2: Maximum observed values of TEXH , PMAX and PIGN at (a) 5800 W,
(b) 3380 W, (c) 1770 W engine load. (2200 RPM, Ta= 21ºC.)
Brake Load (W)
1770
3380
5080
5800
TEXH (ºC)
318
335
363
372
PMAX (bar)
47.5
55.5
68.5
76.0
PIGN (bar)
46.0
46.5
47.8
48.8
From Figure 4.3 and Table 4.2 it is possible to observe an increase of the
ignition pressure of the cylinder charge with the increase in load. As
expected, maximum combustion pressure and exhaust gas temperature
increase with load. Figure 4.4 shows the engine thermal efficiency as a
function of load, showing a maximum value of approximately 26% at 5800
104
W. This value served as a reference of comparison with the other modes of
operation at the same load.
75
70
65
In-Cylinder Pressure (bar)
60
55
50
45
40
35
30
25
20
15
10
5
0
8678,39
8678,4
8678,41
8678,42
8678,43
8678,44
6382,785
6382,795
5185,5
5185,51
Time (ms)
65
60
In-Cylinder Pressure (bar)
55
50
45
40
35
30
25
20
15
10
5
0
6382,745
6382,755
6382,765
6382,775
Time (ms)
60
55
In-Cylinder Pressure (bar)
50
45
40
35
30
25
20
15
10
5
0
5185,46
5185,47
5185,48
5185,49
Time (ms)
Figure 4.2: Cylinder pressure diagram for diesel operation at (a) 5800 W,
(b) 3380 W, (c) 1770 W engine load. 2200 RPM, Ta= 21ºC.
105
80
380
75
360
70
340
65
320
60
300
55
280
50
260
1500
2000
2500
3000
3500
4000
4500
5000
5500
Ignition pressure (bar); Max. Combustion Pressure (bar)
Exhaust gas temperature (ºC)
400
45
6000
Engine break load (Watts)
Figure 4.3: Exhaust gas temperature, ignition pressure and maximum
combustion pressure at (a) 5800W, (b) 5080W, (c) 3380W, (d) 1770W.
p
y
2,1
28,5
2,05
27
2
25,5
24
1,9
22,5
1,85
21
1,8
19,5
1,75
18
1,7
16,5
1,65
15
1,6
13,5
1,55
12
1,5
10,5
1,45
9
1,4
Brake thermal efficiency (%)
Diesel Fuel Consumption (Litres/hour)
1,95
7,5
1,35
6
1,3
4,5
1,25
3
1,2
1,5
1,15
0
0
500 1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000
Engine Brake load (Watt)
Figure 4.4: Diesel fuel consumption and brake thermal efficiency as a
function of engine load. (2200 RPM, Ta= 21ºC.)
106
In-Cylinder. Pressure bar. Rate of pressure rise bar/º
diesel oil operation
75
70
65
60
55
50
45
40
35
30
25
20
15
10
5
0
-5
-10
0
50
100 150 200 250 300 350 400 450 500 550 600 650 700 750
Crank Angle (degrees)
Figure 4.5: Open pressure diagram and its derivative of test engine
operated in diesel mode at 5.8 kW load.
Figure 4.5 shows the pressure plot for diesel operation, as well as the rate
of pressure rise. The maximum rate of pressure rise is 8.59 bar/ºCA and the
development of combustion from ignition to its maximum pressure takes
approximately 6.4ºCA.
4.5 Dual fuel operation characterization
Having the test rig with the hydrogen port injection system, testing of the
engine in dual fuel (hydrogen and diesel) mode is possible. A series of such
tests were carried out to generate experience with the test rig and as part
of the development of commercial dual fuel engine technology, described
further in Appendices A and B. Although dual fuel operation is not the main
focus of this thesis, this section will present a summary of the results
obtained during these initial tests.
Dual fuel engines run simultaneously on two different fuels, typically one
gaseous and one liquid, where the liquid is used as a source of ignition for
the pre-mixed cylinder charge. In addition to more flexibility in the fuel
107
supply one can, if the combustion properties of the two fuels complement
each other, obtain a better engine performance than with either of the
fuels alone.
The operation of diesel engines in dual fuel mode with a gaseous fuel
inducted in the intake air is particularly attractive if the gaseous fuel can
improve the diesel combustion, since this may improve on some of the
weak points of this engine type, such as high particulate matter (PM)
emissions. This is particularly attractive if using diesel fuels of low quality
with poor ignition and combustion properties, such as bio-oils or heavy fuel
oils. In the case of biofuels, minimising the need for costly fuel processing
is of great importance, and an engine which can use, for example,
unprocessed vegetable oils directly will provide a substantial advantage in
the overall energy balance.
Diesel engines operating in dual fuel mode with natural gas have been
studied by a number of authors (See Appendix A for a short overview), and
such systems with hydrogen as the inducted fuel have also been reported
(see e.g. Roy et al., 2010; Varde and Frame, 1983; Saravanan et al.,
2007,2008). In general, it is reported that engine fuel efficiency is
comparable to that under normal diesel operation and nitrogen oxides
(NOx) emissions are similar or slightly lower. Substantial advantages have,
however, been seen in the particulate matter emissions, with reductions of
above 50% frequently reported. Lower PM emissions follows naturally from
the substitution of diesel with a pre-mixed fuel (which does not produce
PM), however changes in the in-cylinder processes will also influence this,
potentially further enhancing the PM reductions. Studying the combustion
process and interaction between the two fuels in dual fuel engines is
therefore worthwhile, in order to fully understand the mechanisms
governing the formation of emissions.
4.5.1 Experimental setup
The experimental setup was based around the same system as described
above. The hydrogen injection rate was adjusted to avoid knock,. The
108
fraction of energy supplied by the hydrogen gas is therefore limited by the
knock detection system based on the knock intensity.
4.5.2 Test results
One of the first tasks carried out was the characterisation of the engine
operation for different amounts of hydrogen injection into the inlet air
manifold. The hydrogen flow rate was varied from 1 to 9.8 dm3/min, and
the diesel oil consumption was adjusted automatically as hydrogen flow
rate changed by normal, simple action of the speed governor. This was
repeated for a range of engine loads to provide an extensive data set of
engine performance parameters for varying operating conditions. Table 4.3
shows an example of the percentage of energy based on hydrogen and on
diesel fuel as a function of load.
Table 4.3: Energy share ratios for hydrogen and diesel fuel at different engine
loads for a constant hydrogen flow of 6.0 dm3/min.
Load (%)
Diesel Fuel (%)
Hydrogen energy (%)
0
75.0
25.00
25
84.2
15.01
50
88.0
12.0
75
90.9
9.1
100
92.9
7.1
4.5.2.1 Combustion and energy efficiency
Figure 4.6 shows the brake thermal efficiency values for all the engine test
loads and hydrogen flow rates as well as those under normal diesel
operation. A clear trend towards increased efficiency with increasing
hydrogen flow rate can be seen. This increase in thermal efficiency
indicates a substantial improvement in the combustion process.
109
30
Diesel
2 dm3/min
7.6 dm3/min
9.6 dm3/min
Brake thermal efficiency, %
25
20
15
10
5
0
0
10
20
30
40
50
60
70
80
90
100
Break Load, %
Figure 4.6: Brake thermal efficiency as a function of load for various
hydrogen flow rates.
Maximum Combustion Pressure, bar
85
Diesel
2 dm3/min
7.6 dm3/min
9.6 dm3/min
80
75
70
65
60
55
50
0
10
20
30
40
50
60
70
80
90
100
Break Load, %
Figure 4.7: Maximum combustion pressure for different hydrogen flow rates
compared with diesel-only operation.
In Figure 4.7 the variation of maximum in-cylinder gas pressure at varying
loads and for different hydrogen flow rates is presented. Under stable dual
fuel operation, i.e. without knocking, the peak gas pressure differs only
very little from that under normal diesel operation. This indicates that the
dual fuel engine will have similar performance as a conventional diesel
engine in terms of noise, and no mechanical challenges are expected
(which can be the case in relation to e.g. bearing loads when using very
fast-burning fuels creating high rates of pressure rise).
110
Figure 4.8 shows the variation of exhaust gas temperature for different
hydrogen flow rates and for different brake loads. It was found that the
operation
with
hydrogen
results
in
slightly
higher
exhaust
gas
temperatures, and this temperature increases with the hydrogen flow rate.
Exhaust Temperature, Deg C
500
Diesel
2 dm3/min
7.6 dm3/min
9.6 dm3/min
450
400
350
300
250
200
0
10
20
30
40
50
60
70
80
90
100
Break Load, %
Figure 4.8: Comparison of exhaust gas temperatures between diesel and
various hydrogen flows.
4.5.2.2 Exhaust gas emissions
Although hydrogen is only injected during the induction stroke, i.e. when
the intake valves are open and exhaust valves are closed, some hydrogen
may pass through the engine and into the exhaust. This is known as
hydrogen slip. Figure 4.9 shows the hydrogen slip expressed in ppm at the
exhaust gases. As can be seen from the figure, the level of hydrogen in the
exhaust gases is generally higher than for diesel operation for the cases
with hydrogen injection, however the level of unburnt hydrogen in the
exhaust is in all cases acceptably low. Regarding this effect, the use of port
injection results in a better use of hydrogen energy and also in a system
much safer than hydrogen fumigation system as engine manifold is not full
of an hydrogen air mixture, and therefore the cylinder is not scavenged
with such mixture.
111
60
Diesel
2 dm3/min
7.6 dm3/min
9.6 dm3/min
55
Hydrogen Slip, ppm
50
45
40
35
30
25
20
0
10
20
30
40
50
60
70
80
90
100
Break Load, %
Figure 4.9: Hydrogen slip into the exhaust gases for different hydrogen flow
rates and engine loads.
2500
Diesel
2 dm3/min
7.6 dm3/min
9.6 dm3/min
NOx, ppm
2000
1500
1000
500
0
0
10
20
30
40
50
60
70
80
90
100
Break Load, %
Figure 4.10: Comparision of the effect of hydrogen addition on the NOx
emissions for different engine loads.
Figure 4.10 shows the variation in NOx concentration in the exhaust gases
for various engine loads and hydrogen flow rates compared with those
resulting from normal diesel operation. It can be seen that for up to 50%
load hydrogen injection resulted in a slight reduction of NOx compared with
the emissions resulting from diesel operation. But for engine loads above
112
50%, injection of hydrogen results in a small increase of NOx production.
However, it can be observed that for 100% engine load the NOx emissions in
dual fuel mode are approximately the same as for diesel operation.
3,5
Diesel
2 dm3/min
7.6 dm3/min
9.6 dm3/min
Smoke, Bosch Number
3
2,5
2
1,5
1
0,5
0
0
10
20
30
40
50
60
70
80
90
100
Break Load, %
Figure 4.11: Particulate matter emissions compared for various hydrogem
flowrates and diesel operation.
As has been reported by other researchers, the use of hydrogen has a
strong impact on the formation of particulate matter, and the test results
for the current system are shown in Figure 4.11. The smoke emissions are
found to be lower for any hydrogen flow rate and throughout the tested
load range. It can be seen that the higher the flow rate of hydrogen, the
lower the emissions of particulate matter are, as one would expect. At full
load and for a hydrogen flow rate of 9.6 dm3/min, the measured Bosch
smoke number is found to be 2 for the dual fuel engine, compared with 3.9
for diesel operation. Notably, it can be seen that even a small hydrogen
flow rate (providing in the order of 5% of the fuel energy) leads to a
substantial reduction in particulate matter. This indicates that the
combustion of the diesel fuel is improved by operating the engine in dual
fuel mode, and the higher the hydrogen flow rate the lower the carbon
content of the exhaust gases.
113
4.6 HCCI operation characterization
The HCCI mode of operation was tested at a constant speed of 2200 rpm
and at the following loads: 1600W, 2100W and 4132W, corresponding to
different equivalence ratios. This procedure allowed investigations into the
interdependency between the various variables involved.
During the tests, the HCCI engine operation exhibited very high thermal
efficiency values, approaching 49%, making the initial results very
encouraging. However, this mode of operation produced varying and high
rates of cylinder pressure rise and reduced operational stability, even at
constant load.
4.6.1 Inlet air temperature and ignition control
( p
)
(
)}
60
55
T air inlet = 100ºC
(dp/d)max
50
45
T air inlet = 110ºC
40
35
Tair inlet = 90ºC,
30
2
3
4

5
6
7
8
9
10
Figure 4.12: Dependence of the RPR as a function of Tair inlet and λ.
Figure 4.12 shows the maximum rate of cylinder pressure rise as a function
of λ and inlet air temperature Tair. It can be seen that the maximum rate of
cylinder pressure rise is not significantly affected by the temperature of
the air at the cylinder inlet, but that the excess air ratio λ has a strong
influence. As a reference, the rate of pressure rise in a medium speed
Diesel engine should generally not be higher than 12 bar per crank angle
degree. During the experiments on HCCI operation, rates of pressure rise
114
exceeding 40 bar/º were observed, with the highest values occurring for
richer cylinder charges and lower speeds.
370
2200 rpmy=387.5-0.2071x
 IGN (º CA)
368
366
364
362
90
95
100
105
110
115
120
Ta(ºC)
Figure 4.13: Angle of ignition as a function of air inlet temperature, Tair.
The ignition angle, IGN, is governed by the end of compression
temperature, and is therefore influenced by the temperature of the air
entering the cylinder. Figure 4.13 illustrates this dependence: higher inlet
air temperatures will result in advanced ignition. For high pre-heating
temperatures, the ignition can take place before the piston top dead
centre, reducing the cycle efficiency drastically and increasing the
mechanical loads on the engine bearings due to the fast combustion in the
HCCI engine.
115
4.6.2 Operating characteristics and performance
46
Brake thermal efficency @ 2200 rpm 17:1
Brake Thermal Efficiency (%)
44
42
40
38
36
34
3
3,5
4
4,5

5
5,5
6
Figure 4.14: Brake thermal efficiency at constant speed (2200 rpm) for
varying fuel air ratios.
Figure 4.14 shows the brake thermal efficiency of the engine for varying
excess air ratios. It can be seen that the engine is able to maintain high
fuel efficiency even for very lean cylinder charges.
116
150
140
130
120
In-Cylinder Pressure(bar)
110
100
90
80
70
60
50
40
30
20
10
0
0
50
100
150
200
250
300
350
400
450
500
550
600
650
700
750
600
650
700
750
Crank Angle(º)
4132 Watt HCCI (6 Averages)
80
75
70
65
In-Cylinder pressure(bar)
60
55
50
45
40
35
30
25
20
15
10
5
0
0
50
100
150
200
250
300
350
400
450
500
550
Crank Angle (º)
2100 Watt (6 Averages)
75
70
65
60
In-Cylinder. pressure(bar)
55
50
45
40
35
30
25
20
15
10
5
0
0
50
100
150
200
250
300
350
400
450
500
550
600
650
700
750
Crank angle (º)
1600 Watt HCCI (6 Averages)
Figure 4.15: HCCI open cycle diagrams for different loads.
It can be observed in Figure 4.15, that the cylinder maximum pressure
amplitude it is a function of the engine load, has it was observed that
117
higher combustion pressures are generated at higher loads but in any case
accompanied by high rates of pressure rise, therefore explaining the
difficulties encountered into control the engine speed and its lower
thermal efficiency at lower loads .
8
7,5
7
 Pmax (ºCA); 
6,5
6
5,5
5
4,5
4
3,5
3
85
87,5
90
92,5
95
97,5
100
102,5
105
107,5
110
Ta (ºC)
Figure 4.16: Effect of the air inlet temperature on the excess air ratio and
angle of maximum pressure. (At constant speed of 2000 rpm, and mass flow
rate 9g/minute of H2.)
The effect of the air inlet temperature on the excess air ratio and angle of
maximum pressure is illustrated in Figure 4.16. As expected, increasing
inlet air temperature advances the angle of maximum pressure, which can
move as much as 5º crank angle (BTDC), therefore affecting directly the
engine efficiency and the mechanical loads on the bearings.
118
110
MRPR (bar/º);Pmax (bar)
100
90
80
70
60
50
40
85
87,5
90
92,5
95
97,5
100
102,5
105
107,5
110
Ta (ºC)
Figure 4.17: Effect of the air inlet temperature on the maximum
combustion pressure and maximum rate of pressure rise. (At constant
speed of 2000 rpm, and mass flow rate 9g/minute of H2.)
Figure 4.17 illustrates the effect of air inlet temperature on the maximum
combustion pressure and maximum rate of pressure rise. It can be seen
that the rate of pressure rise has some dependence on the temperature of
the air entering the cylinder, and this is accompanied by a linear increase
of the maximum combustion pressure. Also, it was identified through the
simulation that increasing the intake air temperature has a bigger influence
on the final compression temperature than an increase in the compression
ratio. This effect was also observed by (Rottengruber et al., 2004) using an
experimental engine.
119
86
420
84
417,5
82
415
80
412,5
78
410
76
407,5
74
405
72
402,5
1,75
2
2,25
2,5
2,75
3
3,25
3,5
3,75
4
Maximum combustion pressure (bar)
Exhaust Gas temperatures (0C)
422,5
70
4,25
Brake load (kW)
Figure 4.18: Exhaust gas temperature and maximum combustion pressure as
a function of engine load (with constant Ta=90ºC at 2200 RPM).
Figure 4.18 shows the exhaust gas temperature and maximum combustion
pressure over the load range with constant intake air temperature and at
constant speed. Considering that the tests were performed at constant
speed, therefore with the same approximate exhaust gas flow, it can be
observed that for break loads up to 2.2 kW the rate of heat loss through the
exhaust is higher than for loads above 2.2 kW resulting in a lower thermal
efficiency. A similar behaviour can be identified in what concerns the
maximum combustion pressure. A possible explanation of this behaviour
can be a characteristic of the engine particular combustion chamber
design, however some authors (Rottengruber et al. 2004) mentioned the
suitability of the HCCI mode for lower to moderate loads, and the
suitability of DIH2 for moderate to higher loads applications.
An increase in the average cylinder charge temperature causes a reduction
in the ignition delay and an increase in the conversion rate of the hydrogen
fuel. Therefore the heat release rates of consecutive operating cycles
become more even, contributing for a better engine controllability through
the reduction of the cyclic variations. However decreasing the ignition
delay and increasing the conversion rate of the hydrogen fuel will result in
120
higher thermal and mechanical component loading, which must be
In-Cylinder. Pressure Rate of Pressure Rise bar/º
considered at the design stage.
85
80
75
70
65
60
55
50
45
40
35
30
25
20
15
10
5
0
-5
-10
0
50
100 150 200 250 300 350 400 450 500 550 600 650 700 750
Crank Angle (degrees)
Figure 4.19: HCCI open pressure diagram and its derivative of test engine
operated at 4.1kW load.
Figure 4.19 shows the in-cylinder pressure plot and its derivative, the rate
of pressure rise, under HCCI operation. The maximum rate of pressure rise
is 11.6 bar/ºCA, taking place before TDC. The development of combustion
from ignition to its maximum pressure takes only 4.2 ºCA. A number of
sharp pressure peaks after the maximum pressure peak can be observed,
being the result of combustion generated pressure waves due to
uncontrolled HCCI combustion characteristics.
121
110
105
2500 rpm
100
Tair inlet (ºC)
95
2200 rpm
90
85
80
75
1500 rpm
70
12gH2/min
65
3
3,25
3,5
3,75
4
4,25

4,5
4,75
5
5,25
5,5
Figure 4.20: Required intake air temperature to sustain combustion as a
function λ, and speed for a 17:1 compression ratio.
Figure 4.20 illustrates one of the most critical aspects of HCCI engine
behaviour: the dependence of the air inlet temperature on the operational
stability of the engine for varying excess air ratios. In order to use the
temperature control of the intake air to control the ignition timing and
combustion, it is critical to include the relationship between these
variables in the engine control system. In particular, the variation of the
minimum inlet air temperature required to sustain combustion for different
speeds and loads must be studied in detail to achieve good engine control.
122
4.6.3 Emissions
Emissions (g/kWh)indicated = f ()
0,8
2200 rpm 17: 1 Tair inlet = 100ºC
g/kWh @ maxi ITE
0,6
H2/10
NO x
0,4
0,2
VOC´s
CO
0
3
3,5
4
4,5

5
5,5
6
Figure 4.21: Emissions at constant speed (2200 rpm) and air inlet
temperature (100ºC) as a function of air fuel ratio.
The exhaust emissions were measured while the engine was operated in
hydrogen fuelled HCCI mode at a speed of 2200 rpm and Ta of 100ºC. The
results of the test are presented in Figure 4.21. As can be seen, the NOx
emissions increase sharply for λ <3.5, and become negligible for higher
values of λ. The NOx levels are considerably lower that what would be
expected for conventional diesel engine operation for all the cases
investigated. The levels of CO and unburnt hydrocarbons (VOC) emissions
are fairly constant over the investigated load range. The levels of these
emissions are negligible for the hydrogen engine, with the only carbon
source being the burning of the lubricating oil.
Figure 4.21 also shows the presence of some hydrogen in the exhaust gases,
and this is due to hydrogen slip which occurs during the valve overlap
period and the non-optimized hydrogen injection valve period. To minimise
hydrogen slip, more accurate control of hydrogen injection is required.
Values of exhaust emissions for the test engine operating in hydrogen
fuelled HCCI and conventional diesel fuelled engine modes are shown in
Table 4.4, illustrating the significant emissions reductions characteristics of
HCCI engines.
123
Table 4.4: Comparison of emissions for DI Diesel and H2 HCCI operation.
H 2 HCCI engine
Diesel engine
NO x
0.01 g/kWh
6.30 g/kWh
CO
~0.00 g/kWh
2.00 g/kWh
Particulate matter
~0.00 g/kWh
0.36 g/kWh
VOC´s
0.015 g/kWh
0.50 g/kWh
4.6.4 Operational stability
Because of the challenges associated with ignition timing control,
operational stability is one of the main challenges in HCCI engines. Small
variations in ignition timing can lead to large variations in the peak cylinder
pressure and the cycle work output, and therefore have large influence on
engine efficiency and emissions formation.
Figure 4.22: Cylinder pressure-volume plots for H2 HCCI operation.
Figure 4.22 shows the pressure-volume plots for 10 consecutive cycles
under HCCI operation. Some variation between the cycles can be seen,
particularly around top dead centre, due to the poor control of the ignition
timing in the HCCI engine.
124
Cycle-to-cycle variations in cycle work output (W) are commonly measured
by the coefficient of variation, COV, defined as
COVW = σW / mean (W),
(4.7)
where σW is the standard deviation of the cycle work and mean(W) is the
mean work output from the cycles.
Sets of 100 consecutive cycles taken at different operating conditions were
analysed to establish the extent of the cycle-to-cycle variations in cycle
work and in-cylinder gas pressure. It was found that between the highest
and the lowest loads tested the coefficient of variation in cycle work
ranged respectively from 7% to 23%. These are acceptable values,
particularly for the higher loads that tend to produce a more stable
operation. The variations in peak in-cylinder gas pressure were higher,
ranging from 15% to 25% over the same load range. This is due to the high
pressure rise shortly after ignition resulting from a fast combustion of
hydrogen, giving large variations in peak pressure from minor variations in
ignition angle. This effect is not desirable as it can result in some engine
instability during load variations.
4.7 DIH2 operation characterization
Similarly as above the DIH2 mode was tested in such a way to characterize
the engine operation at constant load for different equivalence ratios or
fuelling rates, as well as with different inlet air temperatures, thus giving
knowledge of the interdependency between these variables.
125
4.7.1 Auto-ignition of the hydrogen jet
Due to the limited compression ratio of the test engine (17:1) the
temperature at the end of compression was not sufficiently high for autoignition of the injected hydrogen jet. Therefore, heating of the air entering
the cylinder was required as for the HCCI tests above. The temperature of
the air entering the cylinder plays a key role in the control of the RPR and
smoothness of the engine operation, and the relation between these
variables is a critical aspect for developing appropriate engine control
strategies.
12
-5 12.28 (1000/Ta)
 =0.2911+1.33210
e
Auto ignition delay  (ms)
10
8
6
4
2
0
0,7
0,75
0,8
0,85
0,9
0,95
1
1,05
1,1
1000/Ta (1/K)
Figure 4.23: Effect of the end-of-compression temperature on the ignition
delay of the hydrogen jet.
Based on the work of Tsujimura et al. (2003), Figure 4.23 illustrates the
strong dependence of the hydrogen ignition delay on the end-ofcompression temperature. For temperatures below 1100 K, the auto
ignition delay increases sharply and becomes much longer than for
temperatures above 1100K. This dependence follows an Arrhenius function.
The auto ignition delay is therefore strongly dependent upon the
temperature of the air entering the cylinder and the engine compression
ratio, since these variables determine the end-of-compression air
temperature.
126
According
with
Tsujimura
et
al.
(2003)
for
end-of-compression
temperatures below 1100K, the auto ignition delay is longer than that for
Diesel fuel at the same operating conditions, but much shorter delays are
produced for cylinder charge temperatures above 1100K. Due to the high
diffusion of hydrogen the injected hydrogen is rapidly spread all over the
combustion chamber volume, not requiring time for vaporisation of
droplets, and therefore the combustion process proceeds very quickly after
ignition. This gives the benefit of a fast combustion process, and the high
rate of pressure rise is a characteristic of the hydrogen combustion. A third
stage of combustion typical for diesel oil does not exist is practical terms
because, due to the enhanced fuel-air mixing and low hydrogen quenching
distance, the cylinder charge is completely combusted, even in combustion
chamber crevices.
4.7.2 Engine tests
Various injection timings (start- and duration of injection) were tested. In
order to study fully the characteristics of the DIH2 engine, extensive work
for optimization is required. However, here only a limited amount of
optimization work could be carried out as the hydrogen injector was
leaking badly into the hydraulic oil system, and its operating stability was
deteriorating during the tests.
127
In-Cylinder. P. bar Rate of Pressure Rise bar/º
j
y
g
65
60
55
50
45
40
35
30
25
20
15
10
5
0
-5
0
50
100 150 200 250 300 350 400 450 500 550 600 650 700 750
Crank Angle (degrees)
Figure 4.24: Cylinder pressure diagram and its derivative at 5.0 kW load.
Figure 4.24 presents the cylinder pressure plot and the rate of pressure rise
under DIH2 operation at 5.0 kW load. The maximum rate of pressure rise
was found to be 5.1 bar/º and the development of combustion from ignition
to its maximum pressure takes 5.6 ºCA. Comparing with the results above,
it is clear that the rate of pressure rise is significantly lower than for HCCI
operation, as expected. A number of tests were carried out to understand
the effect of the injection timing and duration on the engine operation.
Table 4.5 shows results for varying ignition timing and duration, indicating
that the closer the injection is to the TDC, the higher the IMEP is, and as a
consequence dp/dθ and Pmax become smaller and Texh increases as the cycle
pressure diagram is shifted to the right. From the experimental results it is
suggested that to take advantage of the fast hydrogen combustion the start
of injection can be closer the TDC, resulting in a higher thermal efficiency.
128
Table 4.5: Combustion characteristics as a function of injection timing and
duration (2000 rpm, λ = 5.395).
10º injection
15º injection
20º injection
duration, 10ºBTDC
duration, 15ºBTDC
duration, 20ºBTDC
start of injection
start of injection
start of injection
4.2
4.8
5.0
107
119
123
10º
5º
2º
IMEP (bar)
6.778
6.472
6.196
Texh (ºC)
583
531
506
Pign (bar)
35.97
42.11
43.31
Dp/dθ
(bar/ ºθ)
Pmax (bar)
αPmax
(ºATDC)
16
Diesel oil operation
H2Operation
14
RPR (bar/º)
12
10
8
6
4
2
3
3,5
4
4,5
5
5,5
6
6,5
7
7,5
8
IMEP (bar)
Figure 4.25: Rate of pressure rise as a function of engine load for diesel and
DIH2 operation.
The rate of pressure rise, shown in Figure 4.25, was found to be almost
constant and small for low loads, but increasing rapidly with increasing
load. The rate of pressure rise was much lower than when the engine was
operated with Diesel fuel over most of the load range, however the
pressure rise rate increases rapidly at high loads for the DIH2 engine. This
behaviour is one of the limiting factors of the Diesel engine fuelled with
129
hydrogen, since the operation becomes noisy and the mechanical loads
increase drastically for high engine loads.
46
2500 RPM
45,5
Indicated thermal efficiency (%)
45
44,5
2200 RPM
44
43,5
43
42,5
42
1900 RPM
41,5
41
40,5
0,1
0,12
0,14
0,16
0,18
0,2
0,22
Equivalence ratio ()
Figure 4.26: Indicated thermal efficiency for different equivalence ratios φ
and different speeds.
Figure 4.26 shows the engine efficiency for different engine speeds and
equivalence ratios. It can be seen that the fast combustion of hydrogen
allows the engine to produce high indicated efficiencies at high speeds,
indicating its potential for high power densities and low high-temperature
emissions such as NOx.
130
4.7.3 Emissions formation
1400
2200 RPM Diesel, Hydrogen
NO x=168.7+0.8053e0.992IMEP
1200
NOx (ppm)
1000
800
600
400
200
0
3
3,5
4
4,5
5
5,5
6
6,5
7
7,5
IMEP ( bar )
Figure 4.27: Emissions as a function of engine load under DIH2 and DI Diesel
operation.
Figure 4.27 shows the nitrogen oxide (NOx) emissions for the engine under
DIH2 and conventional diesel engine modes for a range of loads. As
expected, the NOx formation increases sharply at higher loads, due to the
higher temperature levels in the combustion chamber. The DIH2 engine is
seen to produce approximately 20% lower NOx emissions compared with the
same engine in conventional diesel mode. Despite of higher cylinder peak
pressure characteristic of DIH2 mode, resulting from the faster combustion
of hydrogen, the results suggest that the enhanced fuel-air mixing in the
combustion chamber results in a more homogeneous combustion with less
high-temperature zones, therefore inhibiting the thermal NOx formation.
The time required for the combustion of the cylinder charge with hydrogen
and the temperature developed during its combustion are therefore key
parameters that can explain the lower NOx emissions of the hydrogen
fuelled engines.
131
4.8 Efficiency calculations and comparison
As described in Chapter 3, the engine data acquisition and monitoring
software was designed to collect the variables required to calculate the
engine thermal efficiency. On the instrumentation set-up menu of the
software, the net calorific values of the fuels, as well as their densities are
introduced, allowing online calculation of engine thermal efficiency. The
other monitored and logged engine operational parameters were: exhaust
gas temperature; air inlet temperature; air mass flow rate; hydrogen mass
flow rate; diesel oil mass flow rate; and engine speed and power.
Figure 4.28: Engine energy flows considered for thermal efficiency
calculation.
Figure 4.28 illustrates the energy balance of the engine during operation.
Considering the difficulty in measuring the energy losses due to radiation,
these were lumped with the engine cooling losses for these calculations.
Fuel heat input is given by:
132

Q f  m NCV  C
(4.8)
Exhaust gas heat is given by:

Qexh  m exh  C p  T
(4.9)
Brake power:
Qb  T  
(4.10)
Mass flow of exhaust gases:



m exh  m air  m f
(4.11)
The overall heat equation is then:
QF  Qm  Qcool  Qrad  Qexh
(4.12)
Solving in order to determine the heat losses by radiation and cooling of
the cylinder:
Qcool  Qrad  Q f  Qm  Qexh
(4.13)
Q f - Fuel power input [kJ/s];
Qb - Brake power [kJ/s];
Qexh - Exhaust gas energy [kJ/s];
Qrad - Radiation loss of energy [kJ/s];
Qcool - Cooling energy loss [kJ/s];
Qm - Mechanical energy at the engine shaft [kJ/s];
NCV - Net calorific value of the fuel [kJ/kg];
C - Specific fuel consumption [g/kWh];

m f - Mass flow rate of fuel [kg/s];

m exh - Mass flow rate of gases [kg/s];

m air - Mass flow rate of air [kg/s];
C p - Specific calorific value of the gases [kJ/kg K];
 - Angular speed [rad/s] and
T - Engine torque [Nm];
133
4.8.1 Comparison of thermal efficiencies
In order to minimise the differences in terms of mechanical losses friction
and hysteresis, which are dependent upon the engine speed, the engine
was operated at a constant speed of 2200 RPM for the energy balance
comparison. Table 4.6 summarises the results of the tests for thermal
efficiency calculation.
Table 4.6: Comparison of engine energy balance and thermal efficiency at
the maximum reached power at a speed of 2200 RPM.
Diesel DI
HCCI
DIH 2
(Diesel oil)
(H 2 )
(H 2 )
Shaft [%]
27.9
48.0
42.8
Cooling [%]
42.2
20.4
17.3
35.3
31.6
39.9
9000
7076
10280
LOSSES
Exhaust
[%]
Shaft
Power [W]
The highest thermal efficiency was observed for the hydrogen fuelled HCCI
mode, followed by the DIH2 mode. It can be seen that the cooling losses are
significantly lower for DIH2 operation compared with Diesel DI, leading to a
large efficiency advantage. For the HCCI mode of operation, the efficiency
is higher than in direct injection mode, but the power is considerably less
than the other modes of operation, at just above 7 kW.
134
50
Thermal efficiency (%)
45
40
35
30
25
20
Diesel fuel direct injection
HCCI
Bi-Fuel operation
H2 direct injection
15
10
1000
2000
3000
4000
5000
6000
7000
8000
9000
10000 11000
Brake load (Watts)
Figure 4.29: Comparison of brake thermal efficiencies of the test engine for
four operating modes tested.
Figure 4.29 presents the measured thermal efficiency over the full load
range for the four operating modes tested: Diesel DI, DIH2, HCCI H2 and
dual fuel Diesel & H2. It can be seen that DIH2 mode of operation develops
the highest power and a higher thermal efficiency throughout through the
load range, despite the HCCI engine thermal efficiency is higher for its
maximum rated power, it drops for lower loads. The maximum load
achieved under HCCI mode is lowest due to the displacement of intake air
with hydrogen, which is consistent with the results presented by
Rottengrubber et al. (2004). Figure 22 also refers Dual fuel operation
providing an efficiency improvement compared with the standard Diesel DI
mode, having the lowest fuel efficiency. Dual Fuel operation allow existing
Diesel engines to be fuelled with important hydrogen quantities therefore
being a transition technology for the expansion of hydrogen. As a result of
the present research, the author converted two marine Diesel engines of 4
MW to be operated with heavy fuel oil as the ignition source of the cylinder
charge, and a mixture of natural gas and hydrogen.
135
4.9 Uncertainty of measured variables
Prior to the engine testing, the uncertainty associated with the data to be
collected during the experimental work was calculated, since this helps to
gain confidence in the quality of the results and also an understanding of
the
deviation
between
simulated
and
experimental
results.
The
understanding of the causes of deviations should be considered to
determine which part of the error is due to the model inaccuracy, and what
is due to uncertainty in the measured data.
Since any experimental result involves some level of uncertainty that may
originate from a lack of accuracy in measurement equipment, transducer
hysteresis, transducers thermal instability, resonance frequencies and
approximations in data reduction relations, uncertainty analysis is a vital
part of the experimental program and measurement system design. All
these individual uncertainties eventually translate into an uncertainty in
the final results (the so called propagation of uncertainty). Uncertainty is
also important for the model validation, since must be taken into
consideration during the evaluation of the differences of the simulated and
measured results. For this reason, great care was taken during the
experimental system design to choose appropriate transducers, checking if
their characteristics were adequate to monitor the variables involved.
4.9.1 Quantification of uncertainty
Since no statistical analysis of a series of observations was made, according
with the Guide to the Expression of Uncertainty in Measurement Binp et
al, 1995, a type B evaluation of standard uncertainty was used.
Assuming that all the uncertainties have the same level of confidence of
95%, and that all the variables are independent of each other, then the
overall uncertainty associated with the thermal efficiency calculation can
be obtained from the following equation:

R 

WR =  W Xi
xi 
i=1 
n
2
136
(4.14)
WR represents the overall uncertainty;
X represents the independent variable in the thermal efficiency
equation; and
WXi – represents the uncertainty in that variable alone.
4.9.2 Uncertainty in thermal efficiency
The determination of thermal efficiency is one of the key aims of this
research work and therefore the uncertainty of its measurement is of great
importance. Uncertainty is included in the calculation of thermal efficiency
through three different sources: uncertainty in the power measurement,
fuel flow rate, and hydrogen heating value.
The thermal efficiency or fuel conversion efficiency is defined as the ratio
of the work done by the engine divided by the energy input:
ηth =
work .done
Energy.input
(4.15)
Rewriting expression (4.15) in terms of the rate of work, or power, and the
rate of energy input it becomes:
ηth =
Power.Output
Energy.input.rate
(4.16)
The power output can be calculated as the product of the engine torque
and speed, and the energy input rate by the fuel heating value multiplied
by the flow rate.
137
4.9.3 Measurement of thermal efficiency
The measurement of thermal efficiency using a volumetric fuel flow meter
relies on the following equation:
 th 
N
 g V g Qr
(4.17)
N is engine speed [rad/s];  is torque [Nm];  g is the gas density [kg/m3];
V g is volumetric flow rate [m3/s]; and Qr is the heating value per mass of
fuel [J/kg].
The calculation of the mass fuel flow rate using a volumetric flow device
also requires the determination of the gas density. Gas density is not a
parameter that can be measured directly and has to be inferred from other
gas properties. The density of gas can be determined using the ideal gas
law, as shown in Equation 4.18.
ρ=
P
(4.18)
RU
T
Mm
 is gas density [kg/m3]; P is gas pressure [Pa]; T is gas temperature [K];
RU is the universal gas constant [J/mol K]; and Mm is the molecular weight
[kg].
Pressure and temperature can be measured directly. Molecular weight must
be known or can be calculated in the case of a mixture of gases with the
use of gas chromatography.
Although using the ideal gas law provides a convenient method to calculate
the gas density, it is only an idealization of the gas behaviour. Real gas
behaviour approximates an ideal gas only at relatively low pressures and
high temperatures. At other conditions, the density of the gas deviates
from the ideal. To correct this, the concept of a compressibility factor, Z,
is introduced, as shown in Equation 4.19, and is used to adjust the ideal gas
law to fit actual gas behaviour.
138
ρ=
P
=
ZRT
P
R
Z U T
Mm
(4.19)
Figure 4.30: Hydrogen and methane density as a function of pressure @
300K (Source:www.eere.energy.gov)
As can be seen in Figure 4.30, hydrogen density deviates from the ideal gas
relationship substantially for pressures above 15,000 kPa.
4.9.4 Uncertainty in engine power calculation
The calculation of power is usually dependent on two measurements: speed
and torque. The measurement of speed is most likely the least uncertain of
all the measurements made on the engine. Speed is typically measured
digitally, through one or two encoders fitted on the engine camshaft and
crankshaft. The test engine had one encoder fitted at the rear end of the
camshaft with a resolution of 10 Bit and uncertainty equal to ±1/2LSB.
Torque measurement was made trough the use of the developed hydraulic
pressure, generated by a hydraulic pump with fixed displacement. The
139
hydraulic pressure was continuously measured and multiplied by the
volumetric pump efficiency and hydraulic pump flow, which is a function of
pump speed, resulting therefore into shaft power:
P = N H  v H  ηH ,
(4.20)
NH is pump speed [rev/s]; vH is the unitary volume of the pump [3.6
cm3/rev]; and ηH is the volumetric efficiency of the pump.
The hydraulic pump efficiency as a function of load was included into the
online data acquisition and analysis software, thereby accounting for all
the pump inefficiencies as a function of load.
4.9.5 Uncertainty in fuel mass flow rate
The measurement of fuel flow rate is required to calculate thermal
efficiency, and the specific calculation method used depends on what
method is used to measure fuel flow rate, i.e. what type of flow meter is
used. The hydrogen mass flow rate was measured using a Dwyer GFM-1107
mass flow meter with totalizer, shown in Figure 4.31. It is based on a
straight tube sensor with a restrictor flow element to provide high accuracy
(+/-1.5% of full scale) and repeatability (+/- 0.5% of full scale).
The principle of operation of the transducer is based on dividing the flow
by shunting a small portion of the flow through a capillary stainless steel
sensor tube. The remainder of the gas flows through the primary flow
conduit. The geometry of the primary conduit and the sensor tube are
designed to ensure laminar flow in each branch. According to principles of
fluid dynamics the flow rates of a gas in the two laminar flow conduits are
proportional to one another.
140
Figure 4.31: Hydrogen flow meter.
Therefore, the flow rates measured in the sensor tube are directly
proportional to the total flow through the transducer. In order to sense the
flow tube, heat flux is introduced at two sections of the sensing tube by
means of precision wound heating coils. Heat is transferred through the
thin wall of the sensor tube to the gas flowing inside. As gas flow takes
place, heat is carried by the gas stream from the upstream coil to the down
stream coil windings. The resultant temperature dependent resistance
differential is detected by a Wheatstone bridge and amplified. The
measured gradient at the sensor windings is linearly proportional to the
instantaneous rate of flow taking place. An output signal is generated that
is a function of the amount of heat carried by the gases to indicate mass
molecular based flow rates.
The volumetric flow must be converted to a mass flow rate. A great deal of
uncertainty can enter into the calculation of mass flow rate due to the
compound measurements required to calculate the gas density. Pressure
and temperature measurements are needed and taken into account by the
flow meter, also needed is the knowledge of the hydrogen molecular
weight, and lower calorific value.
4.9.6 Uncertainty in the volumetric flow measurement
The thermal efficiency equation for use with a volumetric flow meter was
developed previously. It is shown here explicitly as a function of
measurable variables:
141
 th 
NV H  H  H
 g V g Qr
(4.21)
Taking partial derivatives of the above expression and substituting their
values for a particular test condition and their individual uncertainties
(Kubesh et al., 2002; Wheeler et al., 1996) the calculated uncertainty
associated with the thermal efficiency is calculated by:
  th

 V H

  N H  H p H  g1V g1Qr1

(4.22)
  th

 V N

  V H  H p H  g1V g1Qr1

(4.23)
  th

  H

  V H N H p H  g1V g1Qr1

(4.24)
  th

 p H

  V H  H N H  g1V g1Qr1

(4.25)
  th

 
 g

  VH  H p H N H  g 2V g1Qr1


(4.26)
  th

 V
 g

  VH  H p H N H  g 2V g1Qr1


(4.27)
  th

 Qr

  V H  H p H N H  g 2V g1Qr1

(4.28)
and
142
 ∂
 th  WVH
 ∂V H
WR  
 ∂ th
 W H

 ∂ g
1
2
2
  ∂ th
  
 W pH
  ∂p H
2
2



  ∂ th
  ∂
  
 W N    th  WH
  ∂ H
  ∂N
  ∂ th
  ∂
 
 WVg    th  WQr
  ∂V

  g
  ∂Qr
2
2



2
2






(4.29)
It results from the above calculations that the uncertainty associated with
the engine thermal efficiency is 0.0065, summarised in Table 4.7 for the
test conditions registered during the experiments. Therefore, the thermal
efficiency values calculated from the experimental results have an
uncertainty of ± 0.65%.
Table 4.7: Units and values used for the determination of uncertainties.
The value of ±0.65 % represents a relatively low level of uncertainty,
especially when compared with the values stipulated by the standard ISO
15000, that sets a maximum value of 2 % uncertainty for brake torque, 2 %
uncertainty for speed and 3 % for specific fuel consumption.
A closer inspection of the uncertainty composition reveals that the
individual uncertainty of the hydrogen flow meter (±1.5 %), and the
hydraulic pressure transmitter (±1.0 %) are the main contributors for this
result.
The hydrogen temperature and pressure uncertainties were considered and
combined in the hydrogen flow meter uncertainty calculation. If the
143
hydrogen flow meter measured mass flow instead of volumetric flow, the
result could be improved, since no conversion calculations would be
necessary.
4.9.7 Uncertainty associated with other measurements
The pressure transducer used for the cylinder pressure measurement was a
fibre optic based transducer as described in Chapter 3, with a useful
frequency response range of 0.1 to 25Hz and a maximum housing
temperature of 300ºC, which corresponds an uncertainty of ±1.0% FS under
combustion conditions. Therefore the pressure measurements uncertainty
would be ±3.0 kPa.
Table 4.8: Summary of uncertainties associated with the transducers.
Hydrogen Flow meter
± 1.5%
Hydraulic pressure transmitter
± 1.0%
Cylinder pressure transducer
± 1.0%
Speed sensing system
10Bit ± ½ LSB
Exhaust gas temperature
± 1.0%
Diesel oil flow meter
± 1.5%
4.10 Conclusion
Within the scope of this research project, it was successfully proven that
combustion of the hydrogen in a compression ignition engine is possible by
using any of the three modes (HCCI, DIH2 and Dual-fuel). In addition, the
concept of converting a commercial high speed diesel engine to hydrogen
operation was developed and the tests were conducted on the designed
test bench, operating safely and in acceptable running behaviour.
By means of parameter variations, the influences of the various engine
operating parameters on thermal efficiency, pollutant emissions, and
144
combustion development were studied. The high power, lower emissions
potentials of CI hydrogen fuelled engines was proven.
This chapter also presented the methodology of the data treatment as well
as the data derived from engine tests carried out to characterise each
mode of operation in reference to the diesel fuelled engine.
The performed Dual-fuel operation tests showed a clear efficiency
advantage over the diesel oil operation, with brake fuel efficiency
improvements of up to 5 percentage points. Nitrogen oxides emissions were
comparable to those under normal diesel operation however the emissions
of particulate matter dropped significantly even for small amounts of
hydrogen fuel inducted in the intake air.
The results confirm that dual fuel engines have a significant potential to
improve internal combustion engine performance and reduce exhaust gas
emissions formation. More detailed studies of the mechanisms governing
the in-cylinder processes in dual fuel engines are therefore worthwhile in
order to optimise the design of such engines. Similarly, the use of
alternative fuels, such as bio-oils with poor combustion characteristics,
should be studied to identify potential performance advantages realisable
with hydrogen injection.
Higher flow rates of hydrogen can be achieved, but a control of the Diesel
fuel must be implemented, in such way that the increase of hydrogen flow
rate, will lead to a proportional decrease of the Diesel flow rate,
controlling in this way the amount of energy per cycle. This type of control
is achieved by introducing a transfer function, which above 50% hydrogen
energy per cycle becomes the governing fuel, this implies a transfer of the
PID function, from the diesel governor to the hydrogen governor.
It was found that for HCCI as well as for DIH2 modes of operation the effect
of the temperature of the air entering the engine cylinder has a major
impact in the control of the rate of cylinder pressure rise and this on the
mechanical bearing loads. It was further found that at higher loads the
controllability of the engine is improved but that the rate of pressure rise
can be high.
145
From the energy balances carried to characterise each mode of operation
of the engine it is noticeable that the HCCI mode is characterised by a high
efficiency, but the power is limited by the amount of hydrogen the cylinder
volume can receive. By comparing the losses of heat between the various
modes of operation at the same speed, it was concluded that exhaust
losses are less predominant on the HCCI mode than on Diesel or DIH2 , this
fact can be due to a lower heat input per cycle, and therefore to the lower
temperatures reached at the HCCI exhaust. However DIH2 operation is
characterised by lower cooling losses as the combustion takes place closest
the TDC than the other modes of operation tested. This comparison of heat
losses, can be criticised as the power conditions was not equal for all the
tested modes, and the HCCI mode operation was not optimised in terms of
control The HCCI mode is characterised by an unstable ignition angle that
calls for extensive research to keep it inside acceptable angle variations
not endangering the engine controllability. It can be concluded that the
higher the air temperature at the cylinder inlet, the sooner can be the
ignition of the cylinder charge, therefore contributing for lower engine
efficiency in case of a too early ignition. The direct injection of hydrogen
using high injection pressure allows the control of the ignition angle and if
optimised, can lead to very high thermal efficiencies and high engine
controllability. It was found also that most of that intake air temperature
has a greater effect of the final end of compression temperature that the
engine compression ratio, therefore the use of methodologies for
controlling the cylinder charge at the beginning of compression are very
important. Also it was concluded that cyclic variation are better controlled
when the inlet air temperature is above 70ºC, which necessarily has a
negative effect on the NOx formation. Also it was realised that the increase
in the average mass temperature at the time of the start of injection
causes a reduction of the ignition delay, and an increase in the speed of
the combustion process of the injected fuel, resulting in heat release rates
of the consecutive operating cycles more even. However, a decrease in the
ignition delay results in higher peak pressures, and therefore in higher
mechanical loads on the bearings. The operation of the engine in a Dual-
146
fuel mode, allows the immediate use of hydrogen with important economic
and environment benefits due to the improved exhaust emissions. Hydrogen
can be seen as an energy carrier, which can be produced from renewable or
waste based energy and therefore used as fuel into existing diesel engines
without major modifications of the engine. The controllability of the dual
fuel mode operated engine is satisfactory for hydrogen percentages of 50%
of the energy per cycle. Above that value, a transfer function of the engine
governing actions (PID) needs to be implemented.
147
148
Chapter 5
Modelling and simulation
« Curiosity has its own reason for existing! »
Albert Einstein
This chapter presents of the models used for simulation of the HCCI and
DIH2 hydrogen operated engine, as well as the injector’s dynamic
simulation model used during this research work. A full-cycle simulation
model was developed to investigate the performance of the HCCI and DIH2
engine cycles, but also to serve as a tool for the initial set up of the various
test engine runs and finally to develop practical improvements and
recommendations for future work. The model flexibility allowed the
observation of parameters that are experimentally difficult or expensive to
monitor.
The simulation model allowed the study of the interactions between
various operational variables. The model was kept relatively simple as the
increase in accuracy would require more detailed sub models without
major benefits to the objective of this research work. As a result, the
present simulation model was not developed for in-cycle crankshaft angular
speed variation prediction, since engine and hydraulic brake pump inertia
terms where not considered. The simulation program has a resolution of 0.5
degree of crank angle.
5.1 Modelling of hydrogen HCCI and DIH2 engines
The modelling of compression ignition engines has been investigated and
developed extensively for decades, with varying levels of accuracy,
depending on the modelling objectives ranging from engine dynamics,
combustion development to emissions prediction.
The approach followed to model the HCCI and DIH2 engine cycle was
initially to use a standard air cycle for compression ignition engines. Then a
149
number of improvements around this formulation were implemented to
bring the model closer to the CI engine operation and to HCCI and DIH2
respectively.
The
mathematical
model
was
programmed
using
the
numerical
computation tool box Matlab, which allowed the use of specialised
algorithms for solving the model ordinary differential equations, and also
the construction of a unique and friendly graphic interface.
The following sections describe the details of the simulation model.
5.1.1 Modelling objectives
Engine modelling and simulation was used to reproduce engine operation
and predict its performance. The development of the model was made
based on the actual test engine main characteristics, i.e. bore, stroke,
connecting rod length, piston offset, speed, maximum combustion
pressure, compression pressure, compression ratio, valve angles, valve
dimensions, air inlet temperature, exhaust gas temperatures, specific fuel
consumption, etc.
The objective was to develop a simulation model as close to the actual test
engine as appropriate for the research study. Therefore, the simulation of
the thermodynamic cycle to correlate important cycle events with the
piston position and to determine the engine cycle process parameters was
pursued. The model should also allow the investigation of possible
improvements and solutions for problems specific to the CI engine operated
with hydrogen.
Secondary objectives implemented included data output capability in an
appropriate format to allow detailed analyses and production of highquality graphics. Data logging with a resolution of at least 1º crank angle
was implemented, matching the data acquisition installed on the engine,
and a comprehensive user-friendly human interface was produced.
150
5.1.2 HCCI and DIH2 engine model differences
The differences between HCCI and DIH2 engine operation necessarily affect
B
B
the engine model, however the two modes of operation have in common
the fact that ignition of the cylinder charge occurs due to the high
temperature reached during the compression stroke. The HCCI cycle power
output is limited by the amount of hydrogen that forms the cylinder
charge, because of its high stoichiometric volume at the cylinder charge
pressure and the displacement of intake air. The DIH2 cycle has a much
B
B
higher limit concerning the amount of hydrogen per cylinder charge; in
practice physical component constraints limit the amount of hydrogen that
can be injected in each cycle.
In the HCCI mode of operation, the cylinder charge can be considered a
homogeneous mixture of hydrogen and air. In contrast, the cylinder charge
of a DIH2 hydrogen engine is a stratified charge (despite the hydrogen
B
B
diffusion characteristics), where air is compressed during the main part of
the compression stroke, the hydrogen being injected only when the self
ignition temperature has been achieved and determined by the engine
crank angle.
The cylinder charge ignition in the HCCI engine takes place somewhat
erratically, depending very much on the heat transfer characteristics of the
cylinder to the cylinder charge, ambient air temperature, charge
conditions, etc. The simulation model included a heat loss calculation from
the combustion chamber components, cylinder head and liner to try and
predict this behaviour.
The turbulence of the in-cylinder air and fuel charge is an important factor
in conventionally fuelled engines, however this was not modelled, as for
HCCI the cylinder charge is considered homogeneous and for DIH2 the high
diffusion and short quenching distance of hydrogen would produce a close
to homogeneous cylinder charge within a very short time period. With DIH2
engine operation, combustion coincides with the precise timing of hydrogen
injection as no fuel vaporization or preparation timing phases are required
151
for modelling ignition of hydrogen. However, combustion chamber heat
losses are not so critical in determining the fuel ignition as for the HCCI
mode of operation.
For HCCI engine operation the cylinder hydrogen charge is introduced at
low pressure through inlet port injection, typically below 6 bar, and air and
hydrogen are then compressed until the ignition temperature is reached. A
modified polytropic compression exponent was determined and used for
the working fluid since this is a hydrogen and air mixture which is
compressed. For DIH2 operation, hydrogen at 90 bar is injected near the top
dead centre, into a compressed hot air atmosphere.
5.2 Simulation program description
In the following section the simulation program structure is presented. The
decomposition of the simulation program according to various levels of
detail shows the interaction between various program blocks and the
overall program structure.
5.2.1 High level program structure
Figure 5.1 shows the high level structure of the simulation program which is
described in the next paragraphs. The inputs of the simulation program can
be grouped into, engine parameters, engine ambient conditions and model
parameters. As a result of the resolution and number of iterations, the
thermodynamic cycle is calculated and the results are displayed graphically
and numerically. Using the implemented cursor facility it is possible to
analytically analyse details of the graphic outputs.
152
ENGINE
PARAMETERS
MODEL
PARAMETERS
ENGINE
CYCLE
CALCULATION
AMBIENT
CONDITIONS
GRAPHICAL &
NUMERICAL
OUTPUT
Figure 5.1: Simulation model structure.
5.2.2 Model parameters
Constant model parameters that are required, such as polynomials
coefficients, constants and tables, are grouped together in the program
and allow the initiation of the iterative engine simulation.
153
5.2.3 Engine parameters
These parameters define the engine geometry and include the following:
stroke, connecting rod length, piston offset, piston cylinder head ratio,
piston crown area, bore, compression ratio, clearance volume, exhaust
valve maximum diameter, exhaust valve minimum diameter, inlet valve
dwell angle, exhaust valve dwell angle, inlet valve opening angle, inlet
valve closing angle, exhaust valve opening angle, exhaust valve closing
angle, and engine speed. These parameters also include the start of
injection and duration of injection for DIH2 operation, as well as pulsed
injection frequency and duty cycle for the pulsed injection.
5.2.4 Ambient conditions
These parameters define the ambient conditions of the simulation run,
including: atmospheric pressure and temperature, exhaust back pressure,
fuel net calorific value, fuel specific gravity, specific fuel consumption, and
ignition delay.
5.3 Engine cycle calculation
This part of the simulation performs the calculation of the thermodynamic
cycle using the model and engine parameters at the ambient conditions.
The engine cycle calculation model can be decomposed into the following
structure, as shown in Figure 5.2.
154
Figure 5.2: Engine cycle model structure.
155
5.3.1 Piston crank mechanism
For the thermodynamic simulation of the engine cycle it is vital to know
the precise cylinder volume at any crankshaft angular position.
Figure 5.3: is a representation of the piston crank mechanism.
Figure 5.3 is a representation of the piston crank mechanism and the
variables shown are:
Lcr is the connecting rod with a length (m);F is the projected length of the
B
B
connecting rod (m); Lct is the length of the crank (m); G is the crank
156
projected length over the piston axis (m); E is the crank projected length
over a perpendicular to the piston axis direction (m);  is the crank angle
referred to TDC (rad);  is the angle defined between the connecting rod
and the piston axis (rad); D, P and Q define the position of the connecting
rod pin centre; Lst is the stroke length (m).
At any given crankshaft angle, θ, after the TDC position of the crank, the
connecting rod centre line assumes an angle , to the centre line. This
angle is designated as the “angle of obliquity” of the connecting rod. The
bearing location where the connecting rod attaches to the piston is
denominated the “small end” and may be offset by an amount D to the
cylinder centre line. The offset is positive if that offset is positive toward
the direction of the crank rotation, as in Figure 5.3. The direct implication
of the offset is that when the crank is at the top dead centre, the piston
will be not at its top dead centre, in this case the piston TDC and BDT will
occur at crank angles respectively θtdc , and θbdc. Using the nomenclature
B
B
B
B
and Figure 5.3, it is possible to derive the piston position as a function of
the crank angle θ. With the piston at TDC:
Ftdc + Gtdc =
Lcr + Lct 2  D 2

D
θ tdc = tan 1 
 Ftdc + Gtdc
,
(5.1)

 ,

(5.2)
and
Gtdc =L ct cos θ tdc .
(5.3)
Therefore
Ftdc =  Lct cosθ tdc +
Lcr + Lct 2  D 2
.
(5.4)
With the piston at BDC:
Fbdc =
Lcr  Lct 2  D 2 ,
(5.5)
157
θ bdc = tan 1 (
and
D
),
Fbdc
(5.6)
Gbdc = Lct cosθ bdc .
(5.7)
The stroke of the piston from TDC to BDC is then given by
L st =F tdc +G tdc− F bdc .
(5.8)
If the gudgeon pin offset and/or cylinder axis offset is zero, i.e., if D is
zero, then:
θ tdc= 0
θ bdc = 0
F tdc =Lcr +Lct
G tdc=L ct
F bdc =Lcr -Lct
And the length of the stroke becomes:
Lst = Ftdc + Gtdc - Fbdc = Lcr + Lct  (Lcr - L ct ) = 2L ct
(5.9)
Considering a position within the cylinder of any point on a piston with
respect to its motion from its TDC position to a point where the crank has
turned through an angle  from the TDC angular position of the crank. If for
convenience such a position is marked as X and is located at the small end
bearing centre and its location down the cylinder from its TDC position is Ht
B
as illustrated in the upper sketch of Figure 5.3, the length H t is:
b
H t = Ftdc + Gtdc   F + G  ,
(5.10)
E = Lct sinθ ,
(5.11)
G = Lct cosθ ,
(5.12)
and
F = L2cr  E - D  .
2
(5.13)
158
B
Therefore,
Ht =
Lcr + Lct 2  D 2
 L2cr  Lct sinθ  D   Lct cosθ ,
2
(5.14)
and the angle of obliquity of the connecting rod is given by:
1





Lcr
2
 E-D 
Φ = tan -1  (
)  = tan -1  ((
)  - 1 )
 F 
 Lct sinθ  D  
(5.15)
5.3.2 Heat losses
The derivation of a global heat transfer coefficient has been researched
extensively, but without consensus on the most adequate methodology. A
widely used methodology recommended by Blair (1999), which is based on
Annand´s work, Annand et al. (1963) was implemented in the simulation
model.
The heat losses from an engine combustion chamber are illustrated in
Figure 5.4.
Annand et al. (1963) developed one of the most widely used and accurate
methods for deriving the heat transfer in spark ignition engines while
Eichelberg, Woschni and others developed models for the calculation of
cylinder heat transfer of diesel engines.
The method implemented in this research is based on Annand´s
methodology
as
hydrogen
combustion
using
homogeneous
charge
compression ignition and direct injection is closer to constant volume
combustion than to constant pressure and volume combustion.
This approach separates out the convection and radiation terms, and this
distinguishes it from the heat transfer theories of the other researchers.
The expression of the Nusselt number, Nu, leads to a conventional
derivation for the convection heat transfer coefficient, Ch.
B
159
B
Q L Heat loss
through the
cylinder head
Q L Heat loss
through the
cylinder wall
Q L Heat loss
through the
cylinder wall
Q L Heat loss
through the
piston crown
Figure 5.4: Engine cylinder heat losses.
160
Annand recommends the following relationship between the Reynolds and
Nusselt numbers for a four stroke engine:
N u =a Re 0. 7 ,
(5.16)
the constant a has a value of 0.46. The Reynolds number is calculated by:
Re =
ρ cy C p d cy
μcy
,
(5.17)
The working fluid density is derived from the prevailing cylinder pressure,
temperature and gas properties using Equation 5.18.
ρcy =
p cy
Rcy Tcy
,
(5.18)
pcy is the typical cylinder pressure (Pa), and Tcy is the prevailing cylinder
B
B
temperature (K).
During the compression stroke, the gases inside the cylinder will be a
mixture of air and hydrogen if the engine is operated as a HCCI engine. For
DIH2 operation, the gases will mix only after the compression and injection
B
have taken place. From the induction stroke to the exhaust of the gases,
the gas constant varies, as do other gas properties. Therefore, tracking of
the gas constant during the process is very important. The viscosity of the
gases inside the cylinder with temperature and pressure, but according to
research by Blair (1999), little accuracy is lost if the expression for the
viscosity of air is used. A good approximation of µ of air can be expressed
as:
μ = 7.457  10 6 + 4.1547  10 8 T  7.4793  10 12 T 2 ,
(5.19)
The mean piston velocity is a function of the piston stroke and the engine
B
speed and is given by
Cp =
2L st N
,
60
Lst is piston stroke (m) and N is the engine speed (rpm).
B
161
(5.20)
Once the Reynolds number is calculated, the convection heat transfer
coefficient, Ch, can be obtained from the Nusselt number as follows:
B
B
Ch =
C k Nu
,
d
(5.21)
Ck is the thermal conductivity of the in-cylinder gas (W/m2K) and Tcy is the
B
P
P
B
instantaneous gas temperature (K).
The parameter Ck represents the thermal conductivity of the gas in the
B
B
cylinder that can be assumed equal to that of air at instantaneous cylinder
temperature (Blair, 1999) and it may be determined by:
C k = 6.1944 10 3 + 7.3814  10 5 T  1.249110 8 T 2
(5.22)
Annand also considers the radiation heat transfer coefficient, Cr, to be
B
B
given by:
C r = 4.25  10
9

4
Tcy4  TcW
Tcy  Tcw
,
(5.23)
Tcw is the average cylinder wall temperature (K).
B
It should be noted that Cr is much smaller than Ch, which is why it is
B
B
B
B
neglected by many researchers when a reasonable simplification is desired.
This is particularly the case for this research since hydrogen flame radiation
is extremely small due to the fact that hydrogen combustion do not
produce radiating particulate matter normally originated by the combustion
of hydrocarbon fuels. However due to the lubricating oil combustion some
particulate may exist, although negligible without any impact on the
simulation results.
The value of Tcw is the instantaneous average temperature of the cylinder
B
B
wall, the piston crown and cylinder head surfaces. The infinitesimal heat
transfer dQL during a crankshaft angle interval dθ, corresponding to an
B
B
interval dt can be calculated for the mean value of that transmitted to the
total surface exposed area to the cylinder gases by
162
dQ L = C h  C r Tcy  Tcw Acw dt ,
dt =
(5.24)
dθ 60

360 N
(5.25)
and the surface area of the cylinder, Acw is the summation of the cylinder
B
B
head, piston crown and cylinder liner areas.
It should be noted that the heat transfer coefficients increase dramatically
when combustion is taking place, but that is when a minimum surface area
is exposed, and as the combustion of hydrogen is extremely fast, the time
of exposure of the combustion chamber surfaces to high temperatures is
very short.
Combustion
chamber
Liner
Figure 5.5: Engine cylinder head and liner thermal image.
Figure 5.5 shows the thermal image of the test engine cylinder and it is
evident that there is a greater heat flux in way of the combustion chamber,
and its distribution along the cylinder liner.
163
The average temperatures measured using the thermal imaging camera was
used in the engine model, and the assumption was made that the piston
crown
temperature
was
the
same
as
the
combustion
chamber
temperatures. This assumption and the temperature measured were
applied to HCCI and DIH2 simulated modes of operation.
B
B
5.3.3 Cylinder valves modelling
To model a diesel engine it is essential to calculate the geometrical
passage areas exposed by the valves at any crank angle. There are two
aspects to this requirement. The first is the exposed area at any particular
valve lift, and the second is the valve lift characteristic as a function of the
crank angle.
A representation of the intake and exhaust valves and manifolds is shown in
Figure 5.6.
Figure 5.6: Valve apertures (Blair, 1999).
The intake and exhaust manifold areas are denoted by Aim and Aem
respectively. The flow areas across the ports of the valves are Aip and Aep
for an intake and exhaust valve, respectively. The total flow areas are
Aipt = niv Aip,
(5.26)
and
164
Aept = nev Aip,
(5.27)
niv and nev are the number of intake and exhaust valves respectively.
Therefore, the intake manifold to valve area ratio can be defined as:
C im =
Aim
A
= im ,
niv  Aip Aipt
(5.28)
and exhaust manifold to valve area as:
C em =
Aem
n ev× Aep
=
Aem
Aept .
(5.29)
These ratios are critical to the performance of an engine since they
determine the amplitude of any pressure wave created in the ducting by
the cylinder state conditions.
Figure 5.7: Valve geometry (Blair, 1999).
The inflow or outflow of any valve depends on the areas which correspond
to the side areas of a frustum of a cone. Therefore, according to Figure
5.7, the controlling aperture areas for the exhaust and intake ports at the
values are defined as follows:
Aept = nev
π 2
(d ip  d 2 st ) ,
4
165
(5.30)
π 2
(d ip  d 2 st )
4
Aipt = nip
(5.31)
Poppet valve aperture geometry
Referring to Figure 5.8, the aperture geometry of the poppet valve is a
function of the lift L, above the seat, and the angle ø, which is defined by
the inner and outer diameters dis and dos respectively. Therefore, a
B
B
B
manifold to valve curtain area, At, models the side surface area of a
B
B
cylinder of diameter dis and lift height L and is given by:
B
B
At  π  d is  L
(5.32)
The area of passage is defined as the area of a frustum of a cone defined
by the side length dimensions x, the valve seat angle ø, the inner and outer
seat diameter dis and dos, and the radius r, all of which are dependent upon
B
B
B
B
the value lift L.
The side surface of a frustum of a cone As, defines the maximum
B
B
geometrical flow area and is given by:
As = π(
d major + d minor
2
)x ,
(5.33)
x, is the length of the sloping side and dmajor, dminor are the cone top and
B
B
B
B
bottom diameters.
As can be seen in Figure 5.8, x has two distinct values dependent on valve
position. When the valve lift is very small, x is no longer normal to the
valve seat at angle ø. Therefore, minimum valve lift is given by:
Llim 
d os  d is
d  d is
 os
2 sin  cos 
sin 2
(5.34)
For the first stage of the poppet valve lift where L≤ L lim , the valve curtain
area At is a function of x and r and can be calculated by:
B
B
166
X=Lcos ø ,
r=
d is
+ x  sinø , and
2
At = π  L  cosø(d is + L  sinø  cosø ) .
(5.35)
(5.36)
(5.37)
For the second stage of poppet valve lift where L > Llim the valve curtain
area At is given by:
B
B
At = π(
d os + d is
d  d is
d  d is 2
) (L  os
tanø) 2 + ( os
) .
2
2
2
(5.38)
Since the valves of the test engine have a conventional seat angle of 45º,
Equation (5.38) simplifies to
At = π(
d os  d is
d  d is
d  d is 2
) (L  os
) + ( os
) .
2
2
2
(5.39)
In practice, for valve lift more than 40% of the inner valve seat diameter,
resulting losses that are characterised by a discharge coefficient C d.
B
b
Therefore, the effective area of the valve throat restriction becomes Atc,
B
B
which is defined as follows:
For At < A p :
Atc = C d  At ,
(5.40)
Ate = C d  A p ,
(5.41)
and for At  A p :
At is valve throat area and Ap is valve minimum area (m2).
B
B
B
B
Poppet valve lift
167
Valves cannot instantly lift or drop, therefore the valve commences its lift
at a crank angle Vo and upon closing returns to zero lift at a crank angle V c.
B
b
B
To model the valve operation, the precise valve timing must be known to
compute the valve aperture areas at any instant during crank shaft
rotation.
The poppet valve lift can be described in five different phases, as shown in
Figure 5.8.
Figure 5.8: Valve lift characteristics as a function of crank angle (Blair,
1999).
The ramp up phase has a duration of θºur, with the valve lift going from 0 to
B
B
Lur. The main lift up phase is from the end of the ramp up to the beginning
of the dwell period around the maximum lift point. This period is θur and
B
B
the valve lifts from Lur to L v. The dwell phase occurs around peak lift when
B
B
b
B
the valve remains at Lv for a period θdw. The main lift down phase
B
B
B
B
corresponds to the valve drop from the end of dwell phase to beginning of
the final ramp down phase. It lasts for θdw degrees and the valve falls from
B
B
the lift Lv to L dr. The ramp down is the final phase and lasts for θdr with the
B
B
b
B
B
B
lift decreasing from Ldr to zero.
B
B
It is quite usual for engines to have valve opening and closing ramps which
are similar in crank angle duration and valve displacements Lur and L dr.
B
168
B
b
B
Acceleration characteristics of a poppet valve
Assuming that the movement of the valve at any point of its lift is given by
dL, over an infinitesimal time interval dt, when the engine speed is N.
Since the engine rotates dθ degrees during dt then
1
dt 60 / N
=
=
.
dθ
360
6N
(5.42)
Assuming that the valve lift is in mm, the valve velocity Cv is given by:
B
Cv=
1
dL
×
1000 dt .
B
(5.43)
therefore
Cv=
1
dL dθ
×
×
1000 dθ dt
(5.44)
hence
Cv =
6N dL

.
1000 dθ
(5.45)
Defining the valve velocity variation as dCv during the time interval dt,
B
B
then the acceleration is given by:
dCv 1
 ,
dt
g
(5.46)
dCv dθ 1

 .
dθ
dt g
(5.47)
6N dCv

.
9,81 dθ
(5.48)
av =
where g = 9,81m / s 2 . Therefore
av =
Hence
av =
169
A lookup table of valve lift Lv at one degree increments was required to
model the lift of the engine valves. Figure 5.9 shows lift L1, L2 and L3 and
B
B
B
B
B
B
from Equation (5.49) the mean valve velocity for the two crank angle
intervals is given by:
C v1 =
6N L2  L1

,
1000 θ 2  θ1
(5.49)
C v2 =
6N L3  L2
.

1000 θ 3  θ 2
(5.50)
Figure 5.9: Three adjacent points on a valve lift curve (Blair, 1999).
Therefore the mean acceleration for the lift process as a motion from the
median point of the first element, to the median point on the second
element is given by
a v12 =
C v2  C v1
6N

9,81 0.5θ 2 + θ 3   θ1 + θ 2 
(5.51)
From this, the valve lift can be calculated for each crank angle increment
at any given engine speed.
170
Valve lift characteristics
While developing the simulation model, the option of creating a realistic
valve lift profile was assumed rather than using the measured engine data
of valve lift versus crank angle. The reason was that no angle marks were
available on the engine flywheel, and incorrect angle measurements could
produce an unrealistic valve lift profile.
Since the design of the valve train is a very specialized process, it was
decided to inspect various engine valve lift versus crank angle data to
implement a numerical method of modelling the valve lift. Mathematically,
there is no single function capable of accurately representing the entire
specific lift function. Reverting to the concept of specific lift and specific
angle and referring to Figure 5.7, specific lift and specific angle are given
by:
L s=
θs =
Lθ
Lv
θ
θv
Figure 5.10: Specific lift characteristics of a poppet valve (Blair, 1999).
Figure 5.10 shows the specific lift and specific angle relationships for a
ramp period and a lift period. The same polynomial relationship is used for
171
the “ramp down” as is used for the “ramp up”, and similarly for the “main
lift up” and the “main lift down”. The relationship linking specific lift and
angle is a third-order polynomial in each case, the coefficients of which are
determined from an analysis of measured data. The functions are the
following:
For ramp up and ramp down:
L s = K ro + K r1 θ s + K r2 θ s2 + K r3 θ s3 .
(5.52)
For main lift up and main lift down:
L s = K mo + K m1 θ s + K m2 θ 22 + K m3 θ s3
(5.53)
Procedure for modelling valve lift characteristics
The procedure for modelling valve lift characteristics can be carried out in
the following steps.
Step 1: Opening and closing angles of the valves are “designed” i.e.,
numerical values are assigned to Vo and Vc as crank angles referring to the
B
B
B
B
engine TDC, as θ = 0º.
Step 2: Calculate the valve opening duration using:
θ v =V c− V o
(5.54)
Step 3: Set the duration of the “dwell angle” θdw as the number of crank
B
B
angle degrees, typically in the range of 0º to 10º.
Step 4: The “ramp up” and “ramp down” periods, θvr and θdr are calculated
B
B
B
B
by assuming they take place over a number of crank angle degrees.
Similarly, the main lift periods, up and down, θul and θdl, are each a
B
B
B
B
number of crank angle degrees.
The sum of all these angles must be equal to θv (see Figure 5.8). According
B
B
to Blair (1999), the main lift periods are identical for each ramp period.
Therefore:
θ ul =θ dl=
θ v − θ dw − θ ur − θ dr
2
172
(5.55)
Step 5: Assign the valve lifting associated with the up and down ramp
periods, i.e., Lur and Ldr as a fraction of the maximum lift Lv. These will be
B
B
B
B
B
B
greater than 50% of Lv for CI engines.
B
B
The valve lift ratios for the up and down ramp periods are defined as Cur
B
B
and Cdr respectively.
B
B
Ramp lift ratios
The ramp lift ratios are defined as
C dr =
Ldr
,
Vv
(5.56)
C ur =
Lur
.
Lv
(5.57)
The valve lift curve can be computed by considering each element in
sequence as shown in Figure 5.10, starting with valve opening and the
opening “ramp up”.
Valve lift commences at
Lθ = 0
θ = 0º
(5.58)
The opening ramp up period at any angle such as
0 < θ  θ ur
(5.59)
Therefore the specific angle is calculated by
θs =
θ
θ ur
(5.60)
Inserting θs into Equation (5.53) to calculate L s. Hence the valve lift is
B
B
b
B
given by
Lθ = C ur  Ls  Lv
173
(5.61)
Main lift up period
For
θ ur < θ  θ ur + θ ul
,
(5.62)
the specific angle is given by
θs =
θ  θ ur
.
θ ul
(5.63)
Inserting θs into Equation 5.53 and calculating Ls, the value of the actual
B
B
B
B
lift, Lθ is found by
B
B
Lθ = Lur + Ls (Lv  Lur ) .
(5.64)
θ ur + θ ul < θ  θ ur + θ ul + θ dw
(5.65)
The dwell period
For
and specific angle θs=1, the value of the actual lift, Lθ, is simply the
B
B
maximum lift: L θ=Lv
B
b
B
B
Main lift down
For
(θ ur + θ ul + θ dw ) <   (θ ur + θ ul + θ dw + θ dl ) ,
(5.66)
The specific angle θs is
B
θS =
θ ur +θ ul +θ dw +θ dl − θ
θ dl
.
(5.67)
The angle is calculated in reverse to obtain the valve drop. Inserting the
specific angle θs into Equation 5.53 and calculating Ls, the value of the
B
B
actual lift, Lθ is found by:
B
B
Lθ = Ldr + L s (Lv  Ldr ) .
The ramp down
For
174
(5.68)
θ ur + θ ul + θ dw + θ dl < θ < θ v ,
(5.69)
the specific angle is
θs =
θv  θ
.
θ dr
(5.70)
It should be noted that the values of angle and lift are determined by their
position from the start of the down ramp, i.e. the computation is operated
in reverse for “ramp down” by comparison with “ramp up”.
The final point at θv is not calculated, but is reserved for a positive ”shut”
B
B
in the next segment below. Inserting θs (Equation 5.52) the specific lift Ls is
B
B
B
B
then calculated. The value of the actual lift, Lθ, is found by:
B
B
Lθ = C dr  L s  Lv
(5.71)
Valve shutting
For
L θ= 0
θ=θ v
The use of the positive zeroing of the valve lift curve, at opening and
closing, takes care of the numerical problems caused by the polynomial
coefficient Kro in Equation (5.52) not being an actual zero.
B
B
This problem is evidenced on air flow graphs and charge mass graphs by
showing abrupt “cuts” corresponding to discontinuities of the functions on
their connecting transition points. There are some methods for smoothing
these abrupt cuts, but the routines implemented on the model were not
sufficiently robust to create the desired smoothing of the graphs.
As expected, there will be some function problems due to the fact that
those values of lift are calculated from different functions, giving rise to
unacceptably high levels of velocity and acceleration of the valve.
5.3.4 Ignition delay
Ignition delay is dependent on the auto ignition temperature of the
hydrogen, and for a gaseous injection of hydrogen this is solely dependent
175
on the temperature of the cylinder charge. This dependence was modelled
using experimental data derived by Tsujimura (1999) and is given by the
following expression:
 1000 
) 

T
 a 
12.28 (
τ = 0.2911+1.332  10 5  e
(5.72)
This expression, which is a function of the cylinder charge temperature,
was used to define the ignition crank angle.
For cylinder charge temperatures below the hydrogen self ignition
temperature, there is no ignition and therefore no combustion. For cylinder
charge temperatures above the self ignition temperature, combustion takes
place and is affected by a delay given by the expression above. Therefore,
the angle of ignition is a function of the ignition delay and the cylinder
charge temperature.
In the case of DIH2 it was assumed that the cylinder charge temperature is
always above 1100 K, the self-ignition temperature of hydrogen, therefore
resulting in extremely short ignition delays.
5.3.5 Mass flow calculation
The cylinder conducts an intake stroke, in which a mass of fresh air is
induced through the inlet valve into the cylinder from the atmosphere.
Therefore, this variable varies with atmosphere pressure and temperature.
The ambient conditions Ta and Pa influence directly the air density, and can
be expressed by
ρa =
Pa
.
R a Ta
(5.73)
Therefore, the volumetric efficiency, ηv, of the engine, which is an input
parameter of the simulation program, is related directly to the mass of air
supplied through the inlet manifold and inlet valve during the intake
176
period, divided by mvref which is the mass required to fully fill the cylinder
with a swept volume Vsv under the cylinder prevailing condition. Therefore:
ηv =
ma
,
mvref
m vref = ρ aV sv .
(5.74)
(5.75)
Assuming that the cylinder has induced a mass of fresh air at standard
conditions of pressure and temperature, the reference density ρdref is given
by
ρ dref =
Pdref
Ra Tdref
(5.76)
The delivery ratio, DR, of the engine is defined as the mass of air supplied
through the inlet valve during the intake period divided by the mass of air,
mdref, with a perfect suction into the swept volume and at standard
reference density ρdref. The delivery ratio is then defined as:
DR =
m as
m dref
(5.77)
where
m dref = ρ dref Vsv
(5.78)
5.3.5 Combustion
The CI engine air fuel ratio
Hydrogen (H2) as a CI engine fuel has the following stoichiometric equation
177
79
79


2 H 2  O2 
N 2   2H 2 O 
N2
21  
21

(5.79)
Therefore the air fuel gravimetric ratio is given by
79
 28
21
= 34.332 .
2 1 2
1  32 + 1 
AFR =
Unlike for diesel fuels, particulate matter emissions are not produced by
the hydrogen combustion, allowing operation with richer cylinder charges if
required. Also, due to the hydrogen physical properties such as, high
dispersion, lower explosive limit, low quenching distance and high flame
speed, its possible to run CI engines with very high air fuel ratios.
Cylinder trapping conditions
The air trapping point is considered the point where the intake valve
closes, therefore the total mass of air available for combustion is
dependent on the pressure and temperature conditions at the trapping
point, which is given by:
mtr =
Ptr  Vtr
,
Rtr  Ttr
(5.80)
and:
V tr =V ivc +V cv
.
(5.81)
Considering that air is the prevailing gas inside the cylinder there is little
error assuming that Rtr=Ra. The dominant variable of the trapping process is
the pressure inside the cylinder Ptr. The trapping pressure is directly
controlled by the pressure wave dynamics of the intake and exhaust system
(Blair, 1999).
178
Pressure and temperature profile calculation
The CI engine cycle can be represented by five processes, corresponding to
the air-standard dual cycle as shown in Figure 5.11:
Adiabatic and isentropic compression 1-2;
Constant volume heat addition (combustion) 2-3;
Constant pressure heat addition (combustion) 3-4;
Adiabatic and isentropic expansion 4-5;
Constant volume heat rejection, exhaust process 5-1;
Figure 5.11: Four stroke CI engine pressure volume cycle.
To perform the calculation of the dual cycle, data from the engine
parameters related to the engine geometry are used in Equation 5.14 to
define the piston position Ht at any crank angle, thus allowing the
calculation of cylinder volume and temperature profiles.
To simplify the calculation of the various equations an incremental method
was used. The volume increment is approximated by:
V + V1 
V step = Vn +  n+1

2


179
(5.82)
and the temperature increment is approximated by:
dT
= Tn+1  Tn
dθ
(5.83)
Adiabatic and isentropic compression
It was assumed that only air is compressed, however this is not exactly
correct in the case of the HCCI engine since a mixture of air and hydrogen
is compressed. A correction factor using the ratio of the specific heats was
therefore implemented, using the following expression recommended by
Blair (1999):
k = 1.4373  1.318  10
4
 T + 3.12  10
8
4.8  10 2
T 
,
λ
2
(5.84)
T is the average cylinder charge temperature and  is the average excess
air factor. The mass of air trapped in the cylinder, mta, is given by
mta =
P1Vsv
,
RT1
(5.85)
Vsv is the swept volume, R is the air universal constant, and T1 and P1 are
the initial temperature and pressure.
5.4 Engine simulation program structure
The simulation program was implemented using Matlab (The MathWorks
Inc., 2006). The implemented program structure is presented in Figure 5.12
and its respective functions are described below.
180
Load model parameters and
extract variables from
parameters structure
STAR
C a lcu la t e va l ve are a
p r o f il es
C a lcu la t e pi s ton dis plac em en t
a nd c y li nde r v o lum e p r o f il es
as a func t io n o f ϴ
C a lcu la t e c y l i nde r pres s u r e
a nd te mp era t ure pro f iles usi ng
T su bfunc tion
C a lcu la t e pi s ton dis plac em en t
a nd c y li nde r v o lum e p r o f il es
as a func t io n o f ϴ
C a lcu la t e P , T , dP/ dϴ as a
fu nc tion o f ϴ
END
Plot graphs P, T,
dP/dϴ as a
function of ϴ
Figure 5.12: Simulation program process.
181
Function Main Engine Cycle
This function calculates the working fluid temperature and pressure as a
function of crank angle, engine speed and ambient conditions. The input
variables of this function are:
RPM: Engine speed (rpm);
P a : ambient air pressure (Pa);
B
b
Ta: ambient air temperature (K);
B
B
Qin: combustion energy input function (J/s);
B
B
Function Barrel
In order to calculate the flow through valves and other parameters, it is
necessary to shift the engine cycle and other parameters, so that the initial
state is one of known, quasi-static conditions. For this state, the inlet and
exhaust valves are both closed and the in-cylinder pressure is known or can
be predicted reliably.
The Barrel function is used to shift all crank-angle related values calculated
so far to a suitable point to enable further calculations. The same function
will then be used to restore the cycle to its original position on the crank
angle axis. The cycle is shifted so that the initial condition is immediately
before the exhaust valve opens.
Function Pressure
This function calculates the cylinder pressure profile over a complete
engine combustion cycle. The initial cylinder pressure is the variable Pstart.
B
B
The function uses an iterative process to calculate the results for the input
() range, using the inlet and exhaust valve area profiles given by the
variables iav area and ev area, previously defined.
B
B
The engine speed is given by the variable omega in rad/sec, and the
cylinder volume is given as a function of theta by volume. The data is
shifted to give the starting point where inlet and exhaust valves are closed,
182
starting immediately after inlet valve closes, and assumes that cylinder
pressure is atmospheric.
Function data
This function returns engine parameters. If an input argument is supplied,
it also plots a graphical representation of the valve timing,
Function Engine model
This function contains the whole engine model as separate sub-functions,
and is divided into three distinct intervals:
From beginning of cycle until start of injection.
From start of injection until end of combustion.
From end of combustion until exhaust valve opens.
Calculations are made for each of these cycle periods.
Functions valve lift and area
The valve lift function calculates the valve lift profile based upon the
duration as defined by the opening and closing angles. The valve area
function calculates the valve areas in relation to the crank angle for given
valve dimensions and lift profiles. The methods implemented are those
given by Blair (1999), as described above.
183
5.4.1 Simulation program interfaces
Figure 5.13: Simulation program human interface.
The simulation program was designed to have a graphical user interface as
shown in Figure 5.13. This allows the geometric engine parameters,
injection timing, valves timing, and all the other parameters to be defined.
Figure 5.14: Pressure volume diagram.
184
Figure 5.15: Open pressure diagram pressure as a function of the crank
angle
A pressure volume diagram, an open pressure diagram and a temperature
angle diagram are shown in Figures 5.14, 5.15, and 5.16 respectively. These
are the basic output diagrams of the simulation program.
Figure 5.16: In-cylinder temperature as a function of the crank angle.
Complementary simulations are performed on the basis of the pressure and
temperature diagrams, allowing further graphical interpretation of the
cycle. This includes rate of pressure rise as a function of the crank angle,
185
as shown in Figure 5.17, rate of energy release as a function of the crank
angle as shown in Figure 5.18, and rate of energy received from the
combustion chamber as a function of the crank angle, as shown in Figure
5.19. These are all diagrams related to heat transfer and provides cycle
analysis information essential to investigate hydrogen combustion.
Figure 5.17: Rate of pressure rise as a function of the crank angle.
Figure 5.18: Rate of energy release as a function of the crank angle.
186
Figure 5.19: Rate of energy transfer fuel combustion and combustion
chamber walls.
Another set of diagrams which can be drawn are related with the air flow
through the engine and valves settings. These diagrams are illustrated in
Figures 5.20, 5.21, and 5.22.
Figure 5.20: Inlet and exhaust valve areas as a function of crank angle.
187
Figure 5.21: Variation of induced mass of air and exhaust gases as a
function of rank angle.
Figure 5.22: Cylinder air mass flow rate and its variation with the crank
angle.
188
5.5 Hydrogen injectors modelling
The next sections present the design of the hydrogen injectors used for
HCCI and DIH2 engine operation. Since each injector has a different design
and requirements, their development work is presented separately. The
injection control system is also presented.
5.5.1 HCCI injector design considerations
The design of the injector for the homogeneous charge compression
ignition mode of operation is simpler than that of the DIH2 injector, as the
pressures, temperatures and time available for injection are less restrictive
than for the direct injection injector.
The HCCI injector used first was a pulse width modulated, two way,
normally closed, solenoid valve, as shown in Figure 5.23. It is made from
corrosion resistant materials and used to inject gaseous hydrogen into the
engine intake air manifold. Pulse width modulation allows fuel quantity to
be controlled with an opening and closing timing precision of +/-25
microseconds.
Figure 5.23: HCCI Hydrogen injection valve.
The valve assembly is composed of the valve body, which holds the solenoid
armature, ball poppet and seat, as shown in Figure 5.23.
189
With the solenoid de-energised, the supply pressure, assisted by a spring,
forces the solenoid ball poppet on its seat, barring gas flow. When the
solenoid is energised, the ball poppet is lifted off the seat and held against
the stop. Gas then passes through the valve seat and outlet port of the
injector.
Figure 5.24: Cross section of the solenoid valve.
Figure 5.25 shows the injection valve fitted on the engine air inlet manifold
where a pressure gauge and thermocouple were fitted to monitor the
hydrogen pressure and temperature.
Digital
pressure
gauge
Injection
valve
Hydrogen
supply tube
Figure 5.25: HCCI injection valve fitted on the engine.
190
As referred before, the electronically controlled fuel injection valve is used
also as a fuel-metering device.
Gaseous fuelling consistency between
injectors and cycles is just as important as with direct injection injectors.
The steady gaseous flow rate is defined by the following expression:
Q steady = C d  A  Vt ,
(5.86)
Cd is the coefficient of discharge, A is the area of throat at valve seat (m2),
B
and Vt is the velocity at the throat (m/s).
B
For sonic flow, the velocity of gas through the orifice is given by:
Vt =
γ  Pt
,
ρt
(5.87)
γ is the ratio of specific heats, P is the pressure at the throat (Pa), and ρ
t
B
is the density at the throat.
Therefore, the steady mass flow rate is given by
Qsteady =
K  A  P1
,
T1
(5.88)
P1 is absolute inlet pressure (Pa), T1 is absolute inlet temperature (K), and
B
B
K is a constant given by injector manufacturer (0.214),
provided that
P2
P1
≤r ,
(5.89)
P2 is absolute outlet pressure (Pa) and R is the critical pressure ratio,
B
and
γ
 2 γ 1

.
r = 
 γ +1 
(5.90)
Therefore, the ratio of manifold absolute pressure to absolute supply
pressure cannot exceed the critical pressure ratio to maintain the sonic
191
flow characteristic. The time delay of this valve is 3.0 ms to open and 2.0
ms to close.
5.5.2 DIH 2 injector design
The production of a hydraulic high speed injection system for engine speeds
above 1000 rpm is problematic because of the dynamic response of the
injector, the lubrication of the moving parts, and the prevention of
hydrogen leakage.
Assuming that hydrogen behaves as an ideal gas and flows isentropically
between the hydrogen manifold and the nozzle hole exit, then there is no
energy loss during the flow process. Assuming also that the enthalpy is
dependent only on the temperature, using the relationship between the
specific heat at constant pressure and the specific heat ratio, the following
energy equation can be written:
1 2
k pe
k po
ue +
=
,
2
k  1 ρe k  1 ρo
(5.91)
p is pressure (Pa),  is density (kg/m3), u is internal energy (kJ/kg). and k is
P
P
specific heat ratio.
192
p e ; T e ;  e ;u e  0
R e s e r vo i r
p 0 ;T 0 ;  0 ;u 0 =0
Ae
Figure 5.26: Isentropic jet development of hydrogen injection.
Figure 5.26 shows a schematic diagram of the hydrogen manifold and
nozzle exit. The subscripts “e” and “o” refer to the conditions at the exit
of the nozzle hole, and inside the hydrogen manifold respectively.
The flow velocity and the mass flow rate at the nozzle can be derived
based on figure 5.26:

 k  1 

 p e   k  
k po
ue = 2
1   

k  1 ρo   p o 




me   eue Ae  Ae
2
 k 1 



k




k
pe 
pe  k  
po  o   1   
2


k 1
 po    p o 


193
(5.92)
(5.93)
As the flow velocity at the exit of the nozzle hole approaches the speed of
sound, in compressible flow theory under choked flow conditions (Yunus A.
et al. 1998), the critical pressure, critical density, critical velocity and
critical temperature are given respectively:
k
 2 k 1
p = po 

 k +1 
1
1
 p 
 2  k +1
ρ = ρo   k = ρo 

1
p
k
+


 o
u= k
p
=
ρ
2k p o
k + 1 ρo
 2 
T = To 

 k +1 
(5.94)
(5.95)
(5.96)
(5.97)
Neglecting the energy losses due to viscosity at the nozzle hole, the
maximum mass flow rate is expressed by equation (5.98) as follows:
1
k +1
 po  2  2  2k  1
m   0 uAe = ρo Ae 
k  

 ρo   k + 1 
(5.98)
In practice, because there are friction losses due to viscosity, at the entry
of the nozzle, the flow velocity is smaller than the critical velocity, and as
a consequence, the discharge coefficient is also less than one, varying
194
according to the geometry of the nozzle hole. Therefore, it is necessary to
investigate the discharge coefficients for varying flow conditions.
5.5.3 Under-expanded gas flow in the proximity of a nozzle
hole
The pressure at the nozzle exit is higher than the cylinder pressure, causing
the hydrogen to expand as it is injected. Since hydrogen flows near the
nozzle exit as if it emanated from a nozzle diameter hole larger source
than the actual nozzle (Tsujimura et al., 2003), the nozzle is referred as
having a pseudo-diameter. Figure 5.27 shows a schematic diagram of the
under-expanded jet behaviour just downstream of the nozzle exit. In the
immediate region downstream of the nozzle exit, the over pressure causes
the flow to accelerate and expand, generating expansion waves, that form
a barrel shaped shock profile. In the hydrogen barrel shaped shock volume,
the flow reaches a supersonic speed, being capped by a so called Mach
disc, that has several times the diameter of the orifice. It can be assumed
that the hydrogen pressure drops to the ambient gas pressure at the Mach
disc and the ambient gas cannot be entrained into the barrel shaped shock
volume.
195
Figure 5.27: Schematic diagram of the under-expanded jet behaviour at the
nozzle hole exit (Ewan et al., 1986).
Table 5.1: The under-expanded flow equations.
No zzle exit
Pr essure
p*
D ensi t y
ρ
Tem p era tu re
T*
V e l o cit y
u*
Mach number
1
p * <p<p a
P
P
P
P
P
P
Barrel shaped shock
P
P
B
B
Mach disc
p M D =p a
B
B
B
B
ρ <<
ρMD = p0 / RoTMD
------
TMD = To
>>u *
u MD = kRoTMD
P
P
>>1
1
In Table 5.1 the equations for under-expanded flow are summarised.
Therefore the mass flow rate at the Mach disc is given by:
m =
since
π
d 2 ρ MDU MD ,
4 MD
P
U MD  K  MD
  MD
196

 .

(5.99)
(5.100)

m
Therefore
P
2
 MD  MD
d DN
4
  MD


 ,

(5.101)
the subscript MD stands for Mach Disc.
According to (Ewan et al., 1986) flow conditions at the nozzle hole exit are
related with those at the Mach Disc as follows:
 p 
π
m = d 2 C D ρo  k o 
e
4
 ρo 
 d MD

 de
1/ 2
2

p T 
 = C D o  MD 
p a  TD 

k  1
 2  2k  1


,
 k +1 
1/ 2
k  1
 2  2k  1


,
 k +1 
(5.102)
(5.103)
the subscript a refers to the ambient conditions inside the injection
volume.
Assuming that the temperature at the reservoir is approximately the
temperature at the Mach Disc, the diameter of the Mach disc is given by:
d MD
1
1

2

po  2  2 
= d c C D

  .
pa  k +1  



197
(5.104)
The direct injection of hydrogen into a diesel engine cylinder is a process
different compared to the injection of a diesel liquid fuel since there is no
evaporation and the inertia of the particles or droplets does not have the
same magnitude, the momentum theory must be modified for hydrogen.
This fact of the particles low inertia affects the penetration of the jet into
the cylinder volume. Also, the injection process of a gaseous fuel is related
to the auto ignition and consequent combustion. The injection of a highly
pressurised jet into a pressurised atmosphere has been investigated (Ewan
et al., 1996), but there is a limited amount of data available, therefore a
modified momentum theory is presented to derive a gas jet penetration
correlation.
Momentum theory applied to the jet penetration is based on momentum
conservation, i.e., the jet entering the compressed atmosphere of the
cylinder, changes its momentum due to the cylinder pressure, also called
ambient atmosphere. Momentum theory (Wakuri et al., 1960) is often used
to analyse the spray penetration, but for application to gaseous fuels, some
modifications are needed which require some assumptions to be made.
These are: (a) the spray dispersion angle is constant during its development
process, (b) the velocity differences between the fuel and the ambient gas
entrained into the spray are negligible, (c) the momentum at the nozzle
exit is constant up to the spray tip, and (d) the velocity profile of the spray
is uniform. Assuming that hydrogen is injected through the Mach disc, as
shown in Figure 5.28, the spray correlation x’ based on momentum theory
is modified to estimate the jet penetration as follows:
198
Figure 5.28: Schematic diagram of the jet development model
(Tsujimura T. et al., 2003).
k  1 



p
2


x ' = a ' C p k 0 
 2k  1 
ρa  k + 1 




0.25
 d et 


 tanθ 
0.5
+ LBARREL ,
(5.105)
(Tsujimura T. et al., 2003)
a´ is the experimental constant,  is the jet dispersion angle (rad) , and
LBARREL is the distance between the actual nozzle exit and the Mach disc
B
(m).
The length of the barrel shaped shock profile can be represented by means
of the following equation (Ewan et al. 1986):
LBARREL = 0.77d e + 0.068d e1.35
k
p f  2 k 1
.


pa  k +1 
(5.106)
Figure 5.30 presents a schematic diagram of the jet development model.
199
Half Jet Dispersion Angle  [deg.]
15
96
po[MPa]=8; d e[mm]=1.0; o[k g/m3]=11.25
90
14
84
13
78
12
72
11
66
10
60
9
54
8
48
7
42
6
36
5
30
4
24
3
18
2
12
1
6
0
0.2
Jet Penetration X' [mm]
16
0
0.28
0.36
0.44
0.52
0.6
0.68
0.76
0.84
0.92
1
Time After Start of Injection [ms]
Figure 5.29: Profile of jet penetration and half jet dispersion angle for an
orifice with 1.0 mm diameter derived from experimental data (Tsujimura
et al., 2003; Shao et al., 2003).
As a consequence of the presented theory, one can say that the ambient
density increase, results in a decrease of the spray penetration, as shown in
Figure 5.29, and that the hole diameter has a significant influence over the
jet penetration volume just after the start of injection. According to
Rottengrubber et al. (2004), the number of holes in the hydrogen injector
does not significantly affect behaviour of the hydrogen jet, and this was
confirmed by using only one hole injector nozzle.
5.5.4 Injector hydraulic actuation modelling
With the objective of understanding the dynamics of the hydrogen
hydraulic injector, a dynamic simulation model was created using Simulink.
With this dynamic simulation model, the variables that govern the injector
dynamics were identified, enabling injector design decisions to be
made.The injector speed of response is influenced by two main injector
200
components, the hydraulic solenoid valve and the actuator piston. The
forces that act on the actuator piston are shown in Figure 5.30.
Kp
A
x
Pc
PS
a
Figure 5.30: Top - Injector hydraulic actuator free body diagram. Bottom
Forces acting on the needle
201
Therefore, the Newton second law of motion in vertical direction equation
is given by
1
dv
=
(FS  FH  Fhyd ) ,
2
dt mt
(5.107)
mt is the mass of the needle actuator group (kg) and
B
FS = K  X ,
(5.108)
K is the spring elastic constant (N/m);
F H =P H ASACK ,
2
(5.109)
2
PH2 is the hydrogen supply pressure (Pa) and ASACK is the lateral area of the
B
B
cone inside the sack volume (m2); and
P
P
Fhyd = (A  a)  PHYD ,
(5.110)
A is the area of the piston (m2), a is the cross section area of the piston rod
P
P
(m2), and PHYD is the hydraulic oil pressure (Pa).
P
P
B
Applying the momentum equation to the piston and spring movement
depicted at figure 5.32, the speed of response of the injector can be
represented by the following equation:
V t  =
t =t




1
K  X   PH 2  ASACK   A  a PHYD  dt ,
m t=1
(5.111)
V is the velocity of the needle actuator group (m/s) and X is the vertical
displacement (m).
It can be concluded from this analysis that to increase the speed of
response of the injector it is required to reduce the mass of the actuator
assembly, increase the spring stiffness for a fast closing, and increase the
hydraulic pressure to increase the opening response.
To model the hydraulic poppet valve as illustrated in figure 5.33, it is
necessary to consider the armature response as a function of the solenoid
force, as well as the effect of hydraulic and spring forces.
202
Flux Pa th
A i r Ga p
coil
coil
A rmatu re
Exhaust
Cont rol
Pr ess ur e
Pc
Suppl y
p re ss ure P s
Figure 5.31: Cross section of solenoid actuated hydraulic valve.
Figure 5.31 illustrates the cross section view of a typical solenoid valve
which will be used to derive its magnetic circuit equations. The enclosure
armature and pole piece are made of mild steel, and the coil is wound
around the armature/pole axis. With no current, the internal spring forces
press the armature and ball downwards, against the hydraulic force. This
blocks the supply pressure Ps, and opens a path from control pressure to
exhaust. When the solenoid is energized, the armature and pole come
together and the ball opens the supply port and blocks the exhaust port.
According to Faraday´s law, the magnetic flux can be determined,
assuming eddy currents and leakage flux are negligible. Therefore
φ=
v sol  iR
,
N
203
(5.112)
φ is magnetic flux (Wb); vsol is solenoid voltage (V); i is solenoid current
B
(A); R is coil resistance (Z); and N is number of turns.
The magneto motive force required to develop this flux is decomposed into
two components; one for steel and another for the air gap. Although the
majority of the induction is concentrated at the air gap, the nonlinear
properties of the steel components, such as saturation, and hysteresis, may
limit the solenoid valve performance.
Therefore the magneto motive force can be expressed as follows:
MMF = MMFair + MMFsteel
B
B
B
MMFair= Hair x g
B
B
B
(5.114)
B
MMFsteel = Hsteel x Lsteel
B
B
B
B
(5.113)
B
B
(5.115)
B
MMF is the magneto motive force (N); H is the magnetic field intensity
(A/m); g is the length of the air gap (m); and Lsteel is the magnetic circuit
B
length in steel (m).
Within the steel, the flux density B is a nonlinear function of H, dependent
upon the material properties. Assuming that the steel path has a uniform
area A, which relates to φ , B and the air gap, the flux density can be
calculated as
B=
φ
= μ 0 × H air ,
A
(5.116)
A is the cross section area in the air gap (m2) and μ0 is the magnetic
P
P
permeability of air (H/m).
The solenoid force, Fsol, and respective current, i, are given by:
B
B
2
F sol =
0 .5× B × A
, and
μ0
i=
MMF
.
N
(5.117)
(5.118)
The armature response to the solenoid force, as well as to the hydraulic
and spring forces, is given by
204
mX '' =F sol +A0 P s − K s X − C v X ' ,
(5.119)
X is armature position (m), M is mass (kg), A0 is the supply orifice area
B
2
(m ),Ps is the supply pressure (Pa), Ks is the spring constant (N/m), and Cv is
P
P
B
B
B
the damping rate.
The net oil flow directed from the valve to the actuator is equal to the
difference between the supply flow and the exhaust flow which causes the
valve needle to lift from its seat.
q net =q s − qexh
(5.120)
q S = K 0 A0 sign(PS  PC ) PS  PC  , for X>0
(5.121)
q S = 0, for X=0
(5.122)
qexh = K 0 A0 PC for X>balltravel
(5.123)
q exh= 0, for X=balltravel
(5.124)
PC is the control pressure (Pa) and K0 is the flow coefficient.
The actuating assembly moves the piston against the spring as a function of
the control pressure developed behind it. Neglecting the leakage, the
variation of the control pressure as a function of time can be expressed as:
PC' =


β
q net  X P' AP ,
V
(5.125)
V=XP AP is fluid volume (m3), XP is actuator piston position (m); and AP is
B
B
B
P
P
B
B
B
actuator piston area (m2).
The equation of motion is dominated by the relatively large hydraulic and
spring forces, neglecting the existence of leakages and according with the
equation of motion (5.126):
M P X ''P =PC A P − K SP X P ,
(5.126)
MP is the net actuator mass (kg) and KSP is the spring constant (N/m).
B
B
205
5.5.4.1 Assumptions for the hydraulic injector simulation
In the absence of detailed information on the mechanical and electrical
properties of the solenoid valve, the basic timing information from the
manufacturer was used as the basis of the solenoid model. Therefore, it
was assumed that, time constants for the valve to reach the fully open
position was 3 ms and to reach fully closed position was 2ms. The
displacement of the armature between fully open and fully closed is
assumed to be proportional to the elapsed time.
The hydraulic oil pressure upstream of the solenoid valve and the injector
is assumed to be constant since an accumulator is included in the hydraulic
circuit. The injector control pressure, Pc, is assumed to range between the
upstream hydraulic pressure, Phyd, when the solenoid valve is fully open and
atmospheric pressure, Patm, when the valve is fully closed.
V a lv e clo sed , X=X1 I n j ec t o r N e edl e X = X1
N e edl e d isp lace men t, X V a lv e s e at Effective a re a – t ru n c at ed c on e Figure 5.32: Definitions of angles of passages of the injector nozzle
The area of passage for the injector nozzle, as illustrated in Figure 5.32,
can be calculated by:
206




2
2
Ao = π  R R 2 + Rtanφ   R  Xsinφsinφ  R  Xsinφsinφ  + Rtanφ  Xsin 2 φ 


(5.127)
R is nozzle hole radius (m), θ is needle tip angle (rad) and Φ is θ/2.
The maximum area is the nozzle area, given by:
Ao Max =  R2
For non-choked flow, mass flow through the orifice is given by

m = K nv AP1
2M  γ   P2

 
RT  γ  1   P1




2/γ
P
  2
 P1



( γ+1)/γ

.

(5.128)
For choked flow condition we have the mass flow expressed by:
m = K nv AP1
γM
RT
 2γ 


 γ 1
( γ+1)/(γ 1)
,
(5.129)
Knv is the flow coefficient, A is the orifice area (m2), P1 is the H2 upstream
B
B
B
B
B
B
pressure (Pa), P2 is the in-cylinder pressure (Pa) and T is the upstream H2
B
B
temperature (K).
Choked flow occurs when ratio P1/P2 exceeds the critical pressure ratio Pc,
B
B
given by
 γ +1 
Pc = 

 2 
γ /(  1)
(5.130)
5.5.5 DIH2 injector dynamic simulation
A direct injection injector for gaseous fuels has various objectives to meet:
a) Accurate metering capability. This is the characteristic of the injector to
deliver consistently the same amount of hydrogen, when actuated with the
same input. This allows the fuel delivered to the engine to be accurately
controlled by the Electronic Control Unit, using a control pulse width
signal.
207
B
b) Good dynamic response. This is characterised by its capability to cope
with injection speed and consistency of opening and closing times.
c) The injector should be able to inject the hydrogen at the appropriate
pressure.
d) The injector should not have any leakage.
e) The injector should be constructed with materials compatible with
hydrogen.
f) The injector moving components should be adequately lubricated.
To achieve the three first requirements, a simulation model using Simulink
was developed using the equations presented above. The structure of the
simulation model is shown in Figure 5.33.
Figure 5.33: Block diagram of hydraulic injector model.
The model is subdivided into the following sub model blocs: orifice area
calculation; inlet valve flow; exhaust valve flow; valve flow; total force
208
calculation; solenoid; critical pressure ratio calculator; choked flow
calculation; subsonic flow calculation; mass flow calculation; hydraulic
actuator; and cylinder pressurization.
The model parameters used are shown in Figure 5.34, and the structure of
the sub models is shown in Figures 5.35-5.44.
Figure 5.34: Simulation model parameters.
Cylinder pressurization
Figure 5.35: Cylinder pressurization.
209
Figure 5.36: Hydraulic actuator sub model.
Flow calculation
Figure 5.37: Choked flow sub model.
210
Critical pressure calculation
Figure 5.38: Critical flow calculation block.
Subsonic flow calculation
Figure 5.39: Subsonic flow model.
211
Figure 5.40: Solenoid sub model.
Resultant force sub model
Figure 5.41: Resultant force calculation sub model.
212
Overall valve flow model
Figure 5.42: Overall valve flow model.
Inlet valve flow
Figure 5.43: Inlet valve flow sub model.
213
Injection valve sub model
Figure 5.44: Injection (exhaust) valve sub model.
5.6 Summary
The results obtained from a simulation program depend on the suitability of
the sub-models used, and assumptions made during the modelling phase.
The accuracy of the model is a trade-off between time and effort;
therefore it is a balance between effort and accuracy of the results to
select the most relevant sub-models. The developed simulation program
was oriented towards the research needs, therefore considered the most
important sub-models, such as valve modelling, cylinder charge mass
variation, Cp variation with temperature, heat release, heat losses,
injection timing adjustment, ambient conditions, and a good graphic
display capability. While a number of sub-models are generic and were
easily adopted from existing models of reciprocating engines, there is a
great deal of controversial models used for hydrogen engines, in particular
in what concerns the adequacy of certain standard approaches to model
heat release and heat losses in hydrogen fuelled engines. It was identified
that despite the connection between some sub-models of the air flow
calculation and valve lift polynomials revealed some discontinuities, it was
very useful as the influence of those discontinuities was not relevant. The
214
program was found suitable and fundamental for the research carried out,
revealing how the engine would perform under operating conditions that
could not be achieved with the actual test engine.
This chapter presented the main sub-models of the simulation program, as
well as the simulation program structure and functionalities. All the
simulation results of this research were based on the sub-models
introduced, and their results and conclusions are presented into Chapter 6
“Performance analysis through simulation”.
As the simulation work developed during this research is divided into two
different systems, engine modelling and injection system modelling, the
chapter presents these two systems models separately. The first modelled
with the Matlab programming language, the second modelled using Simulink
simulation block tools. In what concerns the engine
models, further
detailed investigations based on experimental work, can lead to even
better results through model block improvements, being a recommendation
for future development, in particular on the cylinder charge ignition, flame
development and injection optimization knowledge .
Most of the engine sub-models used are based on commonly recommended
modelling approaches and the engine heat loss model was based on the
behaviour of constant volume combustion rather than constant pressure
combustion as it was found to be more appropriate for hydrogen due to its
combustion behaviour. The injector models introduced in this chapter were
used to perform dynamic simulations studies and their results are
presented on next chapter. The model of the injector for direct injection
of hydrogen is a valuable design tool, and revealed the power of the
simulation as the results obtained from practice coincided with the
expected results from simulation in particular for setting up of the DIH2
injection system.
215
Chapter 6
Performance analysis through
simulation
« There is no such thing as a failed experiment, only experiments with
unexpected outcomes. »
Buckminster Fuller
This chapter describes the HCCI and DIH2 engines performance analysis and
possible design improvements using the developed simulation program
produced from the mathematical models presented in Chapter 5. By using
the simulation program, the limitations of the test engine in relation to
design and operating variables could be studied. In this chapter, a DIH2
injector dynamic simulation program is also presented. Simulation results
and possible improvements of the engine and injector operation are
presented and discussed. For the two modes of engine operation, the
research was oriented to study the following: The effect of injection timing
and pulsed injection on the control of the rate of pressure rise (RPR) and
angle of ignition; the effect of equivalence ratio, compression ratio, air
inlet temperature and pressure on engine performance; and the effect of
engine speed on the ignition and combustion processes. For the DIH2
injector, the simulation study was oriented towards identifying the
limitations of the prototype DIH2 injector.
6.1 Hydrogen HCCI model analyses
Data from the test engine was used for the model input parameters. The
parameters were divided into geometric and operational.
The geometric parameters included: stroke, bore, connecting rod length,
exhaust valve timing, exhaust valve inner and outer diameters, inlet valve
217
timing, inlet valve inner and outer diameters, compression ratio, clearance
volume, speed, volumetric efficiency, inlet valve dwell angle, exhaust
valve dwell angle, piston head area, piston head ratio, and piston off-set .
The operational parameters used were: exhaust gas pressure, exhaust gas
temperature,
cylinder
wall,
piston
head
and
cylinder
head
wall
temperatures, start of injection, duration of injection, fuelling rate, and
lower fuel heating value.
6.1.1 Validation and evaluation of the HCCI model
The HCCI model program was validated by comparing simulated and test
data collected at a number of different engine loads. For a set constant
fuel rate, injection timing, engine speed, and ambient air temperature and
pressure, the following data was recorded: cylinder pressure data, rate of
pressure rise, ignition timing, exhaust gas temperature, and power output
from the engine.
The output variables that served to compare the test engine with the
simulation model were: maximum combustion pressure (P max); angle of
b
maximum combustion pressure (APmax); Iignition pressure (P ign); angle of
B
B
b
ignition (A ign); exhaust gas temperature (TExh); the maximum rate of
b
B
B
pressure rise (MRPR); rate of pressure rise (RPR); engine thermal efficiency;
indicated mean effective pressure (IMEP); and brake power (P b).
b
To assess the accuracy of the simulation model for HCCI mode of operation,
the engine was tested with known ambient and operational conditions that
were replicated in the simulation model.
The input variables for the simulation model were: ambient air
temperature (Tainlet), ambient air pressure (P ainlet), hydrogen mass fuel rate
B
b
(mH2), and engine speed (N e).
B
b
The engine was warmed up and the load and temperatures stabilised for
ten minutes at each load, prior to recording test data.
218
The data was taken from the test engine using the developed data
acquisition system. Six data sets were recorded at each operating point,
and the data was averaged and filtered to remove noise within Labview.
For the tests the engine load was 5.6kW, which was the maximum load
achieved in HCCI mode, and the operating conditions were:
Tair
B
inlet
B
B
= 90ºC, Pa=101.5 kPa; mH2=7.698g/min ; N e=2250rpm. Table 6.1
B
B
B
B
b
B
shows the engine test data recorded and the simulation data at this
operating point and conditions.
Table 6.1: Comparison between simulated and measured results for HCCI
mode of operation.
Engine
Variable
Test engine
simulation
Error
model
P m a x (bar)
84.7
86.4
2.0%
AP m a x (CAº)
354.4º
358º
3.6º
IMEP (bar)
3.6
3.3
7.0%
63.0
35.6
43.5%
8.1
9.2
12.2%
P i g n (bar)
33.0
31.2
5.4%
A i g n (CAº)
348
352
4
T e x h (ºC)
418
469
10.8%
P e x h (bar)
2.0
2.2
10%
36.4
39.5
7.9%
5.6
6.1
7.8%
MRPR (bar/CAº)
RPR (bar/CAº)
Thermal
efficiency (%)
Power (kW)
219
120
Simulated Pressure (bar)
100
Simulated dP_dTheta (bar/deg)
Measured In-Cylinder. P.
80
Measured Norm. Deriv. In-Cyl. P.
60
40
20
0
0
100
200
300
400
500
600
700
800
-20
-40
-60
Figure 6.1: Comparison between predicted and measured
pressure traces and their derivatives for the HCCI compression
ignition engine.
It can be seen by the results presented in Table 6.1 and Figure 6.1 that the
RPR and the MRPR have deviations from the average that are not
acceptable. The inaccuracies in the simulated results for the parameters
are because of the problems in accurately modelling the combustion
process. Apart from these parameters, the model output seems to be
sufficiently accurate for the purpose of this research work.
6.1.2 Hydrogen HCCI engine operation analysis
As identified during the tests, the HCCI mode of operation can reach
thermal efficiencies in the order of 50%, while the emissions are kept to a
minimum. Nevertheless it has certain problems that require a design effort
if the hydrogen HCCI engine is to become a commercial option in the
future.
Problems like difficulty to start, engine control and load limit will be
addressed in the following sections.
220
Due to physical and economic limitations of the experimental engine and
testing, simulation was used to explore possible improvements.
In the following section, the control of the ignition angle is characterised
and three different systems that can contribute to effecting control of
ignition are proposed. The control of the ignition angle through the
addition of a second fuel with different ignition properties (see for example
Olsson et al, 2000) does not seem practical.
6.1.3 Problems associated with HCCI operation
6.1.3.1 Hydrogen slip during the valve overlap period
Hydrogen slip results in inefficient hydrogen use and the presence of
unburnt hydrogen in the exhaust gases. The maximum power developed by
the engine is limited by the amount of hydrogen induced and the amount of
slip.
This problem can only be slightly improved for HCCI operation by reducing
the valve overlap period, i.e. the period when both inlet and exhaust
valves are open. Decreasing the valve overlap period increases the amount
of residual gases trapped inside the cylinder, which can lead to a drop in
engine power and poor combustion. A small amount of this so called
internal exhaust gas recirculation can have the advantage of reducing the
NOx emissions. However, this should not be above 30% of the exhaust mass
flow rate, otherwise the penalty in specific fuel consumption and
exaggerated displacement of oxygen in the cylinder becomes too high.
221
2,79
350
2,785
340
2,78
330
2,775
320
2,77
70
72
74
76
78
80
82
Exhaust gas temperature (ºC)
IMEP (bar)
2200 RPM, Ta = 90 ºC, mH2 = 6.02 g/min
310
84
Exhaust gas valve opening angle BBDC (degrees crank angle)
Figure 6.2: Simulated exhaust gas internal recirculation by
reduction of valve overlap period.
The simulation program was run for different degrees of valve overlap to
study the effects on engine performance. Figure 6.2 shows the results for a
hydrogen flow rate of 6.02 g/min and an air inlet temperature of 90ºC.
There are two obvious effects resulting from the reduction of the valve
overlap period. One is a decrease in the indicated mean effective pressure,
therefore a decrease in power and engine thermal efficiency, and the other
is a substantial increase of the exhaust gas temperature.
6.1.3.2 High rates of pressure rise.
As mentioned before, the cylinder charge ignition angle of the HCCI engine
is dependent on when the ignition temperature is reached during the
compression stroke, and it is therefore influenced by the air inlet
temperature. Figure 6.3 shows the angle of ignition for different inlet air
222
temperatures for an engine running at 2200 rpm and with a hydrogen flow
rate of 6.2 g/min.
370
2200 rpmy=387.5-0.2071x
 IGN (º CA)
368
366
364
362
90
95
100
105
110
115
120
Tair inlet(ºC)
Figure 6.3: Simulated angle of ignition for different air inlet temperatures.
60
MRPR (bar/ºCA)
55
T air inlet = 90ºC
50
45
T air inlet = 110ºC
40
35
Tair inlet = 100ºC,
30
2
3
4

5
6
7
8
9
10
Figure 6.4: Dependence of the MRPR as a function of Tair inlet and λ.
223
Figure 6.4 shows the rate of pressure rise for different air inlet
temperatures and the relationship with excess air ratio λ and cylinder inlet
temperature Tair inlet. This is an important engine operating parameter, since
it indicates the level of piston force transferred to the engine crankshaft. It
can be seen from Figure 6.4 that the rate of pressure rise depends heavily
on the excess air ratio and to some degree on the temperature of the air at
the cylinder inlet.The maximum rate of pressure rise for HCCI engine
operation is more than the double that of diesel operation, and above the
maximum recommended 12 bar/º. It is therefore critical that the operating
characteristics of the HCCI engine are taken into account in the mechanical
design of the engine.
6.1.3.3 Power limitation of the HCCI engine
The power output of the HCCI engine is limited by the volume of hydrogen
that can be induced per stroke. Theoretically, for a stoichiometric
hydrogen-air mixture, the (gaseous) hydrogen makes up approximately 30%
of the cylinder displacement volume. Induction of a larger quantity of
hydrogen using port injection can not increase engine power. Due to the
low density of hydrogen, a stoichiometric hydrogen-air mixture has an
energy content of approximately 83% of that of a gasoline-air mixture. This
reduces the engine power output of a pre-mixed hydrogen engine by
around 30% compared to a gasoline engine.
A possible way to get around this limitation would be the injection of
hydrogen into the cylinder at a pressure slightly higher than the one used
with port injection during the first degrees of crank angle while the piston
is starting the charge compression, in this case it might be considered to
fall in the category of direct injection engine.
224
6.1.4 Possible design improvements using simulation
Following the results presented for HCCI operation, design improvements
are recommended to ensure that the engine can operate reliably. These
are:
a) Stronger piston rings and higher load bearings, to withstand the higher
dynamic shock loads. b) An inlet air heating system utilising the exhaust gas
heat. c) Hydrogen injection control ensuring injection only after exhaust
valve closing, to avoid hydrogen slip. d) Control of the temperature of the
air at the engine inlet, the excess air ratio, and the injection duty cycle,
using engine knock as a feedback signal. e) Flame screens fitted on the air
manifold, and a connecting pipe fitted between the crank case and the
inlet manifold, to improve safety by removing blow-by hydrogen.
Results from the test engine and simulation study show that the HCCI
operation has a number of problems that need to be addressed, namely:

Poor combustion control.

Maximum power limited by the amount of hydrogen induced.

Hydrogen slip during the valve overlap period.

High rates of pressure rise.

Safety issues, such as air inlet manifold backfires.
6.1.4.1 Combustion control and dependence of the inlet air
temperature
These problems are characterised by the absence of any timed ignition
mechanism able to ignite the cylinder charge at a pre-defined crank angle.
The ignition timing on the HCCI engine is dependent on the temperature of
the cylinder charge; therefore the test engine performance is sensitive to
the air temperature entering the cylinder. Increasing inlet air temperature
leads to advanced ignition, and vice versa, as shown in Figure 6.3.
225
This problem can be solved by increasing the engine compression ratio to
an extent that the self-ignition temperature of the hydrogen-air charge is
reached at a certain desired angle without heating of the inlet air. If this
solution is implemented, an intake air heating system may not be
necessary, and the engine will be able to start without pre-heating. If a
variable compression ratio system is implemented, this can be adapted to
control ignition timing based on the actual local temperature of the
ambient air.
Another option involves the use of two different fuels; in this case,
hydrogen and another fuel with a different ignition temperature. Varying
the composition of the fuels mixture can allow the ignition angle to be
controlled (Olsson et al, 2000).
Tair inlet = -18.64 + 363.8 e -0.07408 CR
CYLINDER AIR INLET TEMPERATURE ºC
85
80
75
70
65
60
55
50
45
40
35
17
18
19
20
21
22
23
24
25
COMPRESSION RATIO
Figure 6.5: Simulated relationship between the minimum cylinder air inlet
temperature required to maintain combustion and the engine compression
ratio.
Figure 6.5 shows the simulated relationship between the compression ratio
and the required inlet temperature for a constant speed of 2200 rpm and
hydrogen mass flow rate of 9g/minute. This relationship shows that a
compression ratio of the order of 25:1 could ensure combustion with a
temperature of the air at the cylinder inlet of 38ºC.
226
Operating an engine with such high compression ratio results in the control
of its combustion process being performed by the quantity of hydrogen
injected in the inlet port i.e. becomes a PWM controlled process,
independent of the air inlet temperature.
However, increasing the compression ratio to 25:1 will require stronger
engine components.
Figure 6.6 shows the effect of inlet air temperature over the indicated
power and mean effective pressure. As expected, an increase in inlet air
temperature results a decrease in the engine power output due to the
reduced density of the cylinder charge.
Figure 6.6: Simulated effect of the air inlet temperature on the IMEP and
indicated power.
227
6.2 Hydrogen direct injection engine model validation
The DIH2 engine model is basically the same as that used for the HCCI
engine, as the engine used for testing is the same. The model shares the
same routines and sub models presented previously.
By selecting the mode DIH2 mode of operation, the DIH2 injection
B
B
parameters are enabled, and the model assumes that the chosen quantity
of hydrogen is being introduced in the cylinder according with the injection
timing and duration set. With respect to the ignition delay and timing, the
model assumes that ignition will take place only when the self ignition
temperature of hydrogen is reached. Therefore it is dependent on the
operational and ambient conditions set as inputs. The DIH2 model program
was validated through a comparison between simulated and test data
collected from a number of simulation runs and comparative engine tests.
Data was acquired for a set constant fuel rate, injection timing, engine
speed, and ambient air temperature and pressure. Cylinder pressure data,
rate of pressure rise, ignition timing, exhaust gas temperature, and power
output were recorded from the engine.
To assess the accuracy of the simulation model for the DIH2 mode of
operation, the engine was tested with known ambient and operational
conditions that were replicated in the simulation model.
The input variables used for the simulation model were: ambient air
temperature, ambient air pressure, hydrogen mass fuel rate, and engine
speed.
The output variables that served to compare the test engine with the
simulation model were: maximum combustion pressure (P max), angle of
b
maximum combustion pressure (APmax), ignition pressure (P ign), angle of
B
b
ignition (A ign), exhaust gas temperature (TExh), rate of maximum pressure
B
b
rise (MPRP), rate of pressure rise (RPR), engine thermal efficiency,
indicated mean effective pressure (IMEP) and brake power (P b).From the
b
DIH2 simulation model other variables are also available, such as the excess
228
air factor (, the maximum combustion temperature (T max), rate of heat
b
release (RHR), air mass flow rate and polytropic index (n).
The engine was warmed up and the temperatures and pressures stabilised
for ten minutes at each load, prior to recording test data.
The data was taken from the test engine using the developed data
acquisition system, as described above.
Six data sets were recorded at each operating point. The data was
averaged and filtered to remove noise within Labview. For the tests
conducted at a load of 6kW, the operating conditions were:
Table 6.2: Comparison between simulated and measured results for the
DIH2 mode of operation at 6.0 kW load.
Tair
B
B
inlet
B
B
= 90ºC, Pa=101.5 kPa; mH2=7.698g/min ; N e=2250rpm.
B
B
B
B
b
B
Engine
Variables
Test engine
simulation
Error
model
Pmax (bar)
78.8
81.7
3.5%
APmax (CAº)
364
369
5º
IMEP (bar)
4.0
4.1
3.0%
MRPR (bar/CAº)
72
28
60%
13.4
14.1
6.0%
40
42
4.7%
A i g n (CAº)
361
366
5º
Texh (ºC)
403
410
1.7%
Pexh (bar)
2.4
2.5
3.0%
39.0
41.0
4.8%
6.0
6.4
5.0%
RPR (bar/CAº)
Pign (bar)
Thermal
efficiency (%)
Power (kW)
229
100
80
DI SIMULATION MODEL VAL Pressure (bar)
DI SIMULATION MODEL VAL dP_dTheta (bar/deg)
TEST DI SIMUL VALIDATION In-Cylinder. P.
60
TEST DI SIMUL VALIDATION Norm. Deriv. In-Cyl. P.
SIMULATION DI VALIDATION Pressure (bar)
40
SIMULATION DI VALIDATION dP_dTheta (bar/deg)
20
0
-20
0
100
200
300
400
500
600
700
-40
-60
Figure 6.7: Comparison between predicted and measured pressure traces
and their derivatives for DIH2 mode operation.
It can be seen from the results presented in Table 6.2 and Figure 6.7 that
only the RPR and the MRPR have deviations that are not acceptable. The
inaccuracies in the simulated results are due to the challenges associated
with accurately modelling the combustion process. Apart from these
parameters, the model output seems to be sufficiently accurate for the
purpose of this research work.
6.2.1 DIH2 engine design and operational analysis
The diesel engine operated with direct injection of hydrogen has a number
of advantages over the other modes of operation tested during this
research. These are a higher power to weight ratio, very low exhaust gas
emissions, no back firing and good control of the combustion process. It
also has some problems that need to be overcome.
Like the HCCI mode of operation, the DIH2 engine depends on the inlet air
B
temperature to ensure combustion. This is particularly important in what
regards ignition delay, affecting essentially the MRPR. The angle at which
the ignition of the cylinder charge takes place during the cycle is not so
230
800
dependent on the air inlet temperature since the hydrogen is injected in
the cylinder only when desired, at any crank angle after the ignition
temperature has been reached. However, the mechanical loads related to
the MRPR must be considered and it is important that measures to control
this are taken.
One method to control high MRPR values is by using pulsed injection and
appropriate inlet valve timing. Unfortunately, the hydrogen injector for
direct injection which was manufactured for this research was not fast
enough to perform pulsed injection. However, pulsed injection and
modified inlet valve timing were investigated using the simulation model of
the engine.
The ignition delay period forms the first phase of the combustion process,
and is dependent on the properties of the hydrogen-air mixture. The
second phase consists of the spread of the flame from the initial point of
ignition to the main body of the cylinder charge. There is a rapid increase
in pressure during this phase and the rate of pressure rise depends to some
extent on the availability of oxygen next to the hydrogen spray, which in
turn depends on the turbulence in the cylinder. Since hydrogen is injected
into the cylinder in its gaseous phase, its mixing with air is extremely fast.
Due to the very large excess of air, oxygen is available for combustion
during this phase throughout the whole cylinder volume. Therefore, as the
engine speed increases the rate of pressure rise also increases, and can
result in engine knock. During the third phase of combustion the fuel burns
as it is injected into the cylinder, giving more controlled combustion than
in phase two. One of the main factors affecting the combustion in the
controlled combustion phase is the gas motion (swirl and squish effects)
which is governed by the shape of the combustion chamber.
231
6.2.2 Control of MRPR and engine optimisation
A number of tests and simulation studies were conducted to understand the
effect of the injection timing and the use of pulsed injection over the
thermodynamic cycle, and how these variables can help to control the
MRPR.
Injection profile and timing
Injection is characterised by its timing, frequency of the injection profile
(pulse), its injection duration, and by the angle at which the injection is
started. These parameters are expressed in terms of the crank angle. The
injection profile can be continuous, i.e. normal injection, being
characterised by a single square-shaped pulse with rapid opening and
closing and approximately constant fuel flow, or pulsed; where the injector
is opened and closed a number of times during the injection angle, thus
controlling the rate at which fuel is entering into the cylinder, and
consequently the rate of heat release.
Unfortunately it was not possible to implement experimentally this
solution, as the injector which was manufactured did not have a
sufficiently fast dynamic response. High frequencies of operation, as well
as accurate control, are essential requirements for pulsed injection.
Therefore, this was investigated using the developed engine model and
simulation program.
The frequency of the injection pulse profile during the injection angle
depends on the engine speed and according to the Nyquist theorem
Richards et al. 1979, the injection model needs to run at a frequency at
least three times the injection frequency, if it is required to operate in the
same time interval as the injection process under study. Another important
aspect that was identified during the simulation study of pulsed injection
was that, for high injection frequencies (typically above 1.0 kHz), if the
232
duty cycle is above 40%, then the pulsed injection approaches the
continuous injection behaviour, therefore not producing any effect.
Each of the injection timing combinations resulted in a different engine
cycle performance, and the simulation program was used to investigate the
most appropriate optimal combination that can meet the objectives
previously stated. In order to isolate the influences of the other engine
parameters, from the injection timing results, all the engine operational
parameters were kept constant.
The injection timing and duration simulation studies were performed for
the conditions shown in Table 6.3:
Table 6.3: Operating parameters of the engine for injection timing and
duration simulation studies.
Engine Parameters
Values
Ambient temperature (ºC)
80
Ambient pressure (kPa)
101.3
Engine speed (RPM)
2200
Exhaust gas back pressure (kPa)
101.3
Cylinder wall temperature (ºC)
250
Piston crown temperature (ºC)
300
Cylinder head temperature (ºC)
280
Angular resolution (CAº)
0.5
Ignition delay (ms)
1.2
Heat rate ( kJ/kg)
18.0
The following tables and figures show the simulation results for continuous
injection, pulsed injection and the optimised injection of hydrogen.
The following limits served as guide lines for the simulation:
233
Maximum combustion pressure of the same order as for diesel oil
operation at corresponding loads;
Maximum rate of pressure rise of the same order as for diesel oil
operation at corresponding loads <8ºca;
Angle of maximum pressure > 5ºca;
Continuous injection simulation
Three simulation studies were performed for continuous injection of
hydrogen. The results are presented in Tables 6.4 to 6.6 and Figures 6.9 to
6.17.
Study 1: Continuous injection, duration 30º, start of injection 25º BTDC
Table 6.4: Engine performance for continuous injection, Study 1.
PBmaxB (bar)
98
α PBmaxB (θº)
4
TBmaxB (ºC)
1826
αTBmaxB (θº)
3
IMEP (bar)
4.50
PBiB (kW)
6.8
RPR (bar/º)
6.11
MRPR (bar/º)
61.86
234
Figure 6.8: Open cycle diagram for Study 1.
Figure 6.9: Rate of change of cylinder pressure for Study 1.
235
Figure 6.10: Rate of energy release diagram for Study 1.
Study 2:
Continuous injection, duration 25º, start of injection 25º BTDC
Table 6.5: Engine performance for continuous injection, Study 2.
PBmaxB (bar)
105
α PBmaxB (θº)
-1
TBmaxB (ºC)
1934
αTBmaxB (θº)
-1
IMEP (bar)
4.53
PBiB (kW)
6.85
RPR (bar/º)
13.73
MRPR (bar/º)
73.46
236
Figure 6.11: Open cycle pressure diagram for Study 2.
Figure 6.12: Rate of change of cylinder pressure diagram for Study 2.
237
Figure 6.13: Rate of change of cylinder pressure for Study 2.
Study 3:
Continuous injection, duration 25º, start of injection 30º BTDC
Table 6.6: Engine performance for continuous injection, Study 3.
PBmaxB (bar)
107
α PBmaxB (θº)
-4
TBmaxB (ºC)
2009
αTBmaxB (θº)
-3
IMEP (bar)
4.62
PBiB (kW)
6.98
RPR (bar/º)
21.4
MRPR (bar/º)
85.31
238
Figure 6.14: Open cycle diagram for Study 3.
Figure: 6.15: Rate of change of cylinder pressure diagram for Study 3.
239
Figure 6.16: Rate of energy release diagram for Study 3.
Pulsed injection simulation
Five simulation studies were performed for the case of pulsed injection of
hydrogen. The results are presented in Tables 6.7 to 6.11 and Figures 6.17
to 6.28.
240
Study 4:
Pulsed injection, duration 30º, start of injection = 25º BTDC
Table 6.7: Engine performance for pulsed injection, Study 4, frequency
10kHz, duty cycle 40%.
PBmaxB (bar)
95
α PBmaxB (θº)
5
TBmaxB (ºC)
1775
αTBmaxB (θº)
3
IMEP (bar)
4.42
PBiB (kW)
6.68
RPR (bar/º)
5.92
MRPR (bar/º)
58.92
Figure 6.17: Open pressure diagram for Study 4.
241
Figure 6.18: Rate of change of cylinder pressure diagram for Study 4.
Figure 6.19: Rate of energy release diagram for Study 4.
242
Study 5:
Pulsed injection, duration 25º, start of injection = 25º BTDC
Table 6.8: Engine performance for pulsed injection, Study 5, frequency
10kHz, duty cycle 40%.
PBmaxB (bar)
104
α PBmaxB (θº)
0
TBmaxB (ºC)
1895
αTBmaxB (θº)
0
IMEP (bar)
4.45
PBiB (kW)
6.73
RPR (bar/º)
13.5
MRPR (bar/º)
72.45
243
Figure 6.20: Open pressure diagram for Study 5.
Figure 6.21: Rate of change of cylinder pressure diagram for Study 5.
244
Figure 6.22: Rate of energy release diagram for Study 5.
Study 6: Pulsed injection, duration 25º, start of injection 30º BTDC
Table 6.9: Engine performance for pulsed injection, Study 6, frequency
10kHz, duty cycle 40%.
PBmaxB (bar)
107
α PBmaxB (θº)
-4
TBmaxB (ºC)
2009
αTBmaxB (θº)
-3
IMEP (bar)
4.62
PBiB (kW)
6.98
RPR (bar/º)
21.34
MRPR (bar/º)
85.31
245
Figure 6.23: Open pressure diagram for Study 6.
Figure 6.24: Rate of change of cylinder pressure diagram for Study 6.
246
Figure 6.25: Rate of energy release for Study 6.
Study 7: Pulsed injection, duration 30º, start of injection 25º BTDC
Table 6.10: Engine performance for pulsed injection, Study 7,
frequency 10kHz, duty cycle 40%.
Pmax (bar)
95
α Pmax (θº)
5
Tmax (ºC)
1785
αTmax (θº)
3
IMEP (bar)
4.379
Pi (kW)
6.62
RPR (bar/º)
5.96
MRPR (bar/º)
58.39
247
Figure 6.26: Open pressure diagram for Study 7.
Figure 6.27: Rate of change of cylinder pressure diagram for Study 7.
248
Figure 6.28: Rate of energy release diagram for Study 7.
Optimisation of injection to minimise MRPR
As already introduced, the MRPR (Maximum Rate of Pressure Rise) has a
strong influence on mechanical loads of some components of the crankshaft
connecting rod mechanism such as bearing shells and gudgeon pins, but it
has a strong influence on the engine controllability. Therefore, the control
of this parameter is of upmost importance to achieve a regular engine
operation. It was identified during this research, that a too fast energy
introduction results in high values of MRPR, therefore the controlled
introduction of hydrogen into the cylinder, was an approach to follow,
being implemented through simulation as a pulsed injection.
249
Study 8: Pulsed injection, duration 33º, start of injection 25º BTDC
Table 6.11: Engine performance for pulsed injection, Study 8, frequency
10kHz, duty cycle 40%.
Pmax (bar)
91
αPmax (θº)
8
Tmax (ºC)
1714
αTmax (θº)
3
IMEP (bar)
4.42
Pi (kW)
6.68
RPR (bar/º)
5.44
MRPR (bar/º)
53
Pmax
I g n it io n
Figure 6.29: Open pressure diagram for Study 8.
250
Pu lsed in je ct io n
a ct io n
Pr ess ur e o f ig nit io n
p ea k = MRP R
Figure 6.30: Rate of change of cylinder pressure diagram for Study 8.
I g n it io n pe ak = M R PR
Pu lsed in je ct io n
effect
Figure 6.31: Rate of energy release diagram for Study 8.
251
By inspection of figures 6.29, to 6.31, it can be seen that the effect of
pulsed injection resulted in a time extended energy input into the cylinder,
controlling to some extent MRPR.
6.2.3 Comparison and conclusions regarding the simulated continuous
and pulsed injection
The results of all injection timing and profile simulation studies are
summarised in Table 6.12, showing the effects of injection angle and
duration.
Continuous injection
Considering the results obtained from the simulation studies, it is possible
to conclude that direct injection for a hydrogen fuelled CI engine should
have the injection duration extended after top dead centre and that it
should be pulsed. This injection arrangement will produce a cooler
combustion, and the angle of maximum pressure will meet the objective of
being 5º after top dead centre. Also the maximum combustion pressure will
be of the same order of magnitude as when the engine is operated with
diesel oil. It can be seen from the pressure diagrams that the operating
cycle approaches the constant volume engine cycle, indicating that high
efficiency can be achieved. As hydrogen CI engines require higher
compression ratios to achieve the self ignition temperature of hydrogen
than standard CI engines, it is expected that even higher cycle efficiencies
can be achieved.
From the heat release diagrams it is possible to see that, as the injection
starts at approximately top dead centre, the release of energy will be
faster resulting in higher pressure, and therefore higher values of RPR. The
use of pulsed injection has no apparent effect on the RPR values.
Injection timing plus pulsed injection
As noticed for the continuous injection, the injection timing is an important
parameter in the control of engine RPR and MRPR. The simulation studies
252
show that the injection of the DIH2 engine should be prolonged after the
engine top dead centre and pulsed, thus resulting in a smooth combustion,
with maximum cylinder temperatures around 1700ºC, and maximum
cylinder pressures within the range experienced with diesel oil operation.
Therefore DIH2 operation will not create excessive thermal or mechanical
loads on engine components, and as the developed peak temperatures are
smaller than for the continuous injection, it is expected that the formation
of thermal NOx is extremely reduced, in particular because the thermal NOx
formation takes place above 1700ºC according to Richard (1980) and the
time the cylinder charge temperature is above 1700ºC is extremely small
when compared to hydrocarbon fuels Heywood (1972).
As can be seen from the figures of rate of energy release (RER) and MRPR
(maximum rate of pressure rise), there is an initial peak of energy release,
and only after that does the pulsed injection have some control. The pulsed
injection control is only effective during the second phase of the
combustion process; there is no evidence of any control during the ignition
process. Therefore, it achieves only partially its main objective of
controlling the MRPR originating from the first instant of injection.
It can be also recognised that with the advance of injection the maximum
combustion temperature and pressure increases above acceptable values,
particularly for the MRPR and RPR. The increase of the maximum
combustion temperature Tmax promotes the formation of thermal NOx.
Comparing the results of the various simulations studies (see Table 6.12), it
can be seen that the optimised injection profile and timing for the test
engine was:

start of injection 25º BTDC, duration of injection 33º, plus pulsed
injection at 5 kHz.
This combination results in the lowest maximum combustion temperature,
lower RPR, lower MRPR, while maintaining engine power output.
253
254
Start of injection
25º
Duration of
injection 30º
Pulsed injection
Start Injection 25º
Duration Injection
33º
Pulsed injection
Optimized
Injection
Start Injection 25º
Start Injection 25º
Start Injection 30º
Duration Injection
30º
Duration Injection
25º
Duration Injection
25º
Pmax (bar)
95
91
95
104
107
αPmax (θº)
5
8
5
0
-4
Tmax (ºC)
1785
1714
1775
1885
2009
αTmax (θº)
3
3
3
0
-3
IMEP (bar)
4.38
4.42
4.42
4.45
4.62
PBi (kW)
6.62
6.68
6.68
6.73
6.98
RPR (bar/º)
5.96
5.44
5.92
13.5
21.34
MRPR (bar/º)
58.39
53.07
58.92
72.45
85.31
Table 6.12: DIH2 engine parameters for different injector profile and timing.
6.3 Effect of valve timing (Miller cycle) on DIH2 engine
performance.
Since MRPR is related to the temperatures of the cylinder and combustion
chamber charge, any mechanism for controlling the temperature of the
combustion chamber components internal to the engine is of interest.
Therefore the effect of valve timing and use of the Miller cycle was
investigated as such a possible mechanism.
The Miller cycle was first proposed in 1947 and is achieved by early inletvalve closure to provide improved cooling before compression so as to
reduce compression work (Miller, 1947). Miller et al. (1957) further
proposed the increase of the boost pressure to compensate for the reduced
inlet stroke duration, thus improving the engine thermal efficiency this
effect is illustrated in figure 6.32.
Figure 6.32 Miller cycle illustrations (Source Mazda Corp.)
As a result of the different inlet valve timing, the Miller cycle has lower
combustion temperatures than a conventional engine cycle.
As a direct consequence of this new valve timing, the end-of-compression
temperatures are lower. Therefore, maximum combustion temperatures
are lower too, resulting in a lower MRPR and NOx formation, which is a
function of the temperatures during combustion, time, and turbulence.
253
For the Miller cycle, the compression stroke of the engine is shorter than
the expansion stroke, allowing the compression ratio and the expansion
ratio to be set independently. According to Wang et al. (2005), there are
three different practical ways of implementing the Miller cycle:
a) Using a rotating valve between the air manifold, and the inlet
valve (on the cylinder head) to control the intake air quantity. This is
called early rotary valve closing (ERVC).
b) Closing the inlet valve before the termination of the suction
stroke. This is called early inlet valve closing (EIVC).
c) Keeping the inlet valve open during a portion of the compression
stroke, thus rejecting part of the charge and reducing the net compression
ratio. This is called late inlet valve closing (LIVC).
6.3.1 Effect of the Miller cycle on the DIH2 engine
The EIVC method to accomplish the Miller cycle was used as it was more
readily simulated. Three different inlet valve timings were investigated:
Miller 1: the inlet valve opens 20º late and closes 20º early.
Miller 2: the inlet valve opens 25º late and closes 25º early.
Miller 3: the inlet valve opens 10º late and closes 10º early.
The thermal efficiency, indicated power, and the MRPR relationship to the
hydrogen fuel rate (load) was plotted for each Miller valve setting and for
the conventional valve setting for the DIH2 engine. The results are shown in
Figures 6.33, 6.34 and 6.35.
254
39
Diesel DIH 2
38,5
Thermal efficiency (%)
38
Miller 1
37,5
Miller 3
37
36,5
Miller 2
36
35,5
35
34,5
7
8
9
10
11
12
13
14
15
16
H2 (g/min)
Figure 6.33: Relationship between thermal efficiency and hydrogen fuel
rate for conventional and Miller cycle inlet valve settings.
14
DIH2
M
M ille
ille r
r 1
3
M ille r 2
Indicated Power (kW)
12
10
8
6
4
7
8
9
10
11
12
13
14
15
16
H2 (g/min)
Figure 6.34: Relationship between indicated power and hydrogen fuel rate
for conventional and Miller cycle inlet valve settings.
255
110
DIH2
100
M ille r 1
M ille r 3
M ille r 2
MRPR (bar/º)
90
80
70
60
50
40
30
7
8
9
10
11
12
13
14
15
16
H2 (g/min.)
Figure 6.35: Relationship between MRPR and hydrogen fuel rate for
conventional and Miller cycle inlet valve settings.
As can be seen from the results presented in Figure 6.35, the Miller cycle
can be used effectively to reduce the MRPR, therefore improving the
performance of a DIH2 engine. The inlet valve setting corresponding with
Miller1 has a significant effect on reducing MRPR whilst maintaining high
engine thermal efficiency in comparison to the conventional valve timings,
as shown in Figure 6.33.
For the Miller cycle, there is a slight reduction in indicated power (Figure
6.34), accompanied by a small reduction in thermal efficiency (Figure
6.33). If a Miller cycle is to be implemented for DIH2 engine operation, a
trade-off between a reduction in indicated power and thermal efficiency
and the benefits of a reduction in MRPR needs to be made.
It was found that it was possible to achieve maximum combustion
temperatures as low as 1463ºC (for 7g/minute of H2). This indicates the
benefits the Miller cycle has in reducing NOx formation and combustion
chamber thermal stress. As an example, for full load (16g/minute) in
standard DIH2 mode, the peak combustion temperature is 2593ºC whereas
for the “Miller 2” cycle it is only 2482 ºC (a difference of 111ºC).
256
6.4 DIH2 injector dynamic simulation
As has been discussed previously, a DIH2 direct injection injector has
various objectives to meet, namely:
a) Accurate metering capability. This is the ability of the injector to
deliver consistently the same amount of hydrogen, when actuated with the
same input. This allows the fuel delivered to the engine to be accurately
controlled.
b) Good dynamic response. This is the ability of the injector to
produce rapid and consistent opening and closing times.
c) The injector should be able to inject the hydrogen at the
appropriate pressure.
d) The injector should not leak.
e) The injector should be constructed with materials compatible
with hydrogen.
f) The injector should be auto lubricated.
To investigate the three first requirements, a simulation model using
Simulink was developed using the equations presented in Chapter 5. The
following paragraphs show the simulated dynamic response of the injector
regarding the effect of inertia of its moving parts, duty cycle, actuation
frequency, hydraulic pressure and static force or pre-load.
6.4.1 Effect of the inertia of the moving parts on the injector dynamic
response
The inertia of the moving parts of the injector was studied to understand
how it influences the dynamic response of the injector. The actuator mass
was lumped together with the mass of the spring, and the mass was varied
from 5 g to 50 g, whilst keeping all other injector parameters constant.
The injector parameters were:
257
Table 6.13: Injector parameters for dynamic simulation.
Hydraulic pressure (bar)
50
H 2 pressure (bar)
200
Static spring force (N)
500
Cylinder pressure (bar)
80
Hydraulic
solenoid valve
Pre
loading
bolt
Spring
Actuator
H 2 inlet
Hydraulic
oil in/out
Nozle
Figure 6.36: DIH2 Injector view.
258
Figure 6.36 shows the arrangement of DIH2 injector components, evidencing
the hydraulic actuator and the three way solenoid valve.
Figure 6.37 and 6.38 show that injector actuator speed does not
significantly change for a variation of the actuator mass between 5 g and
50 g.
Figure 6.37: Injector actuator speed for an actuator and spring mass of 5 g.
259
Figure 6.38: Injector actuator speed for an actuator and spring mass of 50g.
Figures 6.39 and 6.40 show that the injection needle displacement is not
significantly influenced by an increase in actuator mass.
Figure 6.39: Injector needle valve displacement for an actuator and spring
mass of 5 g.
260
Figure 6.40: Injector needle valve displacement for an actuator and spring
mass of 50 g.
6.4.2 Effect of duty cycle on the injector dynamic response
For values of duty cycle lower than 30%, the injector is not effective for
regulation of mass flow. Despite the good speed of response of the
actuator, there is no sufficient time for the flow to develop. As can be seen
from the simulation graphs, the injector does the flow modulation in an
acceptable way down to 10% duty cycle. For values of duty cycle lower
than 10%, the injection flow rate per cycle is not regular, potentially
making the engine run erratically.
Figures 6.41-6.44 show the actuator speed for a hydraulic pressure of 50
bar, H2 pressure of 200 bar, static force of 500 N, cylinder pressure of 80
bar, actuator mass 5 g, period 0.0599 s (2000RPM), with the duty cycle
varying from 5% to 30%.
261
Figure 6.41: Actuator speed for a duty cycle of 5%.
In figure 6.41 can be observed that the injector is not sufficiently fast to
open, not injecting any hydrogen.
Figure 6.42: Actuator speed for a duty cycle of 10%.
262
Figure 6.42, shows that when the duty cycle is increased to 10%, the
injector opens, but it does not have time to complete the injection.
Figure 6.43: Actuator speed for a duty cycle of 20%.
In figure 6.43 the injector is opened for a bigger time, but still insufficient
to complete the injection of hydrogen.
Figure 6.44: Actuator speed for a duty cycle of 30%.
263
Figure 6.44 shows that the injector response is complete, evidencing a good
control of the injection cycles.
Figures 6.45 to 6.49 show the injector mass flow rate for varying duty
cycle, with the other parameters constant.
Figure 6.45 shows mass low rate for a 5% Duty cycle, evidencing the
irregular flow rate as a consequence of the poor response of the injector.
Figures 6.46 through 6.49, show that an acceptable mass flow rate control
is achieved only for Duty cycles above 30%. The precise control of the
engine is a function of the mass flow rate and therefore of the amount of
energy per cycle. To achieve such a control it is required to control
accurately the time the injector is open, which is a function of the speed
response of the hydraulic injector.
Figure 6.45: Injector mass flow rate for a duty cycle of 5%.
264
Figure 6.46: Injector mass flow rate for a duty cycle of 10%.
Figure 6.47: Injector mass flow rate for a duty cycle of 20%.
265
Figure 6.48: Injector mass flow rate for a duty cycle of 30%.
Figure 6.49: Injector mass flow rate for a duty cycle of 50%.
266
6.4.3 Effect of the injector actuation frequency on the dynamic
response
The injector actuation frequency is an important parameter since it is
intimately related with the engine operating speed. It is necessary to
determine the limit of injector speed of operation and how the frequency
of its actuation affects the actuator speed and the mass flow rate of
hydrogen in the engine.
In this study, the actuation frequency was varied, while the remaining
operating variables are kept constant.
If pulsed injection of hydrogen is to be implemented, then the injector
must be capable of operating at very high frequencies.
Figures 6.50-6.59 show the effect of actuator frequency on the actuator
speed and mass flow rate for a hydraulic pressure of 50 bar, H2 pressure of
200 bar, static force of 500 N, cylinder pressure of 80 bar, actuator mass of
5 g, and a duty cycle of 50%.
Figure 6.50: Actuator speed with period of injection 0.024 s (5000 RPM).
267
Figure 6.51: Mass flow rate with period of injection 0.024 s (5000 RPM).
Figure 6.52: Actuator speed with period of injection 0.03 s (4000 RPM).
268
Figure 6.53: Mass flow rate with period of injection 0.03 s (4000 RPM).
Figure 6.54: Actuator speed with period of injection 0.0333 s (3600 RPM).
269
Figure 6.55: Mass flow rate with period of injection 0.0333 s (3600 RPM).
Figure 6.56: Actuator speed with period of injection 0.0428 sec (2800 RPM).
270
Figure 6.57: Mass flow rate with period of injection 0.0428 s (2800 RPM).
Figure 6.58: Actuator speed with period of injection 0.0545 s (2200 RPM).
271
Figure 6.59: Mass flow rate with period of injection 0.0545 s (2200 RPM).
From the simulation results, it can be seen that the hydraulic injector is
limited in terms of its speed of response, to a maximum operating speed of
2200 RPM. For higher speeds, the injector will be not even injecting
hydrogen into the cylinder, this condition is represented in figure 6.50 that
illustrates a “vibration” of the actuator around zero.
The simulated results indicate that the current injector is not capable of
operating satisfactorily above a 5 kHz PWM injection.
6.4.4 Effect of the hydraulic pressure on the injector dynamic
response
Another important variable to study is the hydraulic pressure to actuate the
injector. To understand its influence on the operation of the injector the
actuator speed and hydrogen mass flow rate were analysed for various
pressures with all the other operational parameters kept constant.
Figures 6.60-6.67 show the actuator speed and mass flow rate for varying
hydraulic pressure, with a H2 pressure of 200 bar, static force of 500 N,
272
cylinder pressure of 80 bar, actuator mass of 5 g, a period of 0.0599 s
(2000RPM), and a duty cycle of 20%.
Figure 6.60: Actuator speed for 200 bar hydraulic pressure.
Figure 6.61: Actuator speed for 150 bar hydraulic pressure.
273
Figure 6.62: Actuator speed for 100 bar hydraulic pressure.
Figure 6.63: Actuator speed for 50 bar hydraulic pressure.
274
Figure 6.64: Mass flow rate for 50 bar hydraulic pressure.
Figure 6.65: Mass flow rate for 100 bar hydraulic pressure.
275
Figure 6.66: Mass flow rate for 150 bar hydraulic pressure.
Figure 6.67: Mass flow rate for 200 bar hydraulic pressure.
Figure 6.68 summarises the influence of hydraulic pressure on the mass
flow rate. It can be concluded that the increase in hydraulic pressure is
beneficial regarding the mass flow rate up to a pressure of 200 bar. Higher
276
pressures do not produce higher flow rates as indicated in Figure 6.68,
since the actuation delay is reduced to a minimum and the fluid dynamics
delays then dominate.
Duty Cycle 20% @ 2000 rpm
Mass flow per injection (kg/injection)
0,000325
0,0003
0,000275
0,00025
0,000225
0,0002
0,000175
0,00015
0,000125
50
75
100
125
150
175
200
225
250
275
300
Hydraulic pressure (bar)
Figure 6.68: Relationship between hydrogen mass flow rate per injection
and hydraulic actuation pressure.
6.4.5 Effect of the static force (pre-load) on the injector dynamic
response
To understand the influence that the static force or pre load force has over
the operation of the injector, this force was increased from 125 N up to 750
N while all other parameters were kept constant, as above.
Figures 6.69-6.72 show the actuator speed for varying static force, with a
constant hydraulic pressure of 50 bar, H2 pressure 200 bar, cylinder
pressure 80 bar, actuator mass 5 g, period 0.0599 s (2000RPM), and duty
cycle 20%.
277
Figure 6.69: Actuator speed for a static force of 125 N.
Figure 6.70: Actuator speed for a static force of 250 N.
278
Figure 6.71: Actuator speed for a static force of 500 N.
Figure 6.72: Actuator speed for a static force of 750 N.
Figure 6.73 shows the relationship between the static force and the speed
of response. It can be concluded that the increase of the static force or
pre-load of the actuator, through compression of the spring, increases the
279
speed of response of the injector. This force is, however, limited by the
spring allowable compression. Therefore, in the design of the injector, an
optimum trade-off between the static spring force and the hydraulic
pressure with the view of achieving satisfactory dynamic performance must
be made.
Duty cycle 20% @ 2000 RPM
6
5,5
Speed of response (m/s)
5
4,5
4
3,5
3
2,5
2
1,5
1
0,5
0
100
150
200
250
300
350
400
450
500
550
600
650
700
750
Spring Static Force (N)
Figure 6.73: Relationship between the speed of response and the static
spring load.
6.4.6 Summary of injector design analyses
The weight of the moving parts of the present design is not a limiting factor
of the injector operation. The findings indicate that the operating hydraulic
pressure of the injector should be between 80 bar and 100 bar, and the
static spring load should be around 800N to provide adequate dynamic
performance. The operating speed of the engine with this injector is
limited to around 2000 RPM. This limitation is due to the mass spring
system response.
It is not possible to use the present injector to deal with pulsed injection,
due to the high frequencies involved and its poor frequency response.
Further, due to deficient regulation of the flow, the injector can cause the
engine to operate erratically for values of PWM smaller than 30%.
280
6.4.7 Possible injector design improvements
There are a number of possible improvements to this injector. With the
objective of assuring that no seizure of the moving parts will occur, the
parts were machined with clearances to allow some passage of hydraulic oil
to lubricate the needle valve. Because these clearances were too generous,
and the hydrogen pressure was higher than the hydraulic pressure, a
substantial amount of hydrogen passed through the clearances to the
hydraulic oil return pipe. Therefore better machining and tighter
clearances are need for future design of such a type of injector.
Also it is obvious that the dynamic characteristics of the injector need to
be improved. This can be achieved by constructing a new injector with a
different approach, so that pulsed injection can be used. The use of new
magnetostrictive materials, such as TERFENOL D, could provide the
solution, as this material can cope with frequencies in the order of several
MHz and do not need oil lubrication if graphite is used on the needle
rubbing surfaces.
6.5 Summary
This chapter demonstrated the use of simulation as a tool to further
complement
the
experimental
work.
In
particular
they
allowed
investigation where experiments would be too costly, difficult or even
impossible to carry out.
Using the simulation programs developed during this research, it was
possible to determine the range of the physical parameters to control the
hydrogen injection process and therefore the engine performance.
The simulation of the rate of pressure rise demonstrated the difficulties in
predicting operating parameters with such a random behaviour with the
same accuracy as the other variables, such as peak cylinder pressure. This
possibly is due to the limitations of the mathematical model used for the
simulation studies of this research. However, improvements can be made to
281
better predict such variables, although the results obtained were
satisfactory for the present research since the simulation model allowed
the study of the parameter influence that could not be easily tested using
the test engine, giving one insight on the ruling variables of the hydrogen
fuelled CI engine.
It was found that to overcome the difficulties encountered in terms of
control of the ignition angle, as well as in terms of the rate of pressure
rise, that this cannot be solved by the optimisation of one design variable
only. The optimisation of the engine compression ratio, variable valve
timing, and pulsed injection can be used to provide good engine
performances. However, the simulation evidenced that the use of high
speed and accurate injectors is required for DIH2 operation and this
introduces a number of problems to solve. It was concluded that for HCCI
engines, the injectors do not need to be so fast and accurate as there is
plenty of time to inject the hydrogen into the cylinder, being the main
problem of this mode of operation the control of the ignition angle, that by
itself is a function of the residual gases inside the cylinder, the air
temperature, speed and load.
For hydrogen direct injection it is extremely important that the
temperature of the air at the beginning of compression is controlled, and it
was found that the start of injection should be retarded compared to diesel
operation in order to avoid excessive in-cylinder pressures.
It was found that for the hydraulically actuated injectors, the inertia of the
moving parts is not critical, but the pre load force and the hydraulic
pressure play a fundamental role in the performance of such injectors.
282
Chapter 7
Conclusions and recommendations
« Standing still is the fastest way of moving backwards in a rapidly
changing world »
This thesis has presented a research work developed into the use of
hydrogen as a fuel for CI engines.
A revision of the historical research carried around this subject was
performed, identifying a path for the research and the achievements and
requirements of further research. The use of hydrogen as a fuel for CI
engines when this research started was a new area of research, with very
few available data and reports.
To perform such research an existing Diesel engine was adapted to operate
under HCCI and DIH2 modes, requiring the design and manufacture of a
considerable number of systems, components, such as hydrogen injectors,
injector controllers, hydraulic load brake, hydrogen fuel system, hydraulic
actuator system, inlet air temperature control system, hydrogen detection
system, and a data acquisition software.
A simulation model was developed and validated against the actual engine,
or physical model, for both modes of operation HCCI and DIH2, and used at
a later stage as a complementary research tool. The dynamic simulation
model of the engine operating under HCCI and DIH2 was used extensively as
a tool to overcome the limitations of the physical test stand, such has
variable compression ratio, variable valve timing, high speed pulsed
injection and evaluation of some dynamic variables such as the rate of heat
release, temperature profile in the combustion chamber, angle of
maximum rate of heat release, that would be very expensive to measure if
not impossible. The engine dynamic model allowed also the evaluation of
the effect of environment variables such as inlet air temperature, or
283
atmospheric pressure, but also the study of the impact on the cycle of the
change of some characteristic engine parameters such as compression
ratio, injection duration, fuel mass flow rate, bore, stroke, speed and
valves timing performing the Miller cycle.
An important part of this research was devoted to the design of the
injection systems, in particular the injectors for HCCI mode of operation,
operating at low pressures (less than 6 bar) and the injectors for DIH2 mode
of operation operating at pressures above 50 bar, therefore requiring the
use of hydraulic force for their actuation. For the design of such hydraulic
injectors the use of a mathematical model was also a valuable tool, helping
the understanding of how the design parameters were interrelated with
each other, therefore allowing the construction of a working injector.
Once the required hardware for hydrogen operation (HCCI and DIH2
injection systems) was manufactured and characterised, it was fitted on
the test engine, to operate it under HCCI and DIH2 modes. Then, a number
of test runs were performed to gather data for comparison between the
various modes of operation including diesel. These comparisons were
focused in particular in what concerns thermal efficiency, performances
and identification of aspects that needed to be solved to achieve a safe and
reliable operation with hydrogen under each mode of operation.
The existing rules and standards applicable to safety and manipulation of
hydrogen applicable to engines are inexistent, therefore, during the course
of the present research general rules had to be adapted, and above all a
good sense and fair engineering judgement was employed to avoid unsafe
circumstances.
The findings have been summarised at the end of each chapter and will be
recapped and synthesised in this last chapter. In addition, design
considerations for hydrogen-fuelled engines will be discussed, based on the
experience accumulated during the present research, and recommendation
for further work will be presented.
As a result of the present research the author, developed the sufficient
knowledge to design an injection system and respective engine control to
284
convert two marine diesel alternators of 4.000 kW with a 320mm bore
engines to operate as dual fuel engines (natural gas). Later the author
converted to dual fuel operation (natural gas) also one marine diesel
alternator of 11.000 kW and a smaller diesel alternator to hydrogen
operation. Please refer to appendices A and B.
To achieve, an effective engine protection and control, the author
developed a knocking control system called KDS (Knock Detection System),
based on automatic comparison of vibration spectrums measured at the
cylinder head of the experimental engine, controlling the flow rate of
hydrogen and therefore the high RPR characteristic of hydrogen operation.
The present research, resulted into two papers, edited by the International
Journal of Hydrogen Energy, and a two more papers presented in
conferences and symposiums.
7.1 Summary of the results
Chapter 2 presented a review of the reported hydrogen technologies
related to the properties, handling and applications of hydrogen fuel, as
well as to its use in reciprocating engines. Particular features and
challenges associated with hydrogen-fuelled engines were identified, such
as the potential for high fuel efficiency, lean-burn operation and low
emissions formation, as well as challenges of high pressure rise rates and
ignition timing control. This chapter also presented a review of reported
research, from the early stages of hydrogen use as a fuel until recent
studies highlighting modern engine performance and the difficulties in
storing and handling hydrogen fuel.
In Chapter 3, the design and development of the engine and injectors test
rig systems used in the research were described. The purpose-developed
system controllers, the high-speed frequency data acquisition system and
related hardware and software, and experimental safety requirements
were presented in detail.
285
Chapter 4 presented the experimental methodology followed during the
experimental phase of the research, as well as experimental results and
their interpretation. Main conclusions from the experimental work are as
follows:
From the experimentation tests performed, it was found that:
 The hydrogen operated engine, in HCCI and DIH2 modes, have better
thermal efficiencies (48 % and 42.8% respectively) than the same
engine operated with diesel oil (brake thermal efficiency of 27.9 %).
 From a breakdown of the engine heat balance, it was found that the
bigger heat loss is through the exhaust for the hydrogen-fuelled
modes of operation, rather than through the cylinder liner and
cylinder head as in the diesel oil operated engine.
 The experimental tests further showed that extremely high rates of
pressure rise can result from too high air inlet temperatures, and
these can have a negative impact on the piston crank mechanism
bearings.
 It was found that the HCCI engine is limited in power due to the
gaseous hydrogen fuel displacing a significant amount of intake air.
 However, it was demonstrated that the HCCI engine can be operated
successfully with extremely lean cylinder charges, albeit with higher
cycle-to-cycle variations.
 The power output of the DIH2 engine is not limited in the same way,
but the maximum power may be limited by high peak pressures and
high rates of pressure rise which can cause mechanical problems.
 The controllability and operational stability of the HCCI engine was
found to be challenging due to difficulty in controlling the ignition
angle, whereas the controllability of the DIH2 engine is excellent as
ignition takes place at the initiation of the injection as in a
conventional diesel engine.
286
 The tests of Dual-fuel mode of operation, demonstrated that
hydrogen can be used simultaneously with diesel oil as a source of
ignition, improving the thermal engine efficiency, and the emissions
even in small induced quantities.
 Therefore the Dual-fuel mode of operation has a large potential for
the penetration of hydrogen in the industry, namely on board vessels
where there is a considerable amount of wasted energy to recover
and accumulate as hydrogen, that can be used to drive auxiliary
engines in port, therefore complying with the ever stringent local air
pollution laws.
 High flame speed within the engine cylinder over a wide range of
temperature and pressure. High flame speeds were identified even
for lean mixtures. The energy release is also so fast that the
combustion duration is short contributing for the high power output,
high efficiencies and high rates of pressure rise.
 The high rates of pressure rise, are noticeable even at low loads for
both modes of operation, in particular for HCCI.
 The DIH2 mode of operation allow the use of very lean cylinder
charges, that in combination with the fast rates of heat release
around TDC, results in high thermal efficiency values and
simultaneously in high power output.
 The HCCI mode of operation allow the use of extremely lean cylinder
charges, that in combination with the fast rates of heat release, that
in the case of optimised ignition angle, results in high thermal
efficiency values and simultaneously in lower power output when
compared with DIH2 and diesel operation.
 One of the most important features of HCCI and DIH2 are the less
undesirable exhaust emissions when compared with Diesel fuel. The
lubricating oil is the only source of hydrocarbon emissions that in a
well maintained engine tends to be negligible. Only NOx and water
287
vapour were identified as being the main products of the
combustion.
 For the HCCI mode of operation the hydrogen presence on the
exhaust gases is negligible when timed port injection is used. If
hydrogen is simply fumigated ate the cylinder inlet, the presence of
hydrogen in the exhaust can be quite important as the cylinder is
scavenged with a mixture of air and hydrogen.
 For DIH2 the presence of hydrogen in the exhaust is negligible and it
is a function of the combustion chamber crevices.
 The hydrogen fast burn characteristics allow its use in high speed CI
engines as there is no charge preparation time required by the diesel
fuel, therefore allowing an increase in power output with a reduced
penalty for lean cylinder charges.
 As the self-ignition temperature of hydrogen is quite high when
compared with other hydrocarbon fuels, the compression ratio of
these CI engines must be higher than the ones used for diesel fuel
operation, therefore contributing for higher efficiencies and higher
power outputs of the HCCI and DIH2 engines.
 Hydrogen operation therefore is associated with less heat loss than
with diesel fuels.
 Due to the hydrogen high burning rates and its diffusivity, the HCCI
and the DIH2 are less sensitive to combustion chamber shapes, level
of turbulence and charge swirling effects.
 The thermodynamic and heat transfer characteristics of hydrogen
are
accompanied
by
higher
final
compression
temperatures
contributing to improvements in engine efficiency and lean mixture
operation.
 Less cyclic variations are encountered for hydrogen than with other
fuels, even for lean operation, however for very lean operation these
cyclic variations are a reason for the engine control difficulties.
288
 Less cyclic variations lead to the reduction of emissions, smoother
operation and improved efficiency.
The development of the HCCI and DIH2 injectors was essential to carry out
the experimental studies presented, in particular for the DIH2 engine. The
design, construction and testing of the injectors were described in detail.
In Chapter 5, the development of a full cycle simulation model for the HCCI
and DIH2 was introduced and its functional organization, the sub models
implemented and assumptions made were described. The modelling of the
HCCI and DIH2 injectors were also presented.
It was found that to model the combustion of hydrogen, commonly used
compression ignition engine models for liquid fuels do not apply. The
differences in the combustion properties between conventional liquid fuels
and hydrogen include factors such as the atomisation and preparation of
cylinder charge (since hydrogen is already in its gaseous form), in-cylinder
gas dynamics and fuel-air mixing, and radiative heat transfer losses.
The ignition delay characteristics are different for hydrogen-fuelled
engines, since for temperatures below 1000 ºK it can be very long when
compared with a modern diesel oil injection system ignition delays, but for
higher temperatures the ignition delay is extremely short.
Further, the use of a heat loss model calibrated against the measured
temperatures on the liner and piston head and assuming that the
temperature of cylinder head is the same as the temperature of the piston
crown was essential to achieve thermal efficiencies similar to those
registered during the tests.
In Chapter 6, the engine model was validated against the experimental
data and the uncertainties calculated. Based on simulation, the influence
of various operational variables, such as ambient conditions, was
systematically investigated. Mitigation actions of the high rates of pressure
rise were studied with the objective of reducing the dynamic bearing loads
and improving the engine controllability. The optimisation of the engine
289
operation was studied by evaluating the valve timing, resulting in the
implementation the Miller cycle. Other techniques were also evaluated,
such as pulsed injection, inlet air pre-heating, and compression ratio
variation. Chapter 6 also presents the dynamic simulation of the DIH2
injector, resulting in the identification of the dominating variables for the
DIH2 operational characteristics.
 As a consequence of the short ignition delay and rapid
combustion, the hydrogen combustion process for HCCI and
DIH2 approaches the constant volume cycle.
 The simulation analysis indicated that pulsed injection can
partly mitigate problems associated with high peak pressures
and high rates of pressure rise, but does not provide a
complete solution.
 It was proved that controllability of the HCCI engine is
influenced by the ignition angle variations, which are a direct
function of the end-of-compression temperature, which is
therefore affected by all the parameters on which this
temperature depends like residual gases temperature.
The main parameters influencing the angle of ignition include the
temperature of the air at the inlet of the cylinder, engine load,
compression ratio and heat exchange between the cylinder walls and the
cylinder charge. However, the dominant variable is the air temperature
entering the cylinder, with which engine control can be improved using one
of the solutions presented in the recommendations.
 It was confirmed by simulation that unlike the diesel fuelled engine,
the main losses of heat in the hydrogen engine are through the
exhaust gases, rather than through the liner and cylinder head.
 The reason for such exhaust losses is that the hydrogen combustion is
extremely fast compared to diesel combustion and without
production of radiating particulate matter as for hydrocarbon
combustion.
290
 It was found by simulation that due to the fast combustion, hydrogen
injection could be initiated closer to TDC, thereby improving engine
efficiency.
 Through the simulation study, it was found that an efficient way of
controlling the compression ratio and rate of pressure rise is to
implement valve timing control. Adjusting the valve timings to the
engine load can contribute positively to the engine controllability, as
well as for the control of the emissions of NOx.
 Therefore the Miller cycle was simulated, confirming the above
conclusions.
 The DIH2 injector model was found to be adequate, as it allowed the
setting up of the injector for the tests, but also gave knowledge of
its limitations, learning among other things that the inertia of the
moving parts is a less dominant variable compared with the pre-load
and the hydraulic pressure.
 Simulation studies indicated that pulsed injection can contribute to
mitigate the high rates of energy release, and therefore the rate of
pressure rise, but with the increase of the pulses frequency, the
effect approaches the continuous injection.
 The pulsed injection frequency needs to be adapted and optimised
to the engine speed of operation, to sort some effect. From the
simulation studies and for an engine speed of 2200 RPM it was found
that the combination of start of injection 25º BTDC, duration of
injection 33º, plus pulsed injection at 5 kHz, resulted in the lowest
maximum combustion temperature, lower RPR, lower MRPR, while
maintaining engine power output
 Such high injection frequencies and required accuracy, cannot be
met by using common mechanical injectors as the one used on this
research, which whoever can be employed in low speed engines. A
new technology must be developed, which is introduced on the
recommendations.
291
7.2 Hydrogen as a fuel for CI engines, further considerations
Hydrogen has a number of unique features that potentially make it a
particularly interesting CI engine fuel. Some of these features are the
following:

Hydrogen fuelled CI engines are less sensitive to hydrogen purity
than other hydrogen fuelled energy conversion devices such as fuel
cells.

Heat transfer characteristics of hydrogen fuelled CI engines are
significantly
different
from
those
engines
operating
with
hydrocarbon fuels, as hydrogen combustion is characterised by
lower radiation therefore taking a bigger importance the convective
component of the heat transfer specially for lean cylinder charges.

Hydrogen engines are suitable for waste heat recovery applications
since the energy transfer from some water vapour can add up to the
thermal load output and the corresponding energy efficiency.
Despite the above listed advantages of hydrogen as a fuel, there are a
number of limitations and mechanisms associated with its used as a CI fuel
that is worth to mention. There is a need to address equally this limitation
and suggest means to overcome them.
The following points are a list of features associated with hydrogen as a CI
engine fuel that may require some further engineering effort or remedy:

Hydrogen as a compressed gas at 200 bar and atmospheric
temperature has only approximately 5% of the energy of diesel oil
for the same volume.

Unless the energy required to produce the hydrogen is from a
renewable source or from a low temperature waste heat source the
energy required is bigger than the energy produced as ¾ of the
energy contained in one kilogram of hydrogen may be used to liquefy
292
it. Therefore, the existence of a hydrogen fuelled engine must be
associated to some renewable or waste heat recovery source of
energy.

Dual-fuel (hydrogen+ diesel oil) operation mode, results in a brake
fuel efficiency improvement of up to 5%.

Dual-fuel mode of operation results in drastic particulate matter
emission, even in small quantities such as 10% of the energy per
cycle.

Dual-fuel mode of operation is characterised by stable engine
governing for hydrogen quantities up to 50% of the heat input per
cycle.

Dual-fuel mode of operation can constitute an important way of
dissemination of hydrogen as a fuel for CI engines in the industry.

Dual-fuel mode of operation can potentiate projects where hydrogen
is produced from renewable or waste heat recovery projects, as
diesel-oil or bio-oil working as pilot fuels can be reduced to 5% of the
energy required per cycle, being the remaining 95% of the energy
supplied by the hydrogen.

HCCI engines operated with hydrogen suffer from limited power
output, mainly due to the very low heating value on volume basis.
This effect is more noticeable when operating with very lean
cylinder charges.

The
above
effect
is
aggravated
due
to
the
relative
high
stoichiometric hydrogen to air ratio.

There are considerable potential problems associated with backfiring
and pre-ignition into the intake manifold of HCCI engines, noticed
clearly during the tests carried. Hydrogen pre-ignition is mainly due
to its low ignition energy.

The high burning rates of hydrogen may produce high temperatures
and pressures during combustion, in particular for less lean cylinder
293
charges, and may lead to higher rates of thermal NOx emissions.
Therefore an optimised engine control system is required to provide
acceptable engine performance and lower emissions.

The use of cold exhaust gas recirculation, may be a limitation as the
cylinder temperature control has a major importance on the angle of
ignition and therefore on engine control.

There is always some potential for increased safety problems with
hydrogen handling systems and operation.

Materials compatibility problems need to be considered for all
engine components in direct contact with hydrogen.

There is a potential for corrosion problems and lubricating oil
deterioration due to condensed water from its combustion and
viscosity reduction due to hydrogen incorporation by the lubricating
oil molecules.

It was found that the cylinder charge temperature at the end of the
compression stroke is the variable that has the highest influence on
the ignition angle and subsequent rate of pressure rise. This variable
is therefore of critical importance for engine controllability.

Consequently, all the variables influencing the end-of-compression
temperature, like inlet air temperature, temperature of the residual
gases, temperature of combustion chamber, compression ration must
be controlled to control the rate of pressure rise and angle of
ignition. Therefore, one or more of these variables need to be
controlled to achieve an acceptable engine controllability and
efficiency.

Due to the high self-ignition temperature of hydrogen, it was found
that to achieve stable operation of the engine in compression
ignition hydrogen fuelled mode, the intake air had to be heated
above 70ºC.

In the HCCI operation, the higher the air fuel ratio the lower is the
required inlet air temperature to achieve the cylinder charge
294
ignition as higher temperature of the residual gases help on the
required end of compression temperature. Therefore the higher the
engine speed the earlier would be the ignition, contributing for
lower engine controllability.

In HCCI engines the angle of ignition is an inverse linear function of
the air at the cylinder inlet temperature.

In the HCCI, exhaust gas temperature, maximum combustion
pressure and angle of ignition, increase with engine load.

Hydrogen ignition delay is short and its combustion is extremely fast.
To achieve a good thermal efficiency, the angles of injection and
ignition of the cylinder charge need to be very close the TDC,
therefore producing the maximum of positive work.

For HCCI CI engines, the parameters influencing the angle of ignition
need to be accurately controlled, being the most effective way of
doing it through the control of the compression ratio, through the
used of variable valve timing system. Despite the possible
contribution to the control of the angle of ignition and rate of
pressure rise, of variables like cylinder inlet air temperature they
not sufficiently fast.

The variable valve timing control system, it is an extra complication
and cost to consider for a controlled HCCI engine operation.

By comparing the heat losses between the various modes of
operation at the same speed, and load, it was concluded that
exhaust losses are less predominant on the HCCI engines than on
diesel and DIH2. This fact is due to a lower heat input per cycle (high
equivalence ratio) characteristic of HCCI operation.

DIH2 operation has lower cooling losses than the HCCI and diesel
mode of operation as combustion takes place closer the TDC than
the other modes of operation.

For the HCCI engine, the higher the temperature inside the cylinder
the earlier is the ignition, and therefore the lower is the engine
295
efficiency and the higher the rate of pressure rise. Whereas for the
DIH2 the cylinder charge temperature has only effects on the ignition
delay angle and rate of pressure rise, that some how can controlled
if the pressure development takes place only after TDC.

DIH2 mode of operation can deliver more power, than HCCI and
Diesel modes, this is mainly due to the absence of an air fuel ratio
limit, therefore the combustion can be made so rich as required.
Also there is no need for cylinder charge preparation time, as
hydrogen is injected in gaseous state, which allied to its diffusivity
and fast combustion development after TDC contribute for such high
power capability.

For the DIH2 mode of operation, the production of NOx is an
exponential function of the indicated mean effective pressure for
both modes of operation.

For a constant heat input per cycle, the DIH2 thermal efficiency is
higher for higher engine speeds, and the RPR is practically constant
through all the engine load range.
296
7.3 Recommendations for further work
Hydrogen fuelled CI engines can exhibit operational characteristics that are
superior to those associated to conventional diesel fuels. The results
presented in this thesis indicate that further developments of hydrogen
fuelled CI engines should be oriented towards the control of the angle of
ignition and rate of pressure rise, necessarily including the development of
fast response control systems. For the HCCI engine, operational stability
and the minimising of cycle-to-cycle variations is of critical importance,
and these must be resolved before such engines can become commercially
viable. This research has identified the important control variables for
combustion control, including the level of pre-heating of the intake air and
compression ratio control. Significant improvements in the controllability
of such engines can probably be achieved in a more detailed study
focussing on engine control issues related to the control of such variables.
The work presented has shown that satisfactory hydrogen injectors can be
developed using standard techniques and materials, however the designs
can be refined and improved using advanced materials and fast-acting
actuators.
The information contained on this research work have given new insights
into the areas that require further development, but complementary
knowledge concerning the engine applications and integration should also
be studied.
In the next paragraphs, are outlined some possible approaches contributing
for the improvement of the hydrogen fuelled compression ignition engines.
A number of possible design and operation features need to be addressed
to make the hydrogen CI hydrogen fuelled engines serious candidates for
some applications. When most of these following recommendations are
implemented in the design of CI hydrogen fuelled engines, then most of the
apparent limitations associated with hydrogen as a fuel for CI engines will
be minimised or even disappear.
297
7.3.1 Feasibility of the hydrogen fuelled CI engine
As mentioned above, the feasibility of the hydrogen fuelled CI engine
depends on the solving of the problems identified during this and previously
reported research. If these challenges can be solved, this could make
hydrogen engines (with HCCI or DIH2 mode of operation) an interesting
alternative to the fuel cell technologies. The main problems that need to
be solved, not considering the fuel storage and safety aspects, are related
to the control of the combustion process and associated mechanical stress,
as well as the operational control of the HCCI engine, in particular the
angle of ignition in order to obtain stable engine operation.
In what concerns the engine controllability, one of the main difficulties
identified
during
the
research
was
that
the
end-of-compression
temperature may be below the hydrogen self-ignition temperature, leading
to ignition failure. There are various methods to overcome this difficulty,
such as a mechanism for compression ratio adjustment or pre-heating of
the intake air. These will have varying technical difficulties in their
implementation, related with the technologies to employ and their speed
of response.
7.3.2 Engine mechanical loading and controllability
When considering the use of the hydrogen as a fuel to operate the engine
under HCCI or DIH2 mode, the mechanical loads applied to the engine crank
B
B
mechanism and piston rings as a consequence of the combustion
characteristics of hydrogen must be considered. Due to the fast combustion
and rapid pressure rise, mechanical loading must be calculated to confirm
that it will not exceed acceptable limits for conventional engine piston,
crankshaft and bearings design.
As a practical engineering rule, the maximum load-carrying capacity of a
crank bearing should be less than 410 m bar / s (Miralles et al, 1986).
Figure 7.1 illustrates the bearing dimensions and Figure 7.2 illustrates the
298
connecting rod and crank mechanism, in particular the big end bearing
shell.
d
a
Figure 7.1: Main bearing dimensions.
The analysis required to determine the relationship between cylinder peak
pressure and main engine bearing load capacity is as follows. Bearing load
capacity is defined as the product of the peripheral velocity of the bearing
journal and the minimum bearing surface contact pressure (Miralles, 1986).
Therefore
C = v p  PS ,
(7.1)
C is the load carrying capacity (m bar/sec),vp is the peripheral velocity
B
(m/s), and PS is the minimum surface pressure (bar).
B
The peripheral velocity is given by
vp =
ne πd
,
60
(7.2)
ne is the engine velocity (rad/s) and d is the bearing diameter (m).
B
,
The minimum surface pressure is given by
pππ 2
Ps =
,
4ad
Where, a is the bearing width (m) and B is the cylinder bore (m).
299
(7.3)
Substituting Equations 7.2 and 7.3 into Equation 7.1 gives the load carrying
capacity as
C=
4.1B 2 ne p
.
10 5 a
(7.4)
It can be seen from Equation 7.4 that the load carrying capacity of the
engine bearings for a particular engine cylinder bore and speed is directly
related to the maximum cylinder pressure. Accurate control of the ignition
angle and rate of pressure rise is therefore essential.
7.3.3 Compression ratio adjustment
The possibility of using a variable compression ratio mechanism will allow
indirect control of the cylinder charge final compression temperature. By
implementing a valve timing control system, a dynamically optimised
compression ratio can be obtained as a function of the engine operating
parameters. This can give direct control over the angle of ignition and
actual rate of pressure rise and therefore the engine controllability will be
improved. Such mechanism, if optimised, will also help in the control of
the NOx emissions. This mechanism was simulated on Chapter 6.
7.3.4 Control of the inlet air temperature
As mentioned above the control of the temperature of the air entering the
cylinder is another possibility for controlling the angle of ignition and rate
of pressure rise, in particular for the HCCI engine. Figure 7.2 illustrates a
proposal for a system based on the circulation of exhaust gas trough a heat
exchanger fitted at the inlet air manifold. In this way air is heated or
cooled throughout the entire engine load range, by using a fast acting three
way by-pass valve. The speed of response of this by-pass valve needs to be
high, as the heat transfer via an heat exchanger is not a sufficiently fast
300
process to cope with sudden engine load changes. A challenge that is not
solved by this proposed system is the unavailability of hot gases during the
engine start up.
Exhaust Air
heater
ECU
Exhaust By-pass
Valve
Air inlet
Figure 7.2: Ignition angle control through inlet air heating using an exhaust
gases heat exchanger.
To start the engine, the use of auxiliary power must therefore be provided,
for example through the use of an electric heater fitted on the air inlet
manifold, as used on the test engine. Once the engine is started and the
exhaust gas flow is controlled by the ECU through the three-way valve, the
electric heater is switched off.
Figure 7.3 illustrates another system proposal based on the heating of the
cylinder charge using recirculation of exhaust gas. The heating is achieved
by direct contact of the hot gases with the air entering the cylinder. This
solution also has some interesting “side effects”, one of them is the control
of some exhaust chemical species such as NOx. Nevertheless, this solution
has limitations regarding the percentage of exhaust gas that can be
301
recirculated, associated with the availability of oxygen for combustion and
humidity in the lubricating oil and engine components.
Exhaust Byp as s
V a l ve
ECU
A i r i n let
Figure 7.3: Combustion control through cylinder charge heating by
recirculation of exhaust gases.
However, the system proposed in Figure 7.3 has the advantage of being
much faster acting than that proposed in Figure 7.2, since the heat is
transferred by direct contact of the cylinder residual gases with the
cylinder incoming air.
7.3.5 Internal exhaust gas recirculation
Internal exhaust gas recirculation can be implemented using the valve
timing mechanism, requiring a valve control system to let a specific volume
of hot exhaust gases remain inside the cylinder.
302
Care must be taken to avoid increasing hydrogen slip, which results in
poorer fuel efficiency and presence of unburnt hydrogen in the exhaust
gases. This difficulty can be slightly improved by reducing the valve overlap
timing, therefore decreasing the waste of hydrogen and making better use
of it. As a consequence, the decrease of the valve overlap period increases
the amount of residual gases trapped inside the cylinder, resulting in an
internal exhaust gas recirculation.
2,79
350
2,785
340
2,78
330
2,775
320
2,77
70
72
74
76
78
80
82
Exhaust gas temperature (ºC)
IMEP (bar)
2200 RPM, Ta = 90 ºC, mH2 = 6.02 g/min
310
84
Exhaust gas valve opening angle BBDC (degrees crank angle)
Figure 7.4: Exhaust gas internal recirculation by reduction of valve overlap
period.
As can be seen in Figure 7.4, there are two obvious effects on the engine
performance resulting from the reduction of the valve overlap period. One
is a decrease in the indicated mean effective pressure and corresponding
power, the other a substantial increase of the exhaust gas temperature
resulting in a lower thermal efficiency. This was shown in more detail in
Chapter 6 in the simulation of the Miller cycle.
303
7.3.6 Pulsed injection for DIH2
The simulation study of the pulsed injection in Chapter 6 indicated that this
technique can be beneficial for the control of the heat release and the
corresponding rate of pressure rise. The release of the hydrogen in small
controlled quantities will avoid the sudden heat release and uncontrolled
pressure rise if the whole amount of hydrogen is burnt in one process.
If hydraulic injectors can perform well when installed in the slow to
medium speed engines, the same would not be possible for high speed
engines, therefore calling for high speed and highly accurate injectors. To
implement such a high speed and accurate injection system, new technical
solutions are required, as the present technologies are not sufficiently fast
and accurate to satisfy such requirements. Pulses with frequencies above 1
kHz can be achieved by using new magneto-restrictive materials such as
Terfenol-D. These materials can be used to produce fast acting and simpler
hydrogen injectors.
These materials, known as magnetostrictive or giant magnetostrictive
materials (GMM), are alloys that change their physical form (dimensions) as
a function of the magnetic field applied. The physical phenomena that
characterises the effect is illustrated in Figure 7.5. Under a magnetic field
action the dipoles become oriented in such way that the original geometry
of the magnetostrictive material rod is changed.
Figure 7.5: Working principle of magnetostrictive materials
(www.etrema-usa.com).
304
Magnetic field induced strain materials are classically represented by GMMs
such as rare earth-iron discovered by Clark (www.Wikipidia.org) These
materials feature magneto strains which are two orders of magnitude larger
than Nickel. Among them, bulk Tb0.3Dy0.7Fe1.9, called Terfenol-D, has been
commercially available since 1987 and presents the best compromise
between a large magneto strain and a low magnetic field at room
temperature. The electric actuating circuit, illustrated in Figure 7.6, is
very simple. The deformation is a direct function of the intensity of the
magnetic field applied to the material. A coil wounded around the material
rod induces the required magnetic field to obtain the desired rod
deformation. A Zener diode serves to limit the currents induced by the coil.
Figure 7.6: Basic electric actuating circuit.
305
Figure 7.7: Strain magnetic field intensity of Terfenol-D
(www.etrema-usa.com).
Figure 7.7 illustrates the characteristics of strain as a function of the
magnetic field at room temperature for Terfenol D. Positive magneto strain
of 1000 to 2000 ppm (0.1-0.2%) obtained with fields of 50 to 200 KA/m are
reported for bulk materials, giving the possibility of building high power
transducers and actuators. These characteristics have renewed the interest
for magneto striction and much progress in the applications of GMMs has
been made in the last 20 years.
Magnetostrictive materials for actuators
From the commercial point of view, there are three available sources of
Tb0.3Dy0.7Fe1.9, GMM. Etrema Products Inc. (US) produces rods with
dimensions varying from 2 to 68 mm in diameter and from 6 to 250 mm in
length, as well as plates and powder. Figure 7.8 shows some of the
material shapes available and Table 7.1 lists the main physical properties
of Terfenol-D. Gangsu Tianxiang Rare Earth Functional Material Co. Ltd
(China) produces rods with dimensions varying from 5 to 50mm in diameter
306
and up to 200mm in length. The third manufacturer is Materi Tek Co. Ltd
(China). These latter two companies explore the wealthy resources of rare
earth materials in China.
Figure 7.8: Various shapes of Terfenol-D
(www.etrema-usa.com).
Table 7.1: Terfenol-D mechanical properties (www.etrema-usa.com).
Elastic Modulus
20-50 GPa
Thermal Conductivity 10.0 W/mK
Density
9210 kg/m3
Specific Heat
Speed of Sound
1470 – 2330 m/s Relative Permeability 1.5 – 12
Tensile Strength
28 MPa
Electrical Resistivity
350 J/kgK
60 Ω cm
Compression Stregth 700 MPa
Curie Temperature
380 ºC
Bulk Modulus
Energy Density
1-25 kJ/m3
Coupling Factor
0.75
90 GPa
Thermal Expansion 12 ppm/ºC
307
Figure 7.9 shows a comparison of various materials stress strain curves.
Figure 7.9: Stress-strain comparison for various selected active materials
(www.etrema-usa.com).
Another class of materials potentially suitable for this purpose is memory
shape materials, or MSMs. The MSM material is a ternary Heusler-type alloy
of Nickel, Manganese and Gallium with the characteristic of “one way
colossal magnetosctrictive effect” when magnetized in the active plane
(www.etrema-usa.com). These materials have a maximum unloaded
elongation measure of 5.8 %, which is 32 times more strain than the 0.18 %
elongation of the Terfenol–D. There is, however, a trade-off as for greater
strain, the elastic modulus (stiffness) is lower, measured as 0.5 GPa, or 46
times less than the 23.4 GPa of Terfenol-D. Also, the material is not
suitable for high temperature operation, because it displays a full austenite
transition at 48ºC.
308
Figure 7.10 shows the strain temperature characteristics of Terfenol-D as a
function of temperature for different Telurium concentrations.
Figure 7.10: Terfenol-D temperature saturation strain.
(www.etrema-usa.com).
A proposed injector based on Terfenol-D is shown in Figure 7.11, and is
characterised by a broad band frequency range of operation, wide
temperature range, high energy density, fast response, high reliability and
high energy conversion.
However, due to difficulties in buying a rod of Terfenol-D, this injector was
not manufactured and tested, but it constitutes an interesting and
promising area for further development and research.
309
Nu t c ap
Was her s top per
Body
Coil
T er fe no l- D r od
Nut
Nozz le
Needle
Ac tu ator c le are nce
H yd rog en su pp ly
Sack vo lume
Figure 7.11: Terfenol-D based hydrogen injector.
310
References
Allenby S, Chang W-C., Megaritis A. and Wyszynski M. (2001). Hydrogen
enrichment: a way to maintain combustion stability in a natural gas
fuelled engine with exhaust gas recirculation, the potential of fuel
reforming. Proc. Inst. Mech. Eng., Part D: Journal of Automobile
Engineering, Vol. 215, pp 405-418.
Alperstein M., Swim W.B. and Schweitzer P.H. (1958). Fumigation kills
smoke — improves diesel performance. SAE Transactions, Vol. 66, pp
574-588.
Aly H. and Voss A. (1993). Experimental investigation of gaseous hydrogen
utilization in a duel-fuel engine for stationary power plants. Trans.
ASME, Vol. 20, pp 67-79.
Antunes J. and Roskilly A.P. (2004). The use of H2 on compression ignition
engines. In: 3rd European Congress on economics and management
of energy in industry, Lisbon.
Antunes J. and Roskilly A.P. (2006). Opportunities and advantages of the
use of hydrogen on board ships – new concepts. In: X International
Naval engineering conference, Lisbon.
Antunes J., Roskilly A.P., Mikalsen R. (2008). An investigation of
hydrogen fuelled HCCI engine performance and operation.
International Journal of Hydrogen Energy, Volume 33, Issue 20,
October 2008, Pages 5823-5828.
Antunes J., Roskilly A.P., Mikalsen R. (2008). An experimental study of a
direct injection compression ignition hydrogen engine.
International Journal of Hydrogen Energy, Volume 34, Issue 15,
August 2009, Pages 6516-6522.
Arenas V.M. (1978). Motores de Combustion Interna. Tomo I, ed. 78-79,
ETSIN, Spain.
Arnason B. and Sigfusson T. (2003). Application of geothermal energy to
hydrogen production and storage. Bragastofa, Hydrogen Energy
Research Institute, University of Iceland, Iceland.
Arnold W.C., Beadle R.H., Logelin R.L., and Young H.D. (1958). Bifuel
approach to burning residual fuels in diesel engines. SAE
Transactions, Vol. 66, pp 55-64.
Bade Shrestha S. and Karim G. (1999). Hydrogen as additive to methane for
sparking ignition engine applications. International Journal of
Hydrogen Energy, Vol. 24, No. 6, pp 577-586.
Baird B. and Gollahalli S.R. (2000). Emissions and efficiency of a spark
ignition engine fuelled with a natural gas and propane mixture. In:
Proc. 2000 International Joint Power Generation Conference,
Florida, USA.
311
Benson R.S. (1982). The Thermodynamics and Gas Dynamics of InternalCombustion Engines, Volumes I and II. Clarendon Press, Oxford.
Beruoun S., Martins J. (2001). The development of gas (CNG, LPG and H2)
engines for buses and trucks and their emission and cycle variability
characteristics. SAE paper 01-01444.
Bishop, R.H. (1993). Modern Control Systems Analysis and Design Using
MATLAB®. Addison-Wesley Publishing Company, USA
Blair, G.P. (1999). Design and Simulation of Four-Stroke Engines. SAE
International, USA.
Borman G. and Ragland K. (1998). Combustion Engineering, McGraw-Hill
International editions, Mechanical engineering series, Singapore.
Bowns D.E., Cave P.R., Hargreaves M.R.O. and Wallace F.J. (1973).
Transient characteristics of turbocharged Diesel engines. Proc. Inst.
Mech. Eng., CP15, 1973.
Brian D.H. (2002). Essential MATLAB for Scientists and Engineers,
Butterworth Heinemann, 2002.
Brian R.H., Lipsman R.L., Rosenberg J.M. (2001). A guide to MATLAB for
Beginners and Experienced Users. Cambridge University Press, 2001.
Calkins F., Smith R. and Flatau A. (2000). An energy-based hysteresis model
for magnetostrictive transducers. IEEE Transactions on Magnetics,
Volume 36, pp 429-439.
Çengel, Y.A. and Boles, M.A. (2001). Termodinâmica, 3rd edition. McGrawHill de Portugal, Portugal.
Chang J., Güralp O., Filipi Z., Assanis D., Kuo T., Najt P. and Rask R.
(2004). New Heat Transfer Correlation For an HCCI Engine Derived
from Measurements of Instantaneous Surface Heat Flux. SAE Paper
2004-01-2996.
Chesse P., Chalet D., Tauzia X., Hetet J. and Inozu B. (2004). Real-time
performance simulation of marine diesel engines for the training of
navy crews. Marine Technology Society Journal, Vol. 41, pp 95–101.
Chryssakis C. and Assanis D. (2005). Effect of multiple injections on fuel-air
mixing and soot formation in diesel combustion using direct flame
visualization and CFD techniques. Proceedings of ICES2005, ASME
Internal Combustion Engine Division 2005 Spring Technical
Conference, Chicago, IL, USA.
Citron S., O´Higgins J., and Chen L. (1989). Cylinder by cylinder engine
pressure torque waveform determination utilizing crankshaft speed
fluctuations. SAE paper 890486.
College of the desert and SunLine Transit Agency (2001). Hydrogen use in
internal combustion engines. Hydrogen Fuel Cell Engines and Related
Technologies Course Manual, Module 3, Rev. 0, U.S. Department of
Energy - Energy Efficiency and Renewable Energy, Hydrogen, Fuel
Cells and Infrastructure Technologies Program – Technology
Validation, USA.
312
Collier
K. (2001). Hydrogen/Natural Gas Blends for Heavy-Duty
Applications. Proceedings of the 2001 DOE Hydrogen Program
Review.
Das L., Gulati R. and Gupta P. (2000). Comparative evaluation of the
performance characteristics of a spark ignition engine using hydrogen
and compressed natural gas as alternative fuels. Journal
of
Hydrogen Energy Vol. 25, pp. 783-793.
Derry L., Dodds E., Evans E. and Royle D. (1954). The effect of auxiliary
fuels on the smoke-limited power output of diesel engines. Proc.
Inst. Mech. Eng. Vol. 168.
Duane H. and Littlefield B. (2001). Mastering MATLAB® 6: A Comprehensive
Tutorial and Reference. Prentice Hall, USA.
Eastop, T.D. and McConkey A. (1993). Applied Thermodynamics for
Engineering Technologists. 5th Edition, Prentice Hall.
Evenson W. (2004). The potential for a hydrogen energy economy, In: IUPAP
Energy Reports - Report on research and development of energy
technologies, USA, pp 207-232.
Ewan, B.C.R and Moody K. (1986). Structure and Velocity Measurements in
Under expanded Jets. Combustion Science and Technology, Vol.45,
pp 275-288.
Favennec A-G., Minier P. and Lebrun M. (1999). Analysis of the Dynamic
Behaviour of the Circuit of a Common Rail Direct Injection System.
In: Fourth JHPS International Symposium on Fluid Power, Tokyo.
Ferguson C. and Kirkpatrick A. (2001). Internal Combustion Engines: Applied
Thermosciences. Wiley.
Filipi Z., Homsey S., Morrison K., Hoffman S., Dowling D., Assanis D. (1997).
Strain Gage Based Instrumentation for In-Situ Diesel Fuel Injection
System Diagnostics. W.E. Lay Automotive Laboratory, University of
Michigan.
Filipi Z.S. and Assanis D.N. (1997). A non-linear, transient, single-cylinder
diesel engine simulation for predictions of instantaneous engine
speed and torque. In: Proc. Tech. Conf. ASME ICE division, number
97-ICE-8, pages 61-70.
Fiveland S. and Assanis D. (2000). A four-stroke homogeneous charge
compression ignition engine simulation for combustion and
performance studies. SAE paper 2000-01-0332.
Fukuma T., Fujita T., Pichainarong P. and Furuhama S. (1986). Hydrogen
Combustion Study in Direct Injection Hot Surface Ignition Engine. SAE
Paper 861579.
Gao Z. and Schreiber W. (2001). The effects of EGR and split fuel injection
on diesel engine emission. International Journal of Automotive
Technology, Vol.2, pp 123-133.
Geisler O. and Rulfs H. (1993). Research investigations of the combustion of
heavy fuel , natural gas and hydrogen in marine diesel engines.
313
Marine system design and operation Marine Management (Holdings)
Ltd.
Haddad S. and Watson N. (1984). Principles and Performance in Diesel
Engineering. Ellis Horwood Limited, England.
Hann B. (2002). Essential MATLAB for Scientists and Engineers. Second
edition, Butterworth Heinemann, England.
Harris H. L. and Taber W. (1996). Control of industrial engine and gas
turbine exhaust emissions to meet present and future clean and
regulations. American Society of Mechanical Engineers; Book No.
H1048B
Hawley J., Wallace F. and Khalil-Arya S. (2003). A fully analytical
treatment of heat release in diesel engines, Proc. Inst. Mech. Eng.,
J. Automobile Engineering, Vol. 217, pp 701-717.
Heffel J. (2003). NOx emission and performance data for a hydrogen fuelled
internal combustion engine at 1500 rpm using exhaust gas
recirculation. International Journal of Hydrogen Energy, Vol. 25, pp.
901-908.
Heywood J. (1988). Internal Combustion Engine Fundamentals. McGrawHill, USA.
Hicks, T.G. (1985). Standard Handbook of Engineering Calculations. 2nd
Edition, McGraw-Hill Book Company, USA
Hunt B., Lipsman R., Rosenberg J., Coombs K., Osborn J. and Stuck G.
(2001). A guide to MATLAB for Beginners and Experienced Users.
Cambridge University Press, Cambridge, UK.
Imperial, J.M. (1980). Sobrealimentação de Motores. Edições Cetop, MiraSintra Mem Martins.
Imperial, J.M. (1980). Turbo: Sobrealimentação de Motores Rápidos.
Ediciones CEAC, Mem Martins.
International Organization for Standardization (1995). Guide to the
Expression of Uncertainty in Measurement, Switzerland.
Jensen J.P., Kristensen A.F., Serenson S.C., Houbak N., and Hendricks E.
(1971). Mean Value modeling of a small turbocharged diesel engine.
SAE paper 910070.
Kabat D., Heffel J. (2002). Durability implications of neat hydrogen under
sonic flow conditions on pulse-width modulated injectors.
International Journal of Hydrogen Energy, Vol. 27, pp. 1092-1102.
Karim G. (2003). Hydrogen as a spark ignition engine fuel. International
Journal of Hydrogen Energy, Vol. 28, pp. 569-577.
Klass D. (1998). Biomass for Renewable Energy, Fuels, and Chemicals.
Academic Press, USA.
Klein M. and Erikson L. (2002). Models, methods and performance when
estimating the compression ratio based on cylinder pressure.
Vehicular systems, Limkopings Universitet, Sweden.
314
Kubesh J. (2002). Uncertainty in the determination of thermal efficiency in
natural gas engines. Proceedings of ICEF2002, 2002-522 Fall
Technical Conference of ASME Internal Combustion Engine Division,
New Orleans, Lousiana, USA.
Lapuerta M., Hernández J. and Armas O. (2001). Kinetic Modelling of
Gaseous Emissions in a Diesel Engine. SAE Paper 2001-01-2939.
Ledger J. and Walmsley S. (1971). Computer simulation of a turbocharger
diesel engine operating under transient load conditions. SAE paper
710177.
Lee J., Kim Y. and Caton J. (2002). The Development of a Dual Injection
Hydrogen Fueled Engine with High Power and High Efficiency. In: Fall
Technical Conference of the ASME-ICED New Orleans, Louisiana,
USA.
Lyn W. (1952). An Experimental investigation into the Effect of Fuel
Addiction to Intake Air on the Performance of a Compression-ignition
Engine.
Miedema S. and Lu Z. (2002). The dynamic behaviour of a Diesel engine.
Proc. WEDA XXII Technical Conference & 34th Texas A&M Dredging
Seminar, June 12-15, Denver, Colorado, USA.
Moore N. and Mitchell R. (1955). Combustion in Dual-fuel Engines.
Naber J. and Siebers D. (1997). Hydrogen combustion under diesel engine
conditions. Pergamon PII:SO360-3199(97)00083-9
Naber J. and Siebers D. (1996). Effects of Gas Density and Vaporisation on
Penetration and Dispersion of Diesel Sprays. SAE Paper 960034.
Nagaki H., Hitohidde F. and Takahashi S. (1999). Acceptability of Premixed
Hydrogen in Hydrogen Diesel Engine. SAE Paper 1999-01-2521.
Norbeck J., Barth M., Farrel J. and Heffel J. (1997). Development and
Evaluation of a Hydrogen Fuel Power Plant for a Hybrid Electric
Vehicle - Phase II. Report submitted to the South Coast Air Quality
Management District, contract 95073, project 7, University of
California, College of engineering, Center for Environmental
Research and Technology, USA.
Numata A., Kumagai T., Nagae Y. and Osafune S. (2001). Increase of
Thermal Efficiency and Reduction of NOx Emissions in DI Diesel
Engines. Mitsubishi Heavy Industries, Ltd; Technical Review Vol. 38
No.3.
Nwafor O. (2002). Knock characteristics of dual-fuel combustion in diesel
engines using natural gas as primary fuel. Sādhanā - Academy
Proceedings in Engineering Sciences, Vol. 27, part 3, pp 375-382.
Ohta T. (1979). Solar-Hydrogen Energy Systems. Pergamon Press, England.
Olsson J-O., Tunestal P. and Haraldsson G. (2001). A turbo charged dual
fuel HCCI engine. SAE Paper 11896, SAE Fuels & Lubricants Meeting
and Exposition, Orlando FL, USA.
315
Olsson J-O., Erlandsson O. and Johansson B. (2000). Experiments and
simulation of a six-cylinder homogeneous charge compression
ignition (HCCI) engine. SAE transactions, Vol. 109, 3, pp 2046-2056.
Paramo M. and Santavicca, (2002). Gasoline Direct Injection Engine Cold
Start Improvement by Injection of Hydrogen, Available at:
http://forms.gradsch.psu.edu/equity/sroppapers/2002/ParamoMelvi
n.pdf
Passarini L.; Nakajima P. (2003). Development of a high-speed solenoid
valve: an investigation of the importance of the armature mass on
the dynamic response. Journal of the Brazilian Society of Mechanical
Sciences and Engineering, Vol. 25, pp. 329-335.
Peschka W. (1998). Hydrogen: the future cryofuel in internal combustion
engine. International Journal of Hydrogen Energy, Vol. 23, pp. 2743.
Petrovsky N. (1960). Marine Internal Combustion Engines. MIR Publishers,
Moscow.
Pratap R. (2002). Getting Started with MATLAB – A Quick Introduction for
Scientists and Engineers. Oxford University Press, New York.
Pratt J., Flatau A. (1993). Development and analysis of a self-sensing
magnetostrictive actuator design. Proc. SPIE, Smart Structures and
Materials 1993, Vol. 1917, pp 952-961.
Rai G. (1983). Practical Thermodynamics with Questions and Answers.
Khanna Publishers, Delhi.
Reed R. (2004). Study of the feasibility and energy savings of producing and
pre-cooling hydrogen with a 5-kW ammonia based combined
power/cooling cycle. MSc thesis, University of Florida, USA.
Rizzoni G. (1989). Estimate of indicated torque from crankshaft speed
fluctuations: a model for the dynamic of the IC engine. IEEE Trans.
Vehicular Technology, Vol. 38, 168-179.
Rogers G. and Mayhew Y. (1980). Engineering Thermodynamics Work & Heat
Transfer. Third Edition, Longman Group Limited, USA.
Rottengruber H., Berckmüller M., Elsässer G., Brehm N. and Schwarz C.
(2004). Direct-injection hydrogen SI-engine – operation strategy and
power density potentials, SAE Paper 2001-01-2927.
Rottengruber H., Wiebicke U., Woschini G. and Zeilinger K. (2000).
Hydrogen diesel engine with direct injection, high power density and
low exhaust gas emissions. Part 3: engine testing and calculation.
MTZ Worldwide.
Rottengruber H., Wiebicke U., Woschini G., Zeilinger K. (1998).
Investigation of a direct injecting hydrogen diesel-engine. Chair for
Combustion Engines and Road Vehicles (LVK) Technische Universität
München, Germany.
316
Ryan T. (2003). Diesel engine alternatives, In: 2003 Diesel Engine Emissions
Reduction (DEER) Conference Presentations, Session 7: Combustion
and HCCI Regimes; Newport, Rhode Island, USA.
Shao, J., Yan Y., Greebes G. and Smith S. Quantitative characterization of
Diesel sprays using digital imaging techniques. Institute of Physics
Publishing Measurement Science and Technology PII:so9570233(03)55986, 2003.
Shengchang Z., Yangzeng X., Guanglin S. and Tingxiu Z. (2001). Study on
extra-high speed digital valve. In: Proc. Fifth International
Conference On Fluid Power Transmission And Control (ICFP2001),
Hangzhou, China.
Sher E. (1998). Handbook of Air Pollution from Internal Combustion
Engines: Pollutant Formation and Control. Academic Press Limited,
UK.
Shudo T, and Suki H. (2002). Applicability of heat transfer equations to
hydrogen combustion. JSAE Review, 23, pp. 303-308.
Shudo T., Nakajima Y. and Futakuchi T. (2000). Thermal efficiency analysis
in a hydrogen premixed combustion engine. JSAE Review, 21, pp.
177-182.
Shudo T., Shimamura K. and Nakajima Y. (2000). Combustion and emissions
in a methane DI stratified charge engine with hydrogen pre-mixing.
JSAE Review, 21-1, pp. 3-7.
Shudo T., Suzuki H. (2002). New heat transfer equation applicable to
hydrogen-fuelled engines. Proc. of ICEF2002, New Orleans,
Louisiana, USA.
Sorensen H. (1983). Energy Conversion Systems. John Wiley and Sons Inc,
USA.
Stan C., Lefebvre J.-L., Lebrun M. and Martorano L. (1999). Direct gasoline
injection for two-stroke-scooter engines: concept, modelling and
performance, Proceedings 4th International Conference on Internal
Combustion Engines: Experiments and Modeling, ICE 99, Capri 1999.
Stenlåås O., Christensen M., Egnell R., Johansson B. and Mauss F. (2004).
Hydrogen as homogeneous charge compression ignition engine fuel.
In: SAE Fuels & Lubricants Meeting & Exhibition, Toulouse, France.
Stone, R. (1999). Introduction to Internal Combustion Engines. 3rd Edition,
Palgrave MacMillan Press.
Taylor C. (1985). The Internal-Combustion Engine in Theory and Practice Volume 1: Thermodynamics, Fluid Flow, Performance. Second
Edition, MIT Press, USA.
Tazerout M., Le Corre O. and Rousseau S. (1999). TDC determination in IC
engines based on the thermodynamic analysis of the temperatureentropy diagram. SAE paper 1999-01-1489.
317
Tazerout M., Le Corre O. and Rousseau S. (1999). TDC determination in IC
engines based on the thermodynamic analysis of the temperatureentropy diagram. SAE Paper 1999-01-1489.
The MathWorks, Inc. (2003) MATLAB: Reference guide, MA, USA.
Tsujimura T., Mikami S., Achiha N., Tokunaga Y., Senda J. and Fujimoto H.
(2003). A study of direct injection diesel engine fueled with
hydrogen, Proc. 2003 SAE World Congress, Detroit, Michigan, USA.
Tsujimura T., Nakatani K., Mikami S. and Okui N. (2001). Construction of
optimum injection and combustion for gaseous fuel engine. Nihon
Kikai Gakkai Nenji Taikai Koen Ronbunshu, Vol.2, pp. 557-558.
TÜV Süddeutschland (2004). Hydrogen - a world of energy. Munich.
US Department of Energy (2002). National Hydrogen Energy Roadmap. Proc.
National Hydrogen Energy Roadmap Workshop, Washington, DC, USA.
Van Blarigan P. and Keller J. (1998). A hydrogen fuelled internal
combustion engine designed for single speed/power operation. Int.
J. of Hydrogen Energy, Vol. 23, pp. 603-609.
Van Blarigan P. (2000). Advanced internal combustion engine research,
Proc. of 2000 DOE Hydrogen Program Review, San Ramon, California,
USA.
Van Der Drift A., Tjeng S., Beckers G. and Beesteheerde J. (1996). Low-NOx
hydrogen burner. International Journal of Hydrogen Energy, Vol. 21,
pp. 445-449.
Verhelst S. and Sierens R. (2001). Hydrogen engine-specific properties,
International Journal of Hydrogen Energy, Vol. 26, pp. 987-990.
Wakuri Y., Fugii M., Amitani T. and Tsuneya R. (1960). Studies of the
penetration of fuel spray in a diesel engine. Bulletin of JSME, Vol.3,
pp. 123-130.
Wang Y., Zeng S., Huang J., He Y., Huang X., Lin L. and Li S. (2005).
Experimental investigation of applying Miller cycle to reduce NOX
emission from diesel engine. Proc. Inst. Mech Eng., Part A - Journal
of Power and Energy, Vol.219, pp. 631-638.
Watson J. (2004). The Development of Large Technical Systems:
Implications for Hydrogen. Tyndall Centre for Climate Change
Research, Working Paper nº 18.
Watson N. (1984). Dynamic turbocharged diesel engine simulator for
electronic control system development. ASME Journal of Dynamic
Systems, Measurement, and Control, Vol. 106, pp. 27-45.
Welch A. and Wallace J. (1990). Performance characteristics of a hydrogenfueled diesel engine with ignition assist. Proc. SAE Int. Fuels and
Lubricants Meeting, Tulsa, Oklahoma, USA.
Wheeler A. and Ganjianji A. (1996). Introduction to Engineering
Experimentation. Prentice Hall, New Jersey, USA.
318
Winterbone D., Thiruarooran C. and Welltead P. (1977). A wholly dynamic
model of turbocharged diesel engine for transfer function
evaluation. SAE paper 770124.
Woud H. and Stapersma D. (2003). Design of Propulsion and Electric Power
Generation Systems. IMarEST, England.
Xu Y., Nishida K. and Hiroyasu H. (1992). A practical Calculation Method for
Injection Pressure and Spray Penetration in Diesel Engines, SAE
paper 920624.
Yamin J. and Badran O. (2002). Analytical study to minimise the heat losses
from propane powered 4-stroke spark ignition engine. Renewable
Energy, Vol. 27, pp. 453-478.
Yamin J., Gupta H., Bansal B. and Srivastava O. (2000). Effect of
combustion duration on the performance and emission
characteristics of a spark ignition engine using hydrogen as a fuel.
International Journal of Hydrogen Energy, Vol. 25, pp. 581-589.
Yi H., Min K. and Kim E. (2000). The optimised mixture formation for
hydrogen fuelled engines. International Journal of Hydrogen Energy,
Vol. 25, pp 685-690.
Zevenhoven R. (2001). Non-ideal gases in diesel engine processes. In: First
biannual Meeting and General Section Meeting of The ScandinavianNordic Section of the Combustion Institute, Gothenburg, Sweden, pp.
103-108.
Zhang Y. and Rizzoni G. (1993). An on-line indicated torque estimator for IC
engine diagnosis and control. ASME J. Advanced Automotive Tech.,
Vol. 52, pp. 147-162.
Zhenzhong Y., Jianquin W., Zhuoyi F. and Jinding L. (2002). An
investigation of optimum control of ignition timing and injection
system in a in cylinder injection type hydrogen fuelled engine. Int. J.
of Hydrogen Energy, pp. 213-217.
Züttel A. (2003). Hydrogen Storage Methods and Materials, Physics
Department, University of Fribourg, Switzerland.
Zweiri Y., Whidborne J. and Seneviratne L. (1999). Complete analytical
model of a single-cylinder Diesel engine for non-linear control and
estimation, Report EM/99/10, Department of Mechanical
Engineering, King’s College London, Strand, London, UK.
319
320
Appendix A:
Commercial dual fuel engine developments
In this appendix, a commercial task to convert two large-scale industrial
engines for dual fuel operation will be described. The work was based
around the technology for port injection of gaseous fuels in compression
ignition engines, described earlier in this thesis.
The author, through his company TecnoVeritas, was commissioned in 2007
to convert two 4.5MW diesel engines running on heavy fuel oil (HFO) to
dual fuel operation with natural gas. This included development of engine
monitoring and management systems, as well as the use of knock
identification to control the amount of natural gas used. During testing, it
was possible to obtain some experimental results which are presented
below, showing the performance of the engines under different modes of
operation.
A.1. Introduction
Increasing fuel costs and tightening environmental legislation drive an
interest among users of existing diesel engines to look at options to
improve fuel efficiency and reduce exhaust gas emissions formation. In
most places of the world, natural gas is continuously available at a
reasonable price, and its properties make it an interesting alternative fuel
for use in internal combustion engines. The option of converting existing
diesel engines to dual fuel capability gives more flexibility in the fuel
supply, but also potential reductions in operational costs and exhaust gas
emissions. However, the conversion of existing diesel engines to dual fuel
operation implies possible combustion problems that need to be addressed.
Natural gas as a fuel exhibits a number of interesting advantages such as a
reasonable net calorific value (approximately 38,500 kJ/Nm3), low
emissions of CO2, NOx and particulate matter, and generally a good knock
321
resistance (which is necessary to use it in diesel engines without engine
modifications). Where available, its supply is continuous and its price is
commercially appealing. When converting a standard diesel engine to dual
fuel operation, engine operational characteristics can remain largely
unchanged, with the liquid fuel acting as a pilot fuel to ignite the cylinder
charge. Therefore the quantity of pilot fuel must be set to release at least
the minimum ignition energy required to ignite the gaseous fuel, which can
constitute the main energy source of the working cycle. Due to the premixed charge, problems of pre-ignition and detonating combustion (knock)
may, however, occur, and the use of an appropriate engine control system
is therefore essential.
Numerous researchers have studied the use of natural gas as a second fuel
in diesel engines [1-9]. Generally, fuel efficiencies similar to those under
pure diesel engine operation are achieved at high engine loads. At low
loads dual fuel operation gives poorer fuel efficiency. Papagiannakis and
Hountalas [1,2] attribute this to the poor combustion of the gaseous fuel at
part load operation. With a reduction in the use of liquid diesel fuel, the
particulate matter (PM) emissions are significantly reduced; reductions of
above 50% are frequently reported. Also nitrogen oxides (NOx) can be
reduced in dual fuel mode, however increases are usually seen in the
emissions of carbon monoxide (CO) and unburned hydrocarbons (HC).
A.2. Dual fuel operation
The cylinder charge of dual fuel engines (converted existing diesel engines
or simply diesel derived engines) is usually made up of two fuels with
distinct ignition temperatures and different physical states, typically one
liquid and one gaseous. This gives a combustion process different from
those of conventional diesel or spark ignition engines, and gives the dual
fuel engine some particular operating characteristics which need to be
taken into account when working with such engines.
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A.2.1 The dual fuel combustion process
The combustion process in dual fuel engines is more complicated than that
of conventional engines since a combination of premixed and diffusion
combustion occur in this mode of engine operation. The contribution and
characteristics of each type of combustion depends on several parameters,
including fuel properties, injector characteristics, and combustion chamber
design, as well as operational variables such as the engine load, speed,
manifold air pressure and temperature, and the amount of each fuel
present in the combustion chamber.
The combustion process in a dual fuel engine can be divided into three
distinct sub-processes:

ignition of the pilot fuel;

combustion of the gaseous fuel which is in the vicinity of the pilot
fuel cores; and

combustion of the gaseous fuel due to flame propagation into the
premixed lean charge.
The equivalence ratio of the cylinder charge varies spatially from point to
point within the combustion chamber, since a fast homogenisation of the
mixture of natural gas and liquid fuel in most cases is not possible. At low
loads, the air-gas mixtures may be so lean that it causes flame propagation
interruptions, resulting in incomplete combustion. This will lead to loss of
fuel through the exhaust, contributing to a reduction in fuel efficiency and
high levels of unburned hydrocarbons in the exhaust.
A.2.2 Detonation
Detonation occurs during the combustion process when the burnt gas zone,
which is pressurising and heating the unburned part of the cylinder charge
ahead of the flame front, does so at such a rate that the unburned fuel
achieves its auto-ignition temperature before the arrival of the actual
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flame front. The result is that the unburned charge volume ignites
spontaneously over the entire zone where the auto-ignition conditions have
been achieved. The apparent flame speed in this zone is many orders of
magnitude faster than that of conventional combustion initiated by the
normal flame front, resulting in higher rates of pressure and temperature
rise. Knocking is usually associated to small portions of volume of the
combustion chamber, whereas detonation is associated to the entire
compressed cylinder charge.
Engine knock is directly related to the compression ratio, because the
higher the compression, the closer the charge will be to its autoignition
conditions. Moreover, the knock intensity will depend on the intake air
conditions, combustion chamber design and turbulence levels within the
cylinder, as well as the flame speed of the fuel-air mixture. Therefore, for
turbocharged dual fuel engines the temperature of the air entering the
cylinder and its pressure are factors that influence detonation sensitivity.
Detonation is accompanied by a drastic increase in temperature and
pressure within the cylinder, resulting in some cases in serious engine
damage.
In converted diesel engines, the quality of the fuel oil used is of utmost
importance to attain good control of the combustion process as well as a
smooth and clean combustion. Heavy fuel oils with bad ignition quality may
have long, and varying, ignition delays, leading to reduced combustion
efficiency and poorer utilisation of the gaseous fuel. The temperature of
the heavy fuel oil is also an important factor to consider, as viscosity and
therefore atomisation of the fuel play an fundamental role in the
homogenisation of the cylinder charge. If fuels of very low quality are used,
viscosity control may be necessary in order to maintain satisfactory engine
performance.
A.2.3 Pre-ignition
Another dangerous and damaging phenomenon of pre-mixed engine
operation is the pre-ignition of the cylinder charge. This phenomenon
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results from the igniting of the cylinder charge through contact with a high
temperature surface prior to the desired ignition timing (i.e. the time of
pilot fuel injection, or spark discharge in a spark ignited engine). Preignition is particularly dangerous since ignition can occur during the
compression stroke, leading to excessive mechanical stress and damages in
the crank system, piston rings, bearings, etc., as well as increased thermal
stress and detrimental effects on cylinder lubrication due to very high incylinder gas temperatures.
The main cause of such a phenomenon is the presence of carbonaceous
residues on the combustion chamber surface. With the use of low-quality
fuels such as heavy fuel oils, the amount of residues on the combustion
chamber components are likely to increase, and this must be taken into
consideration when converting an engine to dual fuel operation. The
ignition temperature of a natural gas-air mixture will depend on the fuelair ratio, and the fuel substitution ratio and natural gas concentration may
therefore be limited by pre-ignition tendencies of the cylinder charge.
Pre-ignition in a cylinder will also have a self-enhancing effect in that the
increased in-cylinder gas temperatures and pressures resulting from such an
occurrence will increase the temperatures of the combustion chamber
walls and thereby increase the probability of pre-ignition in the next cycle.
With the use of low-quality fuels in a dual fuel engine, continuous
monitoring and control of the combustion process to avoid pre-ignition is
therefore essential in order to maintain engine integrity and operational
stability.
A.3. Engine conversion
TecnoVeritas has performed engine conversions worldwide since 1999 for a
range of engine makes, sizes, and different fuels including dual fuel and
multi fuel using diesel oil, heavy fuel oil, natural gas, producer gas, and
hydrogen. A range of in-house developed technology solutions for engine
monitoring, control, and operational optimisation have been developed.
The following sections present the conversion of two Wärtsilä 9L32 diesel
325
engines to dual fuel operation on heavy fuel oil and natural gas, and
describe technological solutions chosen as well as operational data from
the engines.
In December 2008, TecnoVeritas finished the commissioning of two ninecylinder Wärtsilä type 9L32 diesel engines, each with approximately 96,000
hours of operation on heavy fuel oil (HFO). The objective was to convert
these two HFO engines to use as much natural gas as possible, without
creating mechanical stress higher than that created under normal HFO
operation. Simultaneously, the development of operation software with an
appropriate human interface was required. Figure 1 shows a photograph of
the two engines, and main engine design data are listed in Table 1.
Engine model
Wärtsilä 9L32
Number of cylinders
9
Cylinder bore
320 mm
Stroke length
350 mm
Speed
750 rpm (12.5 s^-1)
Mean effective pressure
21.3 bar (2130 kPa)
Mean piston speed
8.75 m/s
Rated power
4,450 kW
Boost pressure
2.4 bar (240 kPa)
Compression ratio
12
Table 1: Engine data
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Figure 1: The two Wärtsilä 9L32 diesel engines converted at ECE,
Portugal.
A.3.1 Fuel injection system
In order to allow dual fuel operation, modifications were made to the
intake and injection systems. Gas fuel injectors were fitted to the intake
manifold, with one injector per engine cylinder. The gas injection valves
are controlled by an injection controller receiving information from various
sensors and systems installed on the engine, from the gas regulating unit
and from the process controller. The original mechanical controller was
replaced by a hydraulic actuator, controlled by the main injection
controller, allowing seamless transfer from conventional to dual fuel mode
and vice versa. The injection of the correct quantity of gas during the
induction stroke, i.e. after the closing of the exhaust valve and before the
closing of the inlet valve, is continuously adjusted through the injection
angle and gas pressure, thereby allowing a fine adjustment of the gas
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energy supplied per cycle. The system allows for the control of the gas
injection rate individually for each cylinder, therefore offering a high
flexibility in the operation of the engine. The gas fuel injection is
continually controlled using the data from the KDS knock detection system
(described below) to allow immediate adjustments of the injection rate to
a given cylinder in the case of knock or misfiring.
A.3.2 Performance monitoring and control
Performance monitoring and engine control equipment for dual fuel
conversions has been developed by TecnoVeritas, including dedicated
hardware and software such as the VTec multi-point port injection system
for the conversion of existing diesel engines into true dual fuel engines, the
KDS knock detection system, and the EDS engine diagnosis system.
Prior to conversion, the Wärtsilä engines were surveyed and operating
variables such as pressures, temperatures, and specific fuel consumptions
were logged to be used as a reference. Based on the operational values,
the engineering team proceeded with the customised design of the system.
As a consequence of such a demanding contract, a comprehensive cylinder
pressure monitoring system was fitted to monitor the combustion cycles
during system tuning and normal operation. The VTec system is based on
individual cylinder gas port injection, therefore allowing the correction of
cylinder parameters such as temperature, maximum combustion pressure
and knocking, by controlling the quantity of gas being injected in each
cylinder per cycle.
The KDS knock detection system, connected via a CAN-bus network,
includes one accelerometer per cylinder, identifying in each cycle any
knock or the absence of combustion (misfiring). The KDS acceleration
sensors are incorporated into the cylinder head, and give information on
the vibration levels in the proximity of each cylinder. In addition, the KDS
system has two position transducers, one on the flywheel and a second on
the camshaft; both signal are acquired to exactly determine the position
and phase of the TDC of each cylinder. The KDS system uses an algorithm to
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perform a Fast Fourier Analysis at each cycle, therefore allowing the
identification of knocking frequencies an their maximum amplitude before
raising an alarm and output an demand to the injection controller for the
reduction of gas to a given cylinder, or even the switching of operation
from dual-fuel mode to conventional HFO or diesel-fuelled operation.
By using the information on the knock or pre-ignition intensity, the
injection timing and/or gas injection pressure can be adjusted for optimal
engine operation. Another function of the KDS system is to determine if any
misfiring occurs, for example in the case of a gas injector malfunctioning.
In such case, an alarm with indication of the defective cylinder is raised
and the engine immediately transferred from dual fuel to standard HFO
operation.
As the objective is to maximise the use of natural gas, the engine
management system will seek the highest possible substitution ratio of
natural gas to HFO. If the knocking or pre-ignition intensity increases, the
gas flow is reduced after a number of cycles (typically 5). Should the
knocking or pre-ignition continue beyond pre-set limits, the gas will be
switched off and the engine operation is transferred seamlessly to HFO or
diesel. Hence, the use of natural gas is maximised, while taking into
account the varying conditions in the engine, in particular in relation to
carbon deposits which promote pre-ignition of the cylinder charge.
The use of a dual fuel system operated with HFO and natural gas requires
particular attention on the injection equipment condition. For this reason,
both engines were equipped with the EDS engine diagnosis system, which
allows the monitoring of the combustion process in each individual
cylinder. In addition to detecting any injection valve malfunction, the
injection controller is capable of making corrections and optimising engine
operation based on operational variables such as the engine manifold
pressure, exhaust gas temperatures, etc.
329
Figure 2: Graphical user interface of the engine diagnostics and performance
monitoring system.
Figure 2 shows the graphical user interface of the engine management
system. It provides the operator with vital engine performance variables,
including boost air pressure and temperature, natural gas properties,
exhaust gas temperatures, knock intensity, and electric output, as well as
fuel consumption and fuel substitution ratio. The operator can decide the
substitution ratio set-point, as well as switching the engine back to
conventional diesel or HFO operation.
A.4. Experimental results and engine performance
After successful conversion, some performance tests were done to study
engine fuel consumption and exhaust gas emissions under different modes
of operation. As described above, the engines were run with the highest
possible substitution ratio of natural gas to heavy fuel oil, limited by the
knocking intensity in the cylinders.
330
A.4.1 The combustion process
Figure 3 shows in-cylinder pressure traces obtained during the tests. The
graphs show pressure plots for operation on heavy fuel oil only, normal dual
fuel operation, and for dual fuel operation with knocking due to too high
natural gas substitution ratio. At normal dual fuel operation, the fuel
composition is approximately 70% natural gas and 30% HFO (on an energy
basis), and it can be seen that the performance closely resembles that of
normal,
HFO-fuelled
operation.
For
higher
substitution
ratios,
a
significantly faster pressure rise and higher peak pressure can be seen,
illustrating the need for the knock detection system and appropriate gas
injection control.
Figure 3: In-cylinder pressure plots for different engine operating modes.
Figures 4 and 5 further illustrate this, showing the calculated net heat
release rates from combustion, derived from the pressure plots. Again,
with an appropriate substitution ratio, the operation in dual fuel mode
differs only very little from that on heavy fuel oil. However, with too high
331
levels of natural gas, characteristic knocking behaviour is observed, with
detonating combustion at a crank angle of around 190. The knock behaviour
leads to pressure waves in the combustion chamber, compromising the
measurements. Since the heat release rate is calculated from the pressure
data, one gets oscillations as can be seen in the graph.
Figure 4: Net heat release rates.
332
Figure 5: Accumulated heat release in dual fuel mode with and without knock.
A.4.2 Exhaust gas emissions formation
Regarding emissions formation, natural gas has a number of advantageous
features compared with diesel oil or heavy fuel oil. First, the amount of
CO2 produced per unit energy delivered is lower than that of more complex
hydrocarbons (with a higher carbon to hydrogen ratio). In the tests, a
reduction in CO2 emissions of approximately 16% was obtained in dual fuel
mode compared to that of HFO operation. Second, particulate matter
emissions (PM), an issue of great concern in diesel engines, dropped by 50%
compared to the exhaust emissions at the same load on HFO, due to the
replacing of a large fraction of HFO with natural gas, which produces
negligible PM emissions. Finally, NOx emissions dropped by 10%, which is
somewhat lower than the NOx reductions reported by other authors, such as
Mustafi and Raine [9]. This is probably due to the fact that the engines
considered here are significantly larger, and therefore have a lower
operating speed, than most systems described in the literature.
333
A.4.3 Fuel cost and financial viability
The pay-back period of the present conversion project, based on the
achieved substitution ratio and the price of natural gas and HFO at the time
of commissioning (Oct. 2008), was estimated to be less than one year. For
diesel oil operated engines, the payback period can be even shorter; diesel
oil is more expensive than HFO and higher substitution ratios can be
achieved (above 90% has been demonstrated), while maintaining the same
power and slightly reducing the exhaust gas temperatures.
A.5 Summary
A commercial job to convert two industrial, large-bore diesel engines for
operation on heavy fuel oil and natural gas was described. The technical
solutions for engine conversion were outlined, including the fuel injection
system, performance monitoring system, and knock detection system.
Experimental results were presented showing the performance under
standard heavy fuel oil operation and dual fuel operation.
Testing showed that a high substitution ratio could be achieved with
natural gas supplying 70% of the fuel energy under dual fuel operation.
Knocking behaviour was observed for too high substitution ratios,
demonstrating the need for a knock detection system to allow optimised
engine operation and maximised substitution ratio. Significant reduction in
exhaust gas emissions, including NOx, particulates, and CO2 were found
under dual fuel mode compared with conventional heavy fuel oil based
operation.
A.6. References
[1] R. G. Papagiannakis, D. T. Hountalas. Experimental investigation
concerning the effect of natural gas percentage on performance and
334
emissions of a DI dual fuel diesel engine. Applied Thermal Engineering,
Volume 23, Issue 3, February 2003, Pages 353-365.
[2] R. G. Papagiannakis, D. T. Hountalas. Combustion and exhaust emission
characteristics of a dual fuel compression ignition engine operated with
pilot Diesel fuel and natural gas. Energy Conversion and Management,
Volume 45, Issues 18-19, November 2004, Pages 2971-2987.
[3] R. G. Papagiannakis, D. T. Hountalas, P. N. Kotsiopoulos. Experimental
and Theoretical Analysis of the Combustion and Pollutants Formation
Mechanisms in Dual Fuel DI Diesel Engines. SAE Paper 2005-01-1726, 2005.
[4] A.P. Carlucci, A. de Risi, D. Laforgia, F. Naccarato. Experimental
investigation and combustion analysis of a direct injection dual-fuel diesel
natural gas engine. Energy, Volume 33, Issue 2, February 2008, Pages 256263.
[5] M. Y. E. Selim. Pressure–time characteristics in diesel engine fueled
with natural gas. Renewable Energy, Volume 22, Issue 4, April 2001, Pages
473-489.
[6] M. Mbarawa, Brain Edward Milton, Robert Thomas Casey. Experiments
and modelling of natural gas combustion ignited by a pilot diesel fuel spray.
International Journal of Thermal Sciences, Volume 40, Issue 10, 2001,
Pages 927-936.
[7] V. Balasubramanian, K. Sridhara, V. Ganesan. Performance Evaluation
of a Small Agricultural Engine Operated on Dual Fuel (Diesel + Natural Gas)
System. SAE Paper 951777, 1995.
335
[8] R. Papagiannakis, D. Hountalas, C. Rakopoulos. Combustion and
Performance Characteristics of a HSDI Diesel Engine Operating from Low to
High Natural Gas Supplement Ratios at Various Operating Conditions. SAE
Paper 2008-01-1392, 2008.
[9] N. N. Mustafi, R. R. Raine. A Study of the Emissions of a Dual Fuel
Engine Operating with Alternative Gaseous Fuels. SAE Paper 2008-01-1394,
2008.
336
Appendix B:
Development of a dual fuel combined heat and power
research facility
This appendix describes a project to develop a dual fuelled hydrogen and
bio-oil combined heat and power (CHP) system for as a research tool. The
project was carried out as a cooperation between Newcastle University and
TecnoVeritas, and funded by Carbon Connections. TecnoVeritas contributed
the technology for use of gaseous fuels in CI engines, including engine
monitoring and management systems, developed during the course of the
PhD work described in this thesis.
B.1 Introduction
The use of unprocessed bio-oil with hydrogen from renewable sources in a
dual fuel CHP system potentially allows power and heat generation with
near-zero carbon emissions. The known challenges associated with the use
of unprocessed bio-oils in diesel engines, such as their poor ignition and
combustion properties, may be eased by the use of hydrogen in a dual fuel
system, in which the hydrogen will act as a combustion improver.
The objectives of the project were to develop a research tool to allow wide
ranging investigations of the potential of such a system. This included the
construction of an engine system with heat recovery from the exhaust and
engine cooling circuit, development of a bio-oil viscosity control and supply
system, implementation of a timed hydrogen injection system, and the
design of a flexible engine management, performance monitoring and data
acquisition system.
The CHP plant construction and commissioning was completed in March
2009. Preliminary system testing was undertaken in April 2009. These tests
337
were carried out using sunflower oil and rapeseed oil as liquid fuels, as well
as conventional diesel in order to provide a baseline for comparison.
Hydrogen was used as gaseous fuel, but operation of the system with
butane gas was also demonstrated to show the flexibility of the system.
The choice in liquid and gaseous fuels allows studies of CHP system
feasibility in a range of applications, including, for example, the use of
low-calorific gaseous fuels (pyrolysis gas, landfill gas etc.) or different
blends of bio-oils, as well as blends with fossil diesel.
B.2 System description
The system is built in a standard 40-foot freight container, which is divided
into three sections: a fuel storage room, a control room, and an engine
room. Figure 1 shows the outside view of the container.
Figure 1: Outside view of Dual Fuelled CHP system container.
338
B.2.1 Engine and control system
The Dual fuelled CHP system is based around a Deutz turbocharged
industrial diesel engine with a maximum power output of 45kW. The engine
is coupled to an electric generator to produce electric power, and heat
exchangers allow utilisation of the excess heat normally lost to the exhaust
gases and cooling system.
Figure 2: SCADA display of engine system.
Figure 2 shows the SCADA (user interface and control system) display of the
engine system and shows how it is controlled. The system is fully
instrumented with temperature and pressure sensors to allow online
readings of the performance during operation. All operational variables of
the engine are controlled from the control software, such as starts/stop,
load level, and fuel substitution ratio (i.e. percentage of liquid fuel to
percentage gaseous fuel). Figure 3 shows a photograph of the engine room.
In the front of the engine is the electric generator. Above the engine, the
exhaust line can be seen, leading to the exhaust gas heat exchanger, which
339
can be seen in the top of the photograph. On the right hand side, the
engine cooling circuit heat exchanger is seen. The temperature sensors are
visible, which, together with a flow sensor, allow calculation of the heat
flow rejected to this cooler.
Figure 3: View of the CHP system engine room.
340
B.2.2 Fuel supply system
A fuel supply system allowing the use of three different liquid fuels as well
as gaseous fuel has been implemented. The liquid fuel supply system is
illustrated in Figure 4. It allows for the use of a single fuel from one of the
three tanks or the use of a mixture of two fuels at any composition. In
mixed fuel mode, the operator sets the relative fuel flow from each of the
two selected tanks (e.g. 20% fuel A and 80% fuel B), as shown in Figure 4.
Prior to being supplied to the engine, the fuel is passed through a blending
tank to ensure proper mixing. The blending tank includes a recirculation
pump which passes the fuel through a homogenising unit and returns it to
the tank. This allows testing of combination of fuels in which separation
can be a problem.
Figure 4: SCADA display of liquid fuel supply system.
Photographs of the fuel storage room are shown in Figures 5 and 6. In
Figure 5, the three fuel tanks are shown, one of which is used for
conventional diesel fuel (to provide a baseline for comparison during
341
testing and to purge the fuel lines if necessary) whereas two are used for
different bio-oils. The fuel storage room also contains a water tank for the
heat recovery system. On the floor, the fuel pumps controlling the fuel
flow, and thereby the mixing of fuel from the different tanks, can be seen.
Figure 6 shows the blending tank, with the homogenising unit (in the front)
which ensures that any fuel mixture is properly blended before being
supplied to the engine.
Figure 5: View of fuel storage room.
342
Figure 6: View of fuel mixing tank.
The gaseous fuel injection system receives supply gas, for example
hydrogen, from a storage system outside the container for safety reasons.
(Other types of gaseous fuels can also be used, such as natural gas or
butane.) Fuel injectors (one for each cylinder) are fitted to the inlet
manifold and the injection of gaseous fuel is electronically controlled.
Based on the crank angle position reading, the system allows for timed
injection of gas into the pipe between the inlet manifold and each cylinder
during the intake stroke, to avoid build-up of hydrogen gas in the inlet
manifold. Figure 7 shows the gas injection system on the inlet manifold
343
with the lines supplying the four individual cylinders. The gas supply (from
the outside storage system) is the pipe seen on the right hand side of the
unit, while in the front the four gas valves are seen with their individual
power and control signal cables.
Figure 7: Engine gas injection system.
The SCADA user interface and control system allows the operator to
determine the substitution ratio of gaseous to liquid fuel, e.g. operation on
70% bio-oil and 30% hydrogen (all ratios are calculated on an energy basis),
in which the bio-oil can be any mixture from the fuel oil tanks as described
above. This gives very powerful control in the operational optimisation of
the system, in that the fuel composition can be suited for any operating
conditions (e.g. load level) and performance targets (e.g. emissions and
efficiency targets). In order to avoid engine damage due to knock or preignition at high gas substitution ratios, a knock detection system, as
described in Appendix A, is implemented. Knock is detected using
acceleration sensors fitted on the cylinder head of each cylinder. The
344
sensors measure any vibrations due to detonation during combustion, and
the signal is processed using Fast Fourier Transform to obtain a knock
intensity variable in the engine control system. If the knock intensity
exceeds a pre-set limit, the substitution ratio of gaseous to liquid fuel is
reduced. This safety feature ensures that no engine damage (e.g. breakage
of piston rings) occurs if too high gas substitution is demanded by the
operator.
Figure 8 shows the view from the engine control room into the engine
room.
Figure 8: View from engine control room.
B.2.3 Heat recovery
The heat recovery system allows recovery of the exhaust gas heat and the
heat lost to the engine cooling system, to achieve a total system efficiency
of above 80% at full load. Heat is recovered in an exhaust gas heat
345
exchanger located in the exhaust line and in an oil heat exchanger
recovering heat lost to the engine cooling system. Figure 9 shows the
layout of the heat recovery system. The system is fully instrumented with
temperature and flow sensors to allow calculation and logging of the heat
flow and recovery rates. Cooling water for the circuit is supplied from a
storage tank (left hand side of the figure), and the recovered heat is
currently dumped using fan coolers located on the container roof.
Figure 9: SCADA display of the fuel pre-heating and heat recovery
system.
B.2.4 Viscosity control system
A viscosity control system for the liquid fuel tanks is implemented, using
heat from the heat recovery circuit to control the temperature and
viscosity in the fuel storage tanks and in the blending tank. This is
necessary to maintain satisfactory combustion properties of highly viscous
fuels. The viscosity control system is automatically controlled, with the set
points provided by the operator in the control software. Viscosity
346
measurement equipment (viscometer) is provided in the fuel storage room
to allow manual tests of the properties of the different fuels.
B.2.5 Hydrogen-related safety measures
As the project is based on the use of hydrogen, compliance with ATEX /
DSEAR safety standards is required. Therefore the use of extra safety
equipment, such as Zener barriers (galvanic isolators), hydrogen leak
detectors, and ATEX hydrogen valves in the demonstration CHP system was
required.
B.3 Preliminary test results
The hydrogen/bio-oil CHP plant has a maximum continuous electric power
output of 45 kW, and a total efficiency above 80%. (This excludes the waste
heat recovered and used for the controlled heating of the bio-oil storage
tanks.) The inline blending system and homogenisation system allows a
blend of various bio-oils to be used and its quality optimised. With the
construction of the test CHP system, it is possible to study the influence of
the hydrogen as a combustion improver, aiding the combustion of bio-oils
and contributing to engine overall efficiency improvements and emissions
reductions.
B.3.1 Combustion and emissions formation
Preliminary results demonstrated that hydrogen is an excellent combustion
improver, and that only a small quantity of this gas significantly reduces
the ignition delay of the bio-oil combustion. The effect of hydrogen on the
combustion includes an engine performance improvement as well as a
lowering of the thermal NOx formation and particulates emissions as the
fuel is combusted with a higher efficiency and the time at which the
cylinder charge is exposed to high temperatures is reduced. It was further
identified that the hydrogen percentage influences the heat loss
characteristics, with a high hydrogen percentage giving higher heat losses
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through the exhaust and lower heat losses through the combustion chamber
walls. This benefits overall system performance, as the use of an
appropriate exhaust gas economiser allows more efficient heat recovery
when compared with recovery of heat lost to the cooling system, since the
exhaust gases will have a higher temperature.
Tests have been run under a range of operating conditions with encouraging
results. Comparing operation on liquid fuel only with a substitution of 50%
hydrogen (on an energy basis), the following effects were observed:
 The formation of nitrogen oxides (NOx) decreased by approximately 60%,
which is due to the lower peak gas temperatures in the cylinder. A more
homogeneous fuel-air mixture reduces the high-temperature zones in the
burning fuel spray, in which NOx is predominantly formed.
 Formation of particulate matter decreases by approximately 20% with
the introduction of hydrogen, indicating that hydrogen enhances the
combustion of the liquid fuel and the oxidation of carbonaceous material.
B.3.2 Heat recovery
While the engine is operating at full load (45kW electric power), the waste
heat recovered from the engine block and lubricating oil cooling system
amounts to 21.6 kW, and from the exhaust gas the heat recovery reached
32.8 kW. In terms of overall energy efficiency, the value reached during
the first tests was 79.6%.
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