Introduction to
`
`1
`
`'
`
`Internal
`
`Combustion
`
`Engines
`
`\
`
`PAICE 2024
`1
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`BMW v. Paice
`|PR2020-01386
`
`PAICE 2024
`BMW v. Paice
`IPR2020-01386
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`1
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`

`

`M ”
`
`(u Richard Stone 1985, 1992, 1999
`
`All rights reserved. No reproduction. copy or transmission oi this
`publication may be made without written permission
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`transmitted save with written permission or in accordance with the'
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`provisions of the Copyright Designs and Patents Act 1988 or under
`the terms oi any licence permitting limited copying issued by the"'-“
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`London W1? 9HE.
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`Any person who does any unauthorised act in relation to this
`publication may be liable to criminal prosecution and civil claims
`for damages.
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`The author has asserted his right to be identified as
`the author of this work in accordance with the Copyright,
`Designs and Patents Act 1988.
`
`First published 1985
`Second edition 1992
`Reprinted six times
`Third edition 1999
`MACMILLAN PRESS LTD
`Houndmills, Basingstoke, Hampshire R621 6X5
`and London
`Companies and representatives throughout the world
`
`ISBN 0-333—74013—0
`
`A catalogue record for this book is available from the British Library.
`
`This book is printed on paper suitable for recycling and made from
`iully managed and sustained forest sources.
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`Typset by EXPO Holdings, Malaysia
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`Printed in Great Britain by
`Antony Rowe Ltd
`Chippenham, Wiltshire
`
`2
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`

`

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`5%
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`Contents
`
`am
`
`
`
`Preface to the Third Edition
`Acknowledgements
`Notation
`
`1
`
`introduction
`1.1
`Fundamental operating principles
`1.2
`Early internal combustion engine development
`1.3
`Characteristics of internal combustion engines
`1.4
`Additional types of internal combustion engine
`1.4.1
`The Wankel engine
`1.4.2
`Stratified charge engines
`Prospects for internal combustion engines
`Fuel cells
`Question
`
`1.5
`1.6
`17
`
`2 Thermodynamic principles
`2.1
`introduction and definitions of efficiency
`2.2
`ideal air standard cycles
`2.2.1
`The ideal air standard Otto cycle
`2.2.2
`The ideal air standard Diesel cycle
`2.2.3
`The ideal air standard Dual cycle
`2.2.4
`The ideal air standard Atkinson cycle
`Comparison between thermodynamic and mechanical cycles
`Additional performance parameters for internal combustion engines
`Fuel—air cycle
`Computer models
`Conclusions
`
`2.3
`2.4
`2.5
`2.6
`2.7
`
`2.8
`2.9
`
`Examples
`Questions
`
`3 Combustion and fuels
`3.1
`introduction
`
`Combustion chemistry and fuel chemistry
`3.2
`Combustion thermodynamics
`3.3
`3.3a Use of conventional thermodynamic tabulations
`3.3b Use of thermodynamic tabulations in Appendix A
`3.3.1
`The effect of the state of the reactants and products on the
`calorific values
`Dissociation
`
`3.4
`
`3.4.1
`
`Calculation of the equilibrium combustion temperature and pressure
`
`xi
`xiii
`XV
`
`1
`1
`6
`10
`14
`14
`15
`16
`19
`21
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`22
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`25
`25
`26
`29
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`35
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`41
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`42
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`59
`62
`65
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`68
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`3
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`

`
`
`3.5
`
`3.6
`3.7
`
`3.8
`
`Pre-mixed combustion in spark ignition engines
`3.5.1
`Normal combustion
`3.5.2
`Abnormal combustion
`
`Combustion in compression ignition engines
`Fuels and additives
`
`3.7.1
`3.7.2
`
`3.7.3
`3.7.4
`3.7.5
`
`Characteristics of petrol
`In-vehicle performance of fuels, and the potential of
`alcohols
`
`Characteristics of diesel fuel
`Diesel fuel additives
`Alternative diesel fuels
`
`..
`
`‘
`
`. ,. yr
`
`,_
`
`,
`
`Engine emissions and hydrocarbon oxidation
`3.8.1
`Introduction
`3.8.2
`Nitric oxide formation
`
`3.8.3
`3.8.4
`
`Hydrocarbon oxidation
`Carbon monoxide emissions
`
`3.9
`
`Combustion modelling
`3.9.1
`Introduction
`3.9.2
`Zero-dimensional models
`3.9.3
`Quasi-dimensional models
`3.9.4 Multidimensional models
`3.10 Conclusions
`3.11
`Examples
`3.12 Questions
`
`Spark ignition engines
`4.1
`Introduction
`4.2
`Combustion chambers
`4.2.1
`Conventional combustion chambers
`
`4.2.2
`
`High compression ratio combustion chambers and fast burn
`combustion systems
`Advanced combustion systems
`4.2.3
`Direct injection stratified charge engines
`4.2.4
`Catalysts and emissions from spark ignition engines
`4.3.1
`introduction
`
`Development of threeway catalysts
`4.3.2
`Lean-burn NO. reducing catalysts
`4.3.3
`Emissions legislation trends
`4.3.4
`Cycle-by-cycle variations in combustion
`Ignition systems
`4.5.1
`Ignition system overview
`4.5.2
`The ignition process
`Mixture preparation
`4.6.1
`Introduction
`
`Variable jet carburettor
`4.6.2
`Fixed iet carburettor
`4.6.3
`Fuel injection
`4.6.4
`4.6.5 Mixture preparation
`Electronic control of engines
`Conclusions
`
`4.3
`
`4.4
`4.5
`
`4.6
`
`4.7
`4.8
`
`71
`71
`74
`75
`77
`77
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`86
`90
`94
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`98
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`101
`106
`109
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`110
`112
`113
`113
`114
`133
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`142
`142
`147
`147
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`152
`155
`162
`164
`164
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`175
`176
`181
`184
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`190
`192
`192
`194
`195
`198
`202
`207
`211
`
`4
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`

`

`4.9
`
`Example
`4.10 Questions
`
`Compression ignition engines
`5.1
`Introduction
`5.2
`5.3
`5.4
`5.5
`
`Direct injection (DI) systems
`Indirect injection (IDI) systems
`Cold starting of compression ignition engines
`Fuel injection equipment
`5.5.1
`injection system overview
`5.5.2
`Fuel injectors
`5.5.3
`Traditional injection pumps
`5.5.4
`Interconnection of traditional pumps and injectors
`5.5.5
`Common rail and electronic unit injector systems
`Diesel engine emissions
`5.6.1
`Emissions legislation
`5.6.2
`Sources and control of engine out emissions
`5.6.3
`After treatment of diesel emissions
`Conclusions
`
`5.6
`
`5.7
`5.8
`5.9
`
`6.3
`6.4
`
`6.5
`6.6
`
`6.7
`6.8
`6.9
`
`Example
`Questions
`
`Indiiction and Exhaust Processes
`6.1
`Introduction
`6.2
`
`Valve gear
`6.2.1
`Valve types
`6.2.2
`Valve-operating systems
`6.2.3
`Dynamic behaviour of valve gear
`Flow characteristics of poppet valves
`Valve timing
`6.4.1
`Effects of valve timing
`6.4.2
`Variable valve timing
`Unsteady compressible fluid flow
`Manifold design
`6.6.1
`Geneml principles
`6.6.2
`Acoustic modelling techniques
`
`Silencing
`Conclusions
`Questions
`
`—
`
`Two-stroke engines
`7.1
`Introduction
`7.2
`7.3
`7.4
`
`Two-stroke gas flow performance parameters
`Scavenging systems
`Scavenge modelling
`7.4.1
`Perfect displacement scavenging model
`7.4.2
`Perfect mixing scavenging model
`7.4.3
`Complex scavenging models
`Experimental techniques for evaluating scavenge and results for port
`flow coefficients
`
`7.5
`
`Contents agé vii
`
`212
`213
`
`216
`216
`219
`224
`230
`231
`232
`234
`241
`246
`250
`256
`256
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`264
`265
`268
`269
`
`272
`272
`273
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`276
`285
`293
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`297
`301
`309
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`315
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`320
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`325
`326
`326
`327
`329
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`331
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`5
`
`

`

` Contents
`
`7.5.1
`7.5.2
`7.5.3
`
`Firing engine tests
`Non-firing engine tests
`Port flow characteristics
`
`7.6
`7.7
`7.8
`
`Engine performance and technology
`Concluding remarks
`Questions
`
`In-cylinder motion and turbulent combustion
`8.1
`introduction
`
`8.2
`
`8.3
`
`8.4
`
`8.5
`8.6
`
`Flow measurement techniques
`8.2.1
`Background ‘
`8.2.2
`Hot wire anemometry
`8.2.3
`Laser Doppler anemometry
`8.2.4
`Particle Image Velocimetry
`8.2.5
`Comparison of anemometry techniques
`Turbulence
`8.3.1
`Turbulence definitions
`
`In-cylinder turbulence
`8.3.2
`Turbulent combustion modelling
`8.4.1
`A turbulent entrainment model of combustion
`
`Laminar burning velocities
`8.4.2
`The effect of turbulence on flame behaviour
`8.4.3
`Conclusions
`Questions
`
`Turbocharging
`9.1
`Introduction
`9.2
`Radial flow and axial flow machines
`9.2.1
`Introduction and fluid mechanics
`
`Thermodynamics of turbochargers
`9.2.2
`Turbocharging the compression ignition engine
`Turbocharging the spark ignition engine
`Practical considerations and systems
`9.5.1
`Transient response
`9.5.2
`Variable-geometry turbochargers, superchargers and two-stage
`turbocharging
`Conclusions
`
`Examples
`Questions
`
`9.3
`94
`9.5
`
`9.6
`
`9.7
`98
`
`10
`
`Engine modelling
`10.1
`Introduction
`
`10.2 Zero-dimensional modelling
`10.2.1 Thermodynamics
`10.2.2 Gas properties
`10.2.3 Burn rate
`
`10.2.4 Engine gas side heat transfer
`10.2.5
`Induction and exhaust processes
`10.2.6 Engine friction
`
`332
`333
`334
`335
`340
`343
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`345
`' 345
`346
`346
`347
`349
`351
`351
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`352
`357
`361
`361
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`387
`393
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`419
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`429
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`436
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`6
`
`

`

`
`Contents _
`
`10.3 Application of modelling to a turbocharged medium-speed
`diesel engine
`10.3.1
`Introduction
`
`10.3.2 Building and validating the model
`10.3.3 The effect of valve overlap on engine operation
`10.4 Conclusions
`
`11 Mechanical design considerations
`11.1
`Introduction
`
`11.2 The disposition and number of the cylinders
`11.3 Cylinder block and head materials
`11.4 The piston and rings
`11.5 The connecting-rod, crankshaft, camshaft and valves
`11.6 Lubrication and bearings
`11.6.1
`Lubrication
`11.6.2 Bearing materials
`11.7 Advanced design concepts
`11.8 Conclusions
`1 1 .9 Questions
`
`12 Heat transfer in internal combustion engines
`12.1
`ln-cylinder heat transfer
`12.2 Engine cooling
`12.2.1 Background
`12.2.2
`Spark ignition engines
`
`12.2.3 Compression ignition engines
`12.3 Liquid coolant systems
`12.3.1 Conventional coolant systems
`12.3.2 Cooling media performance
`12.3.3 Advanced cooling concepts
`12.4 Conclusions
`
`13 Experimental facilities
`13.1
`Introduction
`13.2 Quasi-steady engine instrumentation
`13.2.1 Dynamometers
`13.2.2 Fuel-consumption measurement
`13.2.3 Airflow rate
`13.2.4 Temperature and pressure
`13.2.5
`In-cylinder pressure measurement
`13.2.6 Techniques for estimating indicated power
`13.2.7 Engine test conditions
`13.2.8 Energy balance
`13.3 Experimental accuracy
`1 3.4 Measurement of exhaust emissions
`13.4.1
`Infra-red absorption
`13.4.2
`Flame ionisation detection (no)
`13.4.3 Chemiluminescence
`13.4.4 Oxygen and air/fuel ratio analysers
`13.4.5 Exhaust smoke and particulates
`
`‘
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`471
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`482
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`518
`519
`521
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`526
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`533
`
`7
`
`

`

` Contents
`
`13.4.6 Determination of EUR and exhaust residual (ER) levels
`13.4.7 Determination of the air/fuel ratio from exhaust emissions
`13.5 Computer—based combustion analysis
`13.5.1
`Introduction
`
`13.5.2 Burn rate analysis
`13.5.3 Heat release analysis
`13.6 Advanced test systems
`13.7 Conclusions
`13.8 Question
`
`14 Case studies
`14.1
`Introduction
`
`.
`
`L
`
`
`
`f
`
`14.2 The RoverKseries engine
`14.2.1
`Introduction
`
`v.
`
`.‘1:-r‘r.’::'
`
`"mg-event».
`
`‘
`
`~1...:
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`14.2.2 The K16 base engine design
`14.2.3 The Rover K series engine coolant and lubrication circuits
`14.2.4 The Rover K series engine manifolding, fuelling and ..~':1
`ignition systems
`14.2.5 The Rover K series engine performance development
`14.2.6 Concluding remarks
`14.3 Chrysler 2.2 litre spark ignition engine
`14.3.1 Background
`14.3.2 The cylinder head
`14.3.3 The cylinder block and associated components
`14.3.4 Combustion control
`
`,
`
`~-
`
`14.4
`
`14.3.5 Engine development
`Ford 2.5 litre DI diesel engine
`14.4.1 Background
`14.4.2 Description
`14.4.3 Combustion system
`14.4.4 Turbocharged engine development
`
`Appendix A: Thermodynamic data
`Appendix B: Answers to numerical problems
`Appendix C: The use of SI units
`Bibliography
`References
`Index
`
`535
`537
`541
`541
`544
`547
`550
`553
`
`554
`
`'555
`555
`555
`555
`556
`558
`
`562
`562
`564
`564
`564
`566
`568
`569
`$70
`570
`570
`571
`574
`577
`
`582
`612
`614
`617
`619
`635
`
`8
`
`

`

`Turbocharging
`
`393
`
`speed reduces the number of gearbox ratios that are needed for starting and hill
`climbing. However,
`if the turbocharger is matched to give high torque at low
`speeds,
`then at high speeds the pressure ratio will be too great, and the
`turbocharger may also over-speed. This problem is particularly severe on passenger
`car engines and an exhaust by-pass valve (waste-gate) is often used. The by-pass
`valve is spring regulated and, at high flow rates when the pressure rises, it allows
`some exhaust to by-pass the turbine, thus limiting the compressor pressure ratio.
`Turbocharging is particularly popular for automotive applications since it enables
`smaller, lighter and more compact power units to be used. This is essential in cars if
`the performance of a compression ignition engine is to approach that of a spark
`ignition engine. In trucks the advantages are even greater. With a lighter engine in
`a vehicle that has a gross weight limit, the payload can be increased. Also, when
`the vehicle is empty the weight is reduced and the vehicle fuel consumption is
`improved. The specific fuel consumption of a turbocharged compression ignition
`engine is better than that for a naturally aspirated engine, but additional gains can
`be made by retuning the engine. If the maximum torque occurs at an even lower
`engine speed, the mechanical losses in the engine will be reduced and the specific
`fuel consumption will be further improved. However, the gearing will then have to
`be changed to ensure that the minimum specific fuel consumption occurs at the
`normal operating point. Ford (1982) claim that turbocharging can reduce the
`weight of truck engines by 30 per cent, and improve the specific fuel consumption
`by from 4 to 16 per cent. Figure 9.20 shows a comparison of naturally aspirated
`and turbocharged truck engines of equivalent power outputs.
`In passenger cars a turbocharged compression ignition engine can offer a
`performance approaching that of a comparably sized spark ignition engine; its
`torque will be greater but its maximum speed lower. Compression ignition engines
`can give a better fuel consumption than spark ignition engined vehicles, but this
`will depend on the driving pattern (Radermacher, 1982) and whether the
`comparison uses a volumetric or gravimetric basis (see chapter 3, section 3.7).
`
`9.4 g Turbocharging the spark ignition engine
`
`Turbocharging the spark ignition engine is more difficult than turbocharging the
`compression ignition engine. The material from the previous section applies, but in
`addition spark ignition engines require a wider air flow range (owing to a wider
`speed range and throttling), a faster response, and more careful control to avoid
`either pre-ignition or self-ignition (knock). The fuel economy of a spark ignition
`engine is not necessarily improved by turbocharging. To avoid both knock and self-
`ignition it is common practice to lower the compression ratio, thus lowering the
`cycle efficiency. This may or may not be offset by the frictional losses representing
`a smaller fraction of the engine output.
`The turbocharger raises the temperature and pressure at inlet to the spark
`ignition engine, and consequently pressures and temperatures are raised
`throughout the ensuing processes. The effect of inlet pressure and temperature
`on the knock-limited operation of an engine running at constant speed, with a
`constant compression ratio, is shown in figure 9.21. Higher octane fuels and rich
`mixtures both permit operation with higher boost pressures and temperatures.
`Retarding the ignition timing will reduce the peak pressures and temperatures to
`provide further control on knock. Unfortunately there will be a trade-off in power
`
`9
`
`

`

`
`
`Figure 9.20
`Comparison of comparably
`powerful naturally
`aspirated and
`turbocharged engines
`(Ford, 1982).
`
`
`
`Enginetorque
`
`Specific fuel
`consumption contours
`(g/kWh)
`
`
`
`
`Engine sis-eed
`Naturally aspirated 8-cylinder Diesel engine
`
`
`
`Enginetorque
`
`
`
`Engine speed
`
`Specific fuel
`
`consumption
`Con‘OUrs
`(g/kWhl
`
`
`
`Turbocharged 6—cylinder Diesel engine
`
`and economy and the exhaust temperature will be higher; this can cause problems
`with increased heat
`transfer in the engine and turbocharger. Reducing the
`compression ratio is the commonest way of inhibiting knock and retarding the
`ignition is used to ensure knock—tree operation under all conditions.
`
`Figure 9.21
`Influence of charge
`temperature on charge
`pressure (knock-limited)
`with different air/fuel ratios
`and fuel qualities (with
`acknowledgement to
`Watson and Janota, 1982).
`
`
`
`
`
`Boostpressure(bari
`
`1.9
`
`1.8
`
`1.7
`
`1.6
`
`
`(D = 1.1 (rich)
`
`RON = 100
`
`
`
`
`
`1.4
`
`1.5
`
`
`9 = 0.9 tweak)
`HON = 91
`
`2500 rpm
`
`7 : 1 comprmion ratio
`
` 1.3
`RON, Research Octane Number
`
`
`lb, Equivalent ratio
`
` 1.2
`20
`4O
`60
`80
`100
`120 140
`Tamnnrmurc of the charne air (°Cl
`
`10
`
`10
`
`

`

`Turbocharging
`
`is very rarely used.
`Inter-cooling may appear attractive, but in practice it
`Compared with a compression ignition engine, the lower pressure ratios cause a
`lower charge temperature, which would then necessitate a larger inter-cooler for a
`given temperature drop. Furthermore, the volume of the inter-cooler impairs the
`transient response, and this is more significant in spark ignition engines with their
`low inertia and rapid response. Finally, a very significant temperature drop occurs
`through fuel evaporation. a process that cannot occur in compression ignition
`engines.
`,
`Water injection has been used in pressure~charged military spark ignition engines
`as a means of cooling the charge, so as to increase the charge density and inhibit the
`onset of knock. Saab have demonstrated the use of water injection in their 2.3 litre
`turbocharged spark ignition engine. Water injection (at up to 0.5 litre/min) is used at
`full throttle conditions above 3000 rpm, permitting an increase in power from
`120 kW at 4200 rpm to 175 kW at 5600 rpm for stoichiometric operation.
`The fuel/air mixture can be prepared by either carburation or fuel injection,
`either before or after the turbocharger. Fuel injection systems are simplest since
`they deduce air mass flow rate and will be designed to be insensitive to pressure
`variations. In engines with carburettors it may appear more attractive to keep the
`carburettor and inlet manifold from the naturally aspirated engine. However, the
`carburettor then has to deal with a flow of varying pressure. The carburettor can be
`rematched by changing the jets, and the float Chamber can be pressurised.
`Unfortunately, it is difficult to obtain the required mixture over the full range of
`pressures and flow rates. In general it is better to place the carburettor before the
`compressor for a variety of reasons. The main complication is that the compressor
`rotor seal needs improvement to prevent dilution of the fuel/air mixture at part
`load and idling conditions. The most effective solution is to replace the piston ring
`type seals with a carbon ring lightly loaded against a thrust face. A disadvantage of
`placing the carburettor before the compressor is that the volume of air and fuel
`between the carburettor and engine is increased. This can cause fuel hold-up when
`the throttle is opened, and a rich mixture on over-run when the throttle is closed,
`as discussed in chapter 4, section 4.6.1.
`The advantage of placing the carburettor or single point fuel injection before the
`compressor are:
`
`(i)
`
`the carburettor operates at ambient pressure
`
`(ii) there is reduced charge temperature
`
`(iii) compressor operation is further from the surge limit
`
`(iv) there is a more homogeneous mixture at entry to the cylinders.
`
`If the carburettor operates at ambient pressure then the fuel pump can be standard
`and the carburettor can be re-jetted or changed to allow for the increased
`volumetric flow rate.
`
`The charge temperature will be lower if the carburettor is placed before the
`compressor. Assuming constant specific heat capacities, and a constant enthalpy of
`evaporation for the fuel, then the temperature drop across the carburettor (ATmb)
`will be the same regardless of the carburettor position. The temperature rise across
`the compressor is given by equation (9.6)
`
`T2 = Tl[1+
`
`(Pz/PIYPWV — 1]
`
`77c
`
`11
`
`11
`
`

`

`..
`
`396
`
`
`I Introduction to internal combustion engines
`
`Lines oi mnstunl sued“: luel emanation (gm/kw hl
`
`Turbocharged
`— - — Naturally aspiroted
`
`Figure 9.22
`Comparative specific fuel
`consumption of a
`turbocharged and
`naturally aspirated engine
`sealed for the same
`maximum torque
`(with acknowledgement to
`Watson and Janota, 1982).
`
`(Yul S
`
`0'0
`
`
`
`Relativetorque
`
`20
`
`30
`
`L0
`
`70
`60
`50
`Relative speed ml
`
`80
`
`90
`
`too
`
`The term in square brackets is greater than unity, so that ATmb will be magnified if the
`carburettor is placed before the compressor. In addition, the ratio of the specific heat
`capacities (y) will be reduced by the presence of the fuel, so causing a further lowering
`of the charge temperature. This is illustrated by example 9.2, which also shows that
`the compressor work will be slightly reduced. The reduced charge temperature is very
`important since it allows a wider knock-free operation — see figure 9.20.
`In spark ignition engines the compressor operates over a wider range of flows,
`and ensuring that the operation is always away from the surge line can be a greater
`problem than in compression ignition engines. If the carburettor, and thus the
`throttle,
`is placed before the compressor the surge margin is increased at part
`throttle. Consider a given compressor pressure ratio and mass flow rate and refer
`back to figure 9.17. The throttle does not change the temperature at inlet to the
`compressor (T1), but it reduces the pressure (p1) and will thus move the operating
`point to the right of the operating point when the throttle is placed after the
`compressor and p1 is not reduced.
`By the time a fuel/air mixture passes through the compressor it will be more
`homogeneous than at entry to the compressor. Furthermore, the flow from the
`compressor would not be immediately suitable for flow through a carburettor.
`Performance figures vary, but typically a mixture boost pressure of 1.5 bar
`would raise the maximum torque by 30 per cent and maximum power by up to
`60 per cent. Figure 9.22 shows the comparative specific fuel consumption of a
`turbocharged and naturally aspirated spark ignition engine. The turbocharged
`engine has
`improved fuel consumption at
`low outputs. but an inferior
`consumption at higher outputs. The effect on vehicle consumption would depend
`on the particular driving pattern.
`
`RS
` 9.5
`
`Practical considerations and systems
`
`9.5.1 Transient response
`
`The transient response of turbocharged engines is discussed in detail by Watson
`and Janota (1982:, The problems are most severe with spark ignition engines
`
`12
`
`12
`
`

`

`Turbocharging ' 397
`
`Figure 9.23
`(a) Compressor
`pressure-relief valve control
`system. (b) Boost
`pressure-sensitive 'waste
`control system (with
`acknowledgement to
`Watson and lanota, 1982).
`
`because of their wide speed range and low inertia; the problems are also significant
`with the more highly turbocharged compression ignition engines. The poor
`performance under changing speed or load conditions derives from the nature of
`the energy transfer between the engine and the turbocharger. When the engine
`accelerates or the load increases, only part of the energy available at the turbine
`appears as compressor work, the balance is used in accelerating the turbocharger
`rotor. Additional lags are provided by the volumes in the inlet and exhaust systems
`between the engine and turbocharger; these volumes should be minimised for
`good transient response. Furthermore, the inlet volume should be minimised in
`spark ignition engines to limit the effect of fuel hold-up on the fuel-wetted
`surfaces. Turbocharger lag cannot be eliminated without some additional energy
`output, but the effect can be minimised. One approach is to under-size the
`turbocharger, since the rotor inertia increases with (length)? while the flow area
`increases with (length)? Then to prevent undue back—pressure in the exhaust, an
`exhaust by-pass valve can be fitted. An alternative approach is to replace a single
`turbocharger by two smaller units.
`The same matching procedure is used for spark ignition engines and
`compression ignition engines. However, the wider speed and flow range of the
`spark ignition engine necessitate greater compromises in the matching of
`turbomachinery to a reciprocating engine. If the turbocharger is matched for the
`maximum flow then the performance at low flows will be very poor, and the large
`turbocharger size will give a poor transient response. When a smaller turbocharger
`is fitted, the efficiency at low flow rates will be greater and the boost pressure will
`be higher throughout the range; the lower inertia will also reduce turbocharger
`lag. However, at higher flow rates the boost pressure would become excessive
`unless modified; two approaches are shown in figure 9.23.
`The compressor pressure can be directly controlled by a relief valve, to keep the
`boost pressure below the knock-limited value. The flow from the relief valve does
`not represent a complete loss of work since the turbine work derives from energy
`that would otherwise be dissipated during the exhaust blow-down. The blow-off
`flow can be used to cool the turbine and exhaust systems. If the carburettor is
`placed before the compressor,
`the blow-off flow has to be returned to the
`compressor inlet, which results in yet higher charge temperatures.
`The exhaust waste-gate system (figure 9.23b) is more attractive since it also
`permits a smaller turbine to be used, because it no longer has to be sized for the
`maximum flow. Turbocharger lag is reduced by the low inertia, and the control
`system ensures that the waste-gate closes during acceleration. The main difficulty
`is in designing a cheap reliable system that will operate at the high temperatures.
`
`Relief valve
`.—
`
`Turbocharger
`—.
`
`Engine
`
`(3)
`
`Turbocharger
`._..
`
`waste gore
`
`
`
`lb)
`
`13
`
`13
`
`

`

`
`
`5 Introduction to internal combustion engines
`
`Variable-area turbines, compressor restrictors and turbine outlet restrictors can
`also be used to control
`the boost pressure;
`restrictors of any form are an
`unsatisfactory solution.
`The transient response of an engine can be modelled by an extension to the type
`of simulation described in chapter 10 (Charlton et (11., 1991). This requires a
`knowledge of the engine, load and turbocharger inertias, and the governor and
`fuel injection pump (or engine management system) dynamic response. At the end
`of each cycle a torque balance needs to be made on each rotating component, so
`that any torque surplus (or deficit) can be used to accelerate (or decelerate) the
`relevant shaft. A change in a parameter, such as the boost pressure, can then be
`used to determine the new fuelling level and injection timing. Transient
`simulations are a useful method of identifying the strategies that might lead to
`better load acceptance on turbocharged engines.
`
`9.5.2 Variable—geometry turbochargers, superchargers and two-
`stage turbocharging
`
`Variable-geometry turbochargers
`
`Variable-geometry turbines have already been discussed in section 9.2.1, and
`illustrated by figures 9.8 and 9.9. Figure 9.24 illustrates the benefits to the low
`speed torque of using the Holset moving sidewall variable-geometry turbine of
`figure 9.9 in a truck engine application. The maximum torque engine speed range
`is extended by 40 per cent, and there is a 43 per cent improvement in the torque at
`1000 rpm.
`
`1200 -—
`
`1100 4
`
`1000 —
`
`900 —
`
`800
`
`l
`
`
`
`EEw
`
`:3
`F."
`e
`
`700 “
`
`600 4——
`500
`
`
`i
`l
`1500
`2000
`
`l
`2500
`
`l
`1000
`
`Figure 9.24
`The increase in low engine speed torque, through the use of a moving sidewall variable—geometry radial flow turbine (courtesy
`of Holset).
`
`Engine Speed (RPM)
`
`14
`
`14
`
`

`

`Turbocharging
`
`Performance comparison of the waste—gate and variable-geometry turbo-
`Table 9.1
`charger Fiat JTD direct injection diesel engines
`
`
`
`The nozzle position controls the flow area, and thus the turbine back-pressure,
`and the work output. This in turn determines the compressor boost pressure and
`the engine torque that is possible with an appropriate fuelling level. It is thus
`possible to control the turbocharger speed and boost pressure without using a
`waSte—gate. The nozzle position is varied by a pneumatic actuator, controlled by
`the engine management system. in response to:
`
`\IO‘U'I#WN—l
`
`engine speed,
`throttle demand,
`
`inlet manifold temperature and pressure,
`
`exhaust manifold pressure (optional),
`ambient pressure,
`
`turbocharger speed, and
`
`fuelling.
`
`The variable-geometry turbine area can also be used to increase the turbine back-
`pressure, thereby increasing the engine braking.
`Variable-geometry turbochargers are also used on smaller automotive diesel
`engines, and in the case of
`the Fiat JTD engines,
`the variable-geometry
`turbocharger is used to maintain a high bmep at high engine speed (Piccone and
`Rinolfi, 1998). The JTD engines are direct injection, with a bore of 82.0 mm, a
`stroke of 90.4 m, an air-to-air inter-cooler,
`two valves per cylinder. an
`electronically controlled common rail
`fuel
`injection system, EGR, and a
`compression ratio of 18.5 2 1. Table 9.1 compares the performance of two JTD
`engines.
`Table 9.1 shows how the variable-geometry turbocharger has led to a higher
`specific output by a combination of a higher bmep occurring at a higher speed.
`
`Supercharging
`
`One way of eliminating turbo-lag is to use a supercharger (a mechanically driven
`compressor), and this is the approach adopted by Jaguar for a spark ignition engine
`(Joyce, 1994). The Jaguar engine has a swept volume of 4 litres, so in normal use it
`will be operating under comparatively low load and speed conditions, the very
`conditions from which turbo-lag would be most significant. Obviously a super-
`
`15
`
`15
`
`

`

`00g Introduction to internal combustion engines
`
`charger will lead to worse fuel economy than a turbocharger. since there is no
`recovery of the exhaust gas expansion work. However, because of the limit to
`boost pressure imposed by fuel quality and knock,
`the pressure ratio will be
`comparatively low, and the fuel economy penalty will not be unacceptable. In any
`case, people who want a supercharged 4 litre engine will probably be more
`concerned with performance than economy. Jaguar adopted a Roots compressor
`with a pressure ratio of 1.5 and an air/coolant/air interscooler. The Roots
`compressor has no internal compression, and this limits the compressor efficiency,
`especially at high pressure ratios, when the effect of compression irreversibilities
`increases (Stone, 1988). The other main loss is leakage past the rotors, which
`depends solely on the pressure ratio and seal clearances, and not on the flow rate
`through the compressor. Leakage losses are most significant at low speeds since the
`flow rates are low and the running clearances are greatest (due to the com-
`paratively low temperature of the compressor): Jaguar found that inlet system
`deposits on the Roots blower rotors led to an in—service reduction in the leakage loss.
`At part—load operation the supercharger is unnecessary, so one option is to use a
`clutch to control its use. However, to avoid a loss of refinement through engaging
`and disengaging the supercharger, Jaguar adopted a permanent drive with a step-
`up speed ratio of 2.5 (the drive has to meet a requirement of 34 kW).
`(If a
`continuously variable ratio drive is available, then throttling losses can be reduced
`at part-load operation by running the supercharger more slowly, and using it as an
`expander.) 50 as to avoid unnecessary compression in the supercharger. a by-pass
`valve opens at part load (in response to manifold pressure); the main throttle is
`upstream of the supercharger.
`An inter-cooler is used since otherwise the performance gains from super-
`charging could be reduced by about 50 per cent. Without an inter—cooler, the
`supercharger outlet temperature would be limited to about 80°C (compared to
`IZOCC with the inter-cooler), and after the inter-cooler the temperature might be
`about 50°C. The inter—cooler thus increases the output of the engine in two ways.
`Firstly, the higher pressure ratio and the cooler air temperature both increase the
`air density, and secondly, the lower temperature allows a higher boost pressure or
`compression ratio for knock-free operation with a given quality fuel.
`Table 9.2 compares the naturally aspirated and supercharged Jaguar AJ6
`engines. Both have an aluminium cylinder head with 4 valves per cylinder, a bore
`of 91 mm and a stroke of 102 mm.
`
`The supercharged engine uses the same camshafts (242°ca valve open period) as
`the naturally aspirated engine, but with zero overlap at tdc, so as to avoid shon-
`circuiting loss of the mixture at high-load conditions, and to give good idle stability.
`The supercharged engine has a higher power and torque output than the
`naturally aspirated 6 litre V12 engine, and table 9.3 compares the brake specific
`
`
`
`Table 9.2 The naturally aspirated and supercharged jaguar AJ6 engines
`
`
`
`Naturally aspirated Supercharged
`
`8.5
`...,
`,
`.
`..'.9.S
`,
`1
`Compression ratio
`510 at 3000 rpm
`’370 at 3500 rpm
`Maximum torque (N m)
`16.0
`" '
`11.6
`'
`bmep (bar)
`‘ 240 at 5000 rpm
`166 at .5000 rpm .
`»
`Maximtim poWer (kW)
`
`'10.0 14.4
`
`bm‘ep (bar)
`
`"
`
`16
`
`16
`
`

`

`Table 9.3 Comparison of the brake specific fuel consumptions (g/kWh) for naturally
`aspirated and supercharged Jaguar engines at 2000 rpm, for a fixed bmep and a fixed torque
`
`
`Engmetype
`I
`’
`V
`Zbarbmep V
`..
`64 N m torque
`
`I uroocnarging
`
`4o I
`
`0.414% f
`
`v
`6.0 lltre‘V‘l 2.x .
`.
`.
`
`
`
`4,0litr'e-supercharged _
`»
`-
`411litrenaturally‘aspirated‘“
`
`0.54