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Thèse de doctorat/ PhD Thesis Citation APA:

Douxchamps, P.-A. (2010). Diesel thermal management optimization for effective efficiency improvement (Unpublished doctoral dissertation). Université libre de Bruxelles, Faculté des sciences appliquées – Mécanique, Bruxelles.

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D 03679

Diesel thermal management optimization for

effective efiiciency improvement

Pierre-Alexis Douxchamps

Aero-Thermo-Mechanics Department Université Libre de Bruxelles

A thesis submitted for the degree of

PhilosophiæDoctor (PhD) in Applied Sciences

on April 2010

'-ibre de Bruxelles

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Abstract

This Work focuses on the cooling of diesel engines. Facing heavy constraints such as émissions control or fossil energy management, political leaders are forcing car manufacturers to drastically reduce the fuel consumption of passenger vehicles. For instance, in Europe, this fuel consumption has to reach 120 by 2012, namely 25 % réduction from today’s level.

Such objectives can only be reached with an optimization of ail engines components from injection strategies to power steering. A classical energy balance of an internai combustion engine shows four main losses: enthalpy losses at the exhaust, heat transfer to the cylinder walls, friction losses and external devices driving. An optimized cooling will improve three of them: the heat transfer losses by increasing the cylinder walls température, the friction losses by reducing the oil viscosity and the coolant pump power consumption.

A model is first built to simulate the engine thermal behavior from the combustion itself to the températures of thedifferent engine components. It is composed by two models with different time scales. First, a thermodynamic model computes the in- cylinder pressure and température as well as the heat flows for each crank angle.

These heat flows are the main input parameters for the second model: the nodal one. This last model computes ail the engine components températures according to the nodal model theory. The cylinder walls température is then given back to the thermodynamic model to compute the heat flows.

The models are then validated through test bench measurements giving excellent résulta for both Mean Effective Pressure and fluids (coolant and oil) températures.

The used engine is a 1.91 displacement turbochaxged piston engine equipped with an in-cylinder pressure sensor for the thermodynamic model validation and ther­

mocouples for the nodal model validation.

The model is then used to optimize the coolant mass flow rate as a fonction of the engine température level. Simulations hâve been done for both stationary conditions with efficiency improvement up to 7% for spécifie points (low load, high engine speed) and transient ones with a heating time improvement of about 2000 s.

This gains are then validated on the test bench showing again good agreement.

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Résumé

Ce travail concerne le refroidissement des moteurs diesel. Face aux contraintes actuelles, à la fois sociétales et environnementales, le monde politique doit prendre des mesures afin de limiter les émissions polluantes (principalement le CO2) ainsi que la consommation énergétique issue des matières fossiles. Ces mesures concer­

nent notamment les véhicules automobiles. La communauté européenne impose ainsi aux constructeurs automobiles de réduire les émissions de CO2 de 25% pour 2012.

L’objectif est ambitieux et nécessite des efforts à tous les niveaux: de la combus­

tion elle-même à l’optimisation de chaque composant afin d’assurer une utilisation rationelle de l’énergie.

Les pertes énergétiques d’un moteur à combustion interne peuvent être classées en quatre catégories: pertes par chaleur sensible, pertes à l’échappement, pertes par frottement et puissances consommées dans l’entrainement des organes nécessaires au bon fonctionnement du moteur.

La gestion thermique du moteur est un des paramètres essentiels influençant son rendement. En effet, toute chaleur excédentaire doit être évacuée vers le milieu ambiant afin que le niveau des températures atteintes garantisse la bonne tenue mécanique des composants. Ce refroidissement nécessaire conduit à accentuer cer­

taines pertes du moteur et à détériorer son rendement.

D’une part, le refroidissement, assuré par la circulation d’un liquide au sein même du moteur, conduit à la diminution de la température des chemises et donc à l’augmentation des pertes par chaleur sensible. D’autre part, un état thermique plus faible conduit à une température d’huile basse et à une viscosité d’huile élevée engendréint des pertes par frottement importantes. Enfin, la circulation du liquide de refroidissement doit être assurée par une pompe consommant une partie de la puissance indiquée.

Optimiser le refroidissement des moteurs à combustion interne revient donc à aug­

menter la température moyenne du moteur tout en gardant une marge par rapport aux limites imposées par la résistance mécanique.

Ce travail présente un modèle thermique précis du moteur. Ce modèle est consti­

tué de deux parties principales ayant chacune des échelles de temps fort différentes.

Tout d’abord, un modèle thermodynamique permet de calculer la température et

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la pression des gaz à l’intérieur du cylindre ainsi que les flux de chaleur au travers des parois pour chaque angle de vilebrequin. Ces flux de chaleur font partie des paramètres d’entrée de la seconde partie du modèle: un modèle nodal de l’ensemble du moteur ainsi que de son circuit de refroidissement et de lubrification. Ce sous- modèle permet notamment de calculer les températures du liquide de refroidisse­

ment, de l’huile et des parties métalliques, en particulier les températures des parois qui sont renvoyées comme paramètres d’entrée au modèle thermodynamique afin de mesurer l’impact de l’état thermique du moteur sur son rendement.

Les modèles sont validés sur un moteur diesel de 1.9 litres de cylindrée équipé d’un turbocompresseur. Ce moteur est équipé d’un capteur de pression intra-cylindre permettant de valider le modèle thermodynamique et de thermocouples permettant de valider les résultats obtenus en terme de température de liquide de refroidisse­

ment et d’huile. Les résultats obtenus donnent entièrement satisfaction avec des erreurs maximum de quelques pour-cents pour la Pression Moyenne Effective et les températures des fluides (liquide de refroidissement et huile).

Le modèle est ensuite utilisé pour optimiser le débit de refroidissement en fonction de paramètres de fonctionnement du moteur tel que vitesse de rotation, état ther­

mique et charge. Les résultats pour les régimes stationnaires montrent des gains pouvant aller jusqu’à 7%. Pour les régimes transitoires, le gain en terme de temps de chauffe atteint 2000 s pour une démarrage à froid.

Ces gains sont également validés au travers d’une campagne d’essai. Cette cam­

pagne d’essai comprend des essais sur des points de fonctionnement stationnaires ainsi que sur des cycles de fonctionnement standardisés qu’ils soient urbains ou extra-urbains. Les résultats obtenus montrent une bonne corrélation avec les ré­

sultats du modèle.

L’outil construit remplit pleinement les objectifs fixés, à savoir un modèle ther­

mique global d’un moteur diesel permettant de prédire l’effet du refroidissement sur son rendement.

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To Caroline and Téo whose smiles supported me during these last years

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Remerciements

Je tiens tout d’abord à remercier les personnes qui m’ont donné l’opportunité de réaliser ce projet et qui m’ont encadré durant ces cinq dernières années. Il s’agit tout d’abord du Professeur Leduc qui a cru en mon projet dès le début, du Professeur Hendrick qui a accepté de devenir mon promoteur de thèse pour les deux dernières années ainsi que des professeurs Degrez et Lambert qui ont fait partie de mon comité d’accompagnement.

Je remercie également l’Université Libre de Bruxelles pour avoir subventionné ma recherche durant ces cinq dernières années.

La subvention accordée par le Fonds Universitaire Jules Reyers, sous la présidence du Professeur Jaumotte, m’a permis de mener à bien mon travail expérimental, qu’il en soit remercié.

Cette thèse n’aurait pu voir le jour sans le soutien quotidien de mes collègues qui par leurs avis, leurs remarques, leurs conseils m’ont permis d’avancer. Cette rencontre professionnelle s’est accompagnée d’une véritable rencontre humaine. J’espère que les liens que nous avons tissés perdureront. Parmi ceux-ci, une pensée particulière va à Christophe Deleplanque, Olivier Berten, Alix Cuvelier, Christophe Riga et François Gruselle.

Pour toute la partie expérimentale, j’ai pu bénéficier du support des techniciens du laboratoire du Service d’Aéro-Thermo-Mécanique, en particulier Pascal Beine, qu’ils en soient remerciés.

Mon entourage a réussi a créer un environnement idéal pour que je puisse m’épanouir Ils ont su me donner confiance afin que je gravisse un à un les échelons pour finir cette thèse. Parmi ceux-ci, Caroline tient une place très spéciale.

Le bon doctorant sait que le chemin est long mais il connaît également son dû.

Je dois à mes parents d’avoir pris cette voie. Je les remercie de m’avoir donné l’éducation et la formation nécessaires à réaliser mes ambitions.

Une pensée va également à tout ceux que j’ai pu oublier en espérant que cela ne soit pas trop long comme lacune.

Enfin, je tiens à remercier les différents lecteurs pour l’intérêt qu’ils ont marqué à ce travail.

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Contents

List of Figures ix

List of Tables xiii

Glossary xv

1 Introduction 1

2 Engine cooling state of the art 7

2.1 History ... 7

2.2 Receiit improvement in engine cooling architecture... 10

2.2.1 Valve control... 10

2.2.2 Electrical coolant puinp & Fan... 11

2.2.3 Others... 11

2.3 Modeling history... 13

2.3.1 Lumped capacity model... 15

2.3.2 Heat losses ... 15

2.3.3 Combustion model... 16

2.3.4 Model coupling... 16

2.4 Current cooling strategies... 17

2.5 Cooling control technologies... 18

2.6 Actual results of controlled cooling... 21

2.6.1 Efficiency increase... 21

2.6.2 Thermal Comfort... 21

2.6.3 Emissions... 22

2.6.4 Warm up time ... 22

2.6.5 Combustion chamber température fluctuation... 22

2.7 Summary... 24

3 Test bench description 25

3.1 Description of the engine... 25

3.2 Description of the test bench ... 32

3.3 Measurement devices... 35

3.3.1 Températures... 35

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CONTENTS

3.3.2 Fuel mass flow rate... 35

3.3.3 Iii-cyliiider pressure ... 36

3.3.4 Coolant flow rate... 41

3.3.5 Manifolds pressures... 41

3.3.6 Engine rotation speed and torque... 42

3.3.7 DAQ System... 42

3.4 Load and speed control... 43

3.5 Coolant pump control... 47

3.6 Measurement error... 50

3.6.1 Fuel mass flow rate... 50

3.6.2 Température... 50

3.6.3 In-cylinder Pressure... 52

4 Thermodynamic cycle model 53

4.1 Model types... 53

4.1.1 One zone model... 54

4.1.2 Multi-zones models... 55

4.1.3 Model choice... 58

4.2

Thermodynamics of a closed System

... 59

4.2.1 First principle... 59

4.2.2 Combustion basics... 60

4.3 External Energy Inputs and Outputs... 62

4.3.1 External work - Volume variation... 62

4.3.2 Heat losses modeling... 62

4.3.2.1 Convection losses... 62

4.3.2.2 Radiation losses... 65

4.3.3 Intake and exhaust strokes... 66

4.3.3.1 Compressible flow through a valve... 66

4.3.3.2 Lift and valve timing ... 68

4.3.3.3 Conditions in the intake and exhaust manifolds... 69

4.3.3.4 Model results... 70

4.4 Combustion model... 72

4.4.1 Jet pénétration and cône angle... 72

IV

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CONTENTS

4.4.2 Droplet Vaporizatioii... 75

4.4.2.1 Vaporization constant... 77

4.4.2.2 Initial droplet diameter... 78

4.4.3 Combustion delay and Chemical kinetics... 79

4.4.4 Ignition delay... 80

4.4.5 Combustion kinetics... 81

4.4.6 Model Results... 83

4.5 Model implémentation... 84

4.5.1 Numericcil methods... 84

4.5.2 Burnt gas fraction... 86

4.6 Summary... 87

5 Global thermal model 89

5.1 Nodal models... 91

5.2 Hydraulic modeling... 93

5.2.1 Pressure drops... 93

5.2.2 Numerical method... 94

5.3 Oil thermal model ... 95

5.3.1 Oil circuit description... 95

5.3.2 Oil properties... 95

5.3.3 Hydraulic circuit... 95

5.3.4 Transfer coefficients ... 97

5.3.5 Nodal subdivision ... 102

5.3.6 Implémentation... 104

5.3.7 Results ... 106

5.4 Coolant and engine components thermal model... 112

5.4.1 Coolant circuit description... 112

5.4.2 Coolant & métal properties... 114

5.4.3 Hydraulic circuit...115

5.4.4 Transfer coefficients ... 118

5.4.5 Nodal subdivision ... 120

5.4.6 Dimension of the different parts... 135

5.4.7 Implémentation... 135

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CONTENTS

5.4.8 Results ... 137

5.5 Coupling between thermodynamic and nodal models...144

5.6 Summary...146

6 Model validation 147

6.1 Thermodynamic model... 147

6.1.1 Experimental results... 147

6.1.2 Simulated results...150

6.2 Global model - steady State conditions... 153

6.2.1 Initial results... 153

6.2.2 Energy balance for the codant circuit...156

6.2.3 Heat fluxes from the combustion... 157

6.2.4 Radiator model...159

6.2.5 Final results...162

6.3 Global model - Transients conditions... 166

6.4 Sensitivity study...172

6.4.1 Combustion model... 172

6.4.2 Nodal model ... 175

7 Model Exploitation 177

7.1 Expected gains... 178

7.1.1 Water pump mechanical power...178

7.1.2 Indicated efRciency increase... 178

7.1.3 Friction losses... 179

7.2 Simulation results...180

7.2.1 Strategy... 180

7.2.2 Implémentation...180

7.2.3 Results for steady state conditions... 184

7.2.4 Results for transients... 188

7.2.5 Fuel economy as a function of engine speed, engine load and cycle length... 190

7.3 Experimental results... 191

7.3.1 Steady state... 191

7.3.2 European Driving Cycle...195

VI

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CONTENTS

7.3.3 Extra-Urban Driving Cycle... 198

7.3.4 Fuel economy as a function of engine speed, engine load and cycle lengtli... 199

7.4 Summary... 201

8 Conclusions &: Perspectives 203

8.1 Main contribution and conclusions... 203

8.2 Perspectives... 206

A Geoinetrical relations 207 B Droplet Vaporization 209

B.l Mass transfer... 209

B. 2 Energy transfer... 211

C Thermodynamic and physical properties of the used species 213

C. l Thermodynamic properties ...213

C.2 Air properties... 215

C. 3 Fuel properties... 216

D Thermo-physical properties for oil, coolant and mechanical compo- nents 217

D. l Oil properties...217

D.1.1 Thermal conductivity ...217

D.l.2 Beat capacity...217

D.l.3 Kinematic viscosity... 218

D.1.4 Density...219

D.2 Coolant properties... 221

D.2.1 Density...221

D.2.2 Kinematic viscosity... 221

D.2.3 Beat capacity... 223

D.2.4 Thermal conductivity ...223

D.3 Métal properties ... 227

E Measurements for validation process 229

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CONTENTS

F Measurements for optimization process 233

Référencés 237

viii

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List of Figures

2.1 Mercedes 35 hp ... 9

2.2 Split cooliiig architecture (1) 13

2.3 Valve controlled cooliiig architecture (2)... 18

2.4 Fuel economy with a valve controlled cooling (2) 19

3.1 Fuel circuit scheme ... 27

3.2 Air and gas circuit scheme ... 29

3.3 Water/oil beat exchanger ... 30

3.4 Coolant circuit scheme ... 31

3.5 Foucault currents brake... 33

3.6 Test Ben ch Scheme ... 34

3.7 LabView interface for température monitoring ... 35

3.8 Fuel mass flow rate measuring device ... 36

3.9 Thermodynamic loss angle définition (3)... 39

3.10 Polytropic coefficient vs. crank angle for different indications (4) . . . . 40

3.11 Polytropic coefficient with respect to the angle gap (4)... 41

3.12 Driving cycle example (5)... 44

3.13 Classical step response ... 45

3.14 LabView interface for engine control... 47

3.15 Timing belt scheme (6)... 48

3.16 Pump speed vs. PWM ... 49

3.17 coolant flow rate vs. PWM ... 50

4.1 Burnt fuel mass fraction as a fonction of the crank angle (7) 54

4.2 Jet geometry - two-zones model (7) 56

4.3 Jet geometry - multi-zones model (7) 56

4.4 Convection coefficient variation as a fonction of the crank angle (8) . . 64

4.5 Beat flux and emissivity for three different crank angles (9) ... 65

4.6 Valve pressure drop coefficient as a fonction of the valve opening (9) . . 68

4.7 Valve lift law (10) 69

4.8 Mass flow rate during the intake stroke as a fonction of the crank angle degree for different engine rotation speeds... 71

4.9 Primary atomization of a water jet (11) ... 73

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LIST OF FIGURES

4.10 Jet pénétration as a function of time (11) ... 74

4.11 Spalding parameters as a function of the droplet température... 76

4.12 Droplet square diameter as a function of time (12) ... 78

4.13 Droplet diameter distribution (9)... 79

4.14 Ignition delay as a function of the cetane number (9) 81

4.15 Hardenberg & Hase équation validation (9) ... 82

4.16 Combustion kinetic (7) 82

4.17 Calculated température during the combustion phase for an engine rota­ tion speed of 3000 rpm and an air-to-fuel ratio of 24.2... 84

4.18 Burnt fuel mass fraction as a function of the crank angle degree for an engine rotation speed of 3000 rpm and a fuel-to-air ratio of 0.6 ... 85

5.1 Volume subdivision for nodal models... 91

5.2 Heat transfer modes... 91

5.3 Oil circuit scheme used for the global model ... 96

5.4 Simplified scheme of the oil hydraulic circuit ... 96

5.5 Heat transfer coefficient between piston and oil jets as a function of the engine rotation speed ... 100

5.6 Oil circuit nodal subdivision ... 103

5.7 Simulink visual interface for oil model... 105

5.8 Oil circuit Simulink block... 106

5.9 Oil circuit nodal model Simulink block ... 107

5.10 Engine oil température for the different nodes ... 109

5.11 Crankcase oil température as a function of the coolant mass flow .... 110

5.12 Evolution of friction losses during warm-up ... 110

5.13 Standard cooling circuit description... 113

5.14 A "Trombone circulation" - B "Diagonal circulation" ... 114

5.15 Coolant circuit scheme... 115

5.16 Coolant hydraulic circuit results...117

5.17 Example of cylinder nodal décomposition (13)...121

5.18 Nodal décomposition for the "Diagonal circulation" ... 122

5.19 Nodal décomposition for the "Pin circulation" ...124

5.20 Engine block model ()... 129

X

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LIST OF FIGURES

5.21 Simulink node configuration ... 137

5.22 Simulink cycle model configuration... 138

5.23 Coolant température history at the engine iiilet and outlet ...139

5.24 Node température history... 142

5.25 Node mass influence on heating time... 143

6.1 Comparison between simulated and measured in-cylinder pressure in a p-V diagram (4) 151 6.2 Comparison between simulated and measured in-cylinder pressure in a p-0 diagram (4) 151 6.3 Setting points for global model validation (14) 153 6.4 Outlet coolant températures with initial model (14) 155 6.5 Inlet coolant températures with initial model (14) 155 6.6 Oil températures with initial model (14)...156

6.7 Heat transfer coefficients between piston and oil for initial and modified corrélations (14) 160 6.8 Modeling of the power evacuated by the radiator (14) 163 6.9 Final results for inlet température (14) 164 6.10 Final results for outlet température (14)...165

6.11 Final results for oil température (14)... 165

6.12 Transients results for step 1 (14) ... 168

6.13 Transients results for step 2 (14) ... 169

6.14 Transients results for step 3 (14) ... 170

6.15 Transients results for step 4 (14) ... 171

7.1 Comparison between integrator controller and integrator -|- proportional controller (14) 182 7.2 Coolant control

System

(14)... 183

7.3 Comparison between coolant flow with and without control (14) .... 185

7.4 Températures for different nodes at 40 kW and 3130 rpni with coolant flow control (14) 185 7.5 Predicted indicated efficiency increase (14) 186 7.6 Predicted effective efficiency increase (14)...187 7.7 Oil température after cold start for different initial coolant flows (14) . 188

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LIST OF FIGURES

7.8 Effective power beiiefit with coolant flow control after a cold start (14) 189 7.9 Coinparison between classical coolant flow rate and optimized coolant

flow rate ...191 7.10 Fuel mass flow rate improvement at 30 kW of effective power ... 192 7.11 Efficiency improvement vs. température increase... 194 7.12 Coinparison between optimized and classical coolant flow rate for an

European Driving Cycle...195 7.13 Coolant température évolution for four consecutive European Driving

Cycles with and without optimized coolant flow rate... 196 7.14 Comparison between optimized and classical coolant flow rate for an

Extra-urban Driving Cycle ... 199 7.15 Coolant température évolution for three consecutive Extra-urban Driving

Cycles with and without optimized coolant flow rate...200 A.l Combustion chamber volume (9) ...207 D.l Engine oil beat capacity as a function of the température ...218 D.2 Engine oil kinematic viscosity as a function of température for 15W40 oil 219 D.3 Density of water/glycol mixture as a function of the température .... 222 D.4 Kinematic viscosity of water/glycol mixture as a function of the tempér­

ature ... 224 D.5 Heat capacity of water/glycol mixture as a function of the température . 225 D.6 Thermal conductivity of water/glycol mixture as a function of the tem­

pérature ...226

xii

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List of Tables

3.1 Engine main characteristic... 26

3.3 Gear ratio values... 44

3.4 Laplace transforin parameters for a load step response... 45

3.5 Load controller parameters ... 46

3.6 Timing belt description ... 48

4.1 Model comparison ... 58

4.2 Coefficients for the Woschni équation (9)... 64

4.3 Intake and exhaust valve characteristics... 69

4.4 Variable engine rotation speed results ... 70

5.1 Equivalent conductance for the different transfer modes ... 92

5.2 NTU beat exchange parameters for oil/coolant exchanger ... 97

5.3 Mean friction power distribution at 4000 rpm (15) 101 5.4 Node description...102

5.5 Cylinder head and engine block oil ducts characteristics... 104

5.6 Simulation input parameters... 108

5.7 Steady state oil températures for an engine rotation speed of 2000 rpm . 108 5.8 Engine - pump pressure coefficients... 116

5.9 Radiator parameters... 116

5.10 Typical coefficients for convective beat transfer between cylinder walls and coolant flow ...119

5.11 Node description - "Diagonal circulation" ... 123

5.12 Node description - "Pin circulation" ... 125

5.13 Node description - "Pin circulation" ... 126

5.14 Mass flow rate distribution in the cylinder head for a "diagonal circulation" 133 5.15 Mass of the different mechanical components ... 135

5.16 Coolant volume répartition ... 136

5.17

Coolant circuit power balance at a given steady State condition

...140

5.18 Coefficients (ai) for this engine geometry...145

6.1 Experimental results for thermodynamic model validation (4)... 149

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LIST OF TABLES

6.2 Comparison between siniulated and measured values for thermodynaiiiic

iiiodel validation...152

6.9 Cylinder wall & coolant températures for different in-cylinder beat trans­ fer coefficients ...173

6.10 Indicated powers & optimized coolant flow rates for different in-cylinder beat transfer coefficients...174

6.11 Indicated powers for different discbarge coefficients...174

7.2 Fuel economy expectation... 190

7.5 Fuel economy expectation... 201

C.l Coefficients giving tbe entbalpy as a température function for tbe used species...214

C.2 Air properties... 215

C. 3 Fuel pbysical properties...216

D. l Coefficients for oil beat capacity corrélation...217

D.2 Coefficients for oil kinematic viscosity corrélation... 218

D.3 Coefficients for oil density corrélation ...219

D.4 Coefficients to compute tbe density as a function of tbe mixture tempér­ ature for different glycol concentrations... 222

D.5 Coefficients to compute tbe viscosity as a function of mixture température for different glycol concentrations... 224

D.6 Coefficients to compute tbe beat capacity as a function of tbe mixture température for different glycol concentrations... 225

D.7 Coefficients to compute tbe thermal conductivity as a function of tbe mixture température for different glycol concentrations... 226

D. 8 Métal Properties... 227

E. l smallcaption... 231

F. l smallcaption... 235

XIV

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Glossary

m Mass flow (^)

A Regular pressure drop coefficient for smooth ducts [-]

Cf Heat capacity mass flow [^]

Hi Engine heat loss [J]

Tip Polytropic coefficient [-]

Pmf Mean friction pressure [Pa]

Rm Clankpin radius [m]

a Thermal diffusivity [^]

P Dilatation coefficient [1/K]

Ai/r Pressure drop [m]

Atinj Injection time [s]

6 Droplet diameter [m]

Ai Node thickness [m]

e Emittivity]-]

T) efficiency [-]

7 Ratio between heat capacity at con­

stant pressure and spécifie volume [-]

H Dynamic viscosity [^]

1/ Kinematic viscosity [^j ui pulsation |^]

$ Heat flow [W]

(j> Jet angle [radian]

’ï' Radiative heat transfer coefficient [^1

P Density ]^[

a Stefan-Boltzman constant [^^4 ] T Ignition delay [ms]

6 Crtmk angle [radian]

4 Volume fr2iction [-]

Çr Pressure drop coefficient [-[

a Crankshaft length [m]

AFR Air to Fuel Ratio [-[

B Bore [m]

Bt Thermal Spalding péirameter [-[

By Species Spalding parameter [-]

Cd Discharge coefficient [-[

Cp Heating capacity at constant pres­

sure

c„ Heating capacity at constant spécifie volume [:^ ]

Ci Aerodynamic coefficient [-[

D Diameter [m]

D* Species diffusivity coefficient]^]

E Heat exchanger efficiency ]-]

e Thickness ]m]

E„ Activation energy ];;j^]

F Heat flux ]^[

Fr Force ]N]

fr rolling résistance coefficient ]-]

g Gravity [9.81 ^ ]

Gy Equivalent conductance between node i and j \^\

Gr Grashoff number ]-]

H Enthalpy ]J]

h Molar enthalpy ]^|

hc Convection coefficient he Height ]m]

hm Mass enthalpy ]jr^]

IMEP Indicated Mean Effective Pres- sure]bar]

k

Thermal conductivity

\-^\

KS Globcd heat exchange coefficient

\^\

L Length ]m]

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GLOSSARY

1 Connecting rod length [m] r Radius [m]

I^m Wet perimeter length [m] Ra Rayleigh number [-]

L„ Vaporization Heat [^] Re Reynolds number [-]

LHV Low Heating Value RR Reaction Rate [ô^]

M Molar mass [-^-1 S Area [m^[

m Mass [kg] S Stroke [m[

Injected fuel mass [kg] T Température [K]

N Engine rotation speed [rpm] T Torque [Nm]

n Droplets number [-] t Time [s]

^cylindera Number of cylinders [-]

U Internai energy [J]

NTU Number of Transfer Units [-]

U Molar internai energy Nu Nusselt number [-]

Um Mass internai energy

[-^]

P Power [W]

V Volume [m®]

P Pressure [Pascal]

V Speed [f ] Px Partial pressure of species x [Pascal]

Vr Radial speed [^]

Pr Prandtl number [-]

w Work[J]

Q Heat [J]

X Jet pénétration [m]

qx Volumétrie flow ]^^]

Yx Molar fraction of species x [- R Perfect gas constant [8.31

XVI

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Le temps d’un monde fini commence.

Paul Valéry

Introduction

Currently the world is facing major challenges leadiiig political world to reduce drasti- cally CO2 émissions and fuel consumption.

• more and more people are living in urban areas;

• more and more people will hâve access to personal vehicles (especially in emerging countries like China or India);

CO2 émissions affect the world climate, leading to natural disasters (hurricanes , heavy rains....):

• fuel consumption will soon or later be higher than fuel production. The oil price will then become quite unstable.

The world population must then lower its fossil energy consumption and find other ways to produce long term, greener, renewable energy sources.

Regarding these worldwide trends, the different political actors hâve taken mea- sures to decrease both CO2 émissions and oil consumption. The most famous one is the Kyoto Protocol. In order to achieve the different objectives set by the Protocol, ré­

gional political decision makers hâve published laws and régulations for the automotive industry.

The objectives set in the Kyoto Protocol are 20% réduction of total CO2 émissions by 2020 compared to the émissions of the year of reference - 1990.. In the European Union, private cars are responsible for 12% of the total CO2 émissions. The environmental commission of the European Parliament (16) has then fixed émissions objectives for cars with a reference mass of less than 2610 kg. These objectives are: 120 g C02/km

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1. INTRODUCTION

by 2012 and 95 g C02/kin by 2020. The current level is 160 gr (702/km. That is to say 25% réduction by 2012 and 41% réduction by 2020. In terins of fuel consuinption 160 gr C02/kni corresponds to 6.03 1/100 km (on the base of classical diesel fuels) or , 120 g (702/km to 4.52 1/100 km and 95 g (702/km to 3.58 1/100 km. Tins huge réduction is said to be achieved by both technological improvements on the engine and other technologies like tire improvement or larger use of biofuels.

In the United States of America, the department of transportation published each year (17) the Corporate Average Fuel Econoniy (or CAFE) standards for passenger cars.

These standards hâve not changed since 1990 and are still set at a ininimum of 27.5 miles/gallon (8.54 1/100 km). These régulations hâve in fact more economical reasons (control of the energy balance) than ecological ones eisthe US hâve not signed the Kyoto Protocol.

In Japan, the government has defined objectives for gasoline engines of 21.4% con- sumption réduction by 2010 in comparison with the levels of 1995. For diesel engines, the objectives are set at 13.1% consomption réduction. The différence of objectives between gasoline and diesel cornes from the fuel consuinption différence between these two engines. It is then expected that gasoline engines and diesel ones will hâve a fuel consuinption of respectively 6.5 1/100 km and 8.3 1/100 km (18).

As it can be seen, environmental, energy and economical constraints hâve forced political decision makers to put heavy requirements on car manufacturers. The objec­

tives of 20% consomption réduction in the European Union and Japan are particularly ambitions. New technologies in every vehicle field is welcome to help to reach these objectives.

An energy balance of a classical diesel engine shows that almost 30% percent of the injected energy (calculated by the Low Heating Value LHV) is really transformed in effective torque, 30% is lost in terms of heat (mainly evacuated by the cooling circuit), 30% is lost in the exhaust and the last 10% are lost in friction and mechanical running components (injection pump, alternator, water pump,...).

An energy balance of a classical vehicle running at 100 km/h shows that 70% goes in the aerodynamic résistance and 30% in the rolling résistance (tire deformation). These values are based on a SCx of 0.64 (actualSedan car) and a tire coefficient of 6 kg/t (actuel low energy consumption tire). Tins répartition is highly speed dépendent, a doubling of the speed resulting in an aerodynamic résistance four times higher. The

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driver’s behavior or tlie use of the vehicle (urban or rural) is also determining for the vehicle consumptioii. Indeed, the inertia effect can represent up to 50% of the energy consuinption for urban driving.

The possibilities of efhciency improvement can be drawn from these two energy balances:

• the indicated torque can be increased with a liigher engine compression ratio or better injection technologies;

• beat losses can be reduced with an optimized cooling (with a cooling flow rate related to the engine thermal state);

• exhaust losses can be reduced with an extended power cycle (Miller cycle);

• friction losses can be reduced with an optimal oil viscosity or thermal management;

• the energy costs of external devices can be reduced with electric driving (for instance the energy cost of power steering can be reduced by 20 % with an electric drive compared to the hydraulic one);

• the aerodynamic résistance can be lowered thanks to an optimal vehicle design;

• the rolling résistance can be reduced with low hystérésis tires;

• the driver’s behavior through his average speed or his speed variations (inertia effect) can greatly influence the fuel consumption;

• the use of the vehicle can be improved with the implémentation of a "Stop & Start"

strategy, the use of biofuel blends, automatic gear boxes and cruise control.

Ail these improvements can be implemented with the current technological know- how.

This Work focuses on the improvement of the diesel engine thermal man­

agement. Indeed, the engine cooling impacts three losses: heat losses, friction losses (viscosity is extremely température dépendent) and mechanical losses by reducing the power consumption of the water pump.

The architecture of this study is as follows. First of ail, in the chapter two, a bibli­

ographie study will présent the history of engine cooling, the current main technologies

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1. INTRODUCTION

and the recent improvements. This work follows the Phd Thesis of Frédéric Pirotais (10) who bas presented a inodel of the sanie engine without studying the link between the engine thermal State and its efficiency.

To assess the impact of new cooling strategies, a model lias been built. This model takes into account the interaction between the combustion, the wall beat fluxes and the global thermal status of the engine. This interaction occurs through a coupling between a cycle model with a high time discretization (down to a crankshaft degree) and a nodal model of the entire engine (with a time step of about 1 second).

This model has been built with two main constraints: first a low complexity linked to a low computation time and secondly a description as physical as possible. This last constraint is set in order to allow to use the presented approach to a different engine.

As the results of the model will be compared to measurements, the model parameters are set for a spécifie diesel engine (1.9 liter displacement turbocharged) . The description of this engine as well as the test bench and its instrumentation are presented in chapter three.

The description of the cycle model, its assumptions, limitations and implémentation are presented in chapter four.

The theory of the nodal models, its application to the engine components, the coolant circuit and the oil circuit are presented in chapter five. In this chapter, the interaction between the two models through the cylinder sleeves beat fluxes and tem­

pératures is also presented as well as its implémentation in the Matlab Simulink envi­

ronment.

A mathematical model has no reason to exist without a validation. The two models are thus validated through measurements on the chosen engine. The main measured parameter for the cycle model is the in-cylinder pressure. The injection law, the volu­

métrie efficiency and the combustion law are adapted to match the measured pressure history for a whole range of engine rotation speed and effective power. Concerning the nodal models, global measurements such as coolant température, oil température and fuel consumption are performed. The model architecture and beat flux laws are then modifîed in order to match the measurements for steady State conditions; the nodes weights are modified to match the measurements for transients conditions. The measurements and the results of the validation are presented in chapter six.

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The model exploitation is presented in chapter seven. A new cooling strategy is pre- sented and implemented on the developed model. The expected benefits for the different losses (heat losses, friction and mechanical losses), are calculated. This optimization takes into account the mechanical constraints in order to prevent engine failure due to overheating. The benefits are presented in terms of fuel consumption improvement for steady state conditions and in terms of cold start heating improvement for transients conditions.

The main conclusions and perspectives are resumed in chapter eight. The future of vehicles driven by fossil energy is also discussed and personal features are presented concerning forecasts on individual mobility.

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1. INTRODUCTION

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I tbink cars today are almost the exact équivalent of tbe great Gotbic catbedrals. I mean tbe suprême création of an era, conceived witb passion by unknown artists, and consumed in image if not in usage by a wbole population wbicb appropriâtes tbem as a purely magical object.

Roland Barthes

Engine cooling state of the art

This chapter will présent the history of the combustion engine and how cooling tech­

nologies hâve evolved in parallel. It is particularly interesting to note that cooling exists since the very first development of the internai combustion engine. As this work con- sists in presenting a new kind of low complexity engine thermal model, particular focus will be reserved to the coolant circuit and beat transfer modeling. This modeling has started around the thirties with the work of Nusselt and Eichelberg but it is only at the start of the eighties with the rise of the modem computer era that the first simulation results were published.

Concerning the results obtained with the model, two sections are dedicated respec- tively for the current cooling strategies and coolant circuit design adopted by manu- facturers and for the latest developments of coolant flow control. This current research topic is of first importance since the engine efhciency has become the main requirement for the engine design.

Finally, a section is dedicated to the approach of Frédéric Pirotais, its limitation and how the présent work adds the efhciency prédiction to that category of models.

2.1 History

The development of cooling circuit started just after the development of the first internai combustion engine. The first engines in the modem history were the ones driven by water vapor. These developments date back from the XVIIth century with the works of Salomon de Caus who created a pump driven by a vapor mill. This idea was improved by Fdward Somerset with a vapor cooling System in 1663.

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2. ENGINE COOLING STATE OF THE ART

A riiilestone in the engine history is the invention of the piston by Denis Papin with a first commercial model in 1712. The principle of cooliiig was the centerpiece of the engine: vapor was admitted in a cylinder where it was cooled down by water jets to produce power. The vapor engine was improved along the XVIIIth century with notably the work of James Watt.

In 1807, two major developrnents brought the engine history one step further. The Works of the Niepce brothers led to an engine called Pyréolophore (19) based on beat transfer only. The water vapor was not the working fluid anymore as it was replaced by air. The working principle of this engine is close to the Stirling cycle. In Switzerland, Isaac de Rivaz created the first engine based on the combustion of an inflammable gas.

The principles of thermodynamic and thermal engines hâve then been theorized in the book of Sadi Carnot published in 1824: "Réflexions sur la puissance motrice du feu et sur les machines propres à développer cette puissance". Prom this moment, the experimental development of gas fired engine will progress slowly during almost 35 years until the works of the Belgian scientist Lenoir. He patented in 1859 a two-stroke engine based on the dilatation of a gas-air mixture.

From this date, the engine development grew exponentially with scientists such as Beau de Rochas, Otto or Benz.

As the engine power was growing, the need for cooling appeared and in 1867 the first water cooling System based on the thermosiphon effect was created on the piston engine of Otto and Langen (20). The development of cooling technologies based on either air or water beat transfer then continued during the following years with the Works of Gottlieb Daimler and Willem Maybach which led in 1884 (20) to:

• the création of a cylinder with blades,

• the invention of the first ventilator based on blades mounted on the crankshaft,

• the first engine with differentiate cooling: cylinder head with water and cylinder walls with air.

In 1892 with a growing effective power, the amount of beat release led to the first beat exchangers: the first water coil appeared on a Daimler engine. This invention preceded by a few years the first radiator patented in 1897 and installed in 1901 in the well known Mercedes 35hp 2.1, "the first car like the modem ones". This model

8

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2.1 History

had anotlier breakthrough feature: the water was provided by a puinp driven by the crankshaft. The cooliiig architecture has alraost not changed since then. This car achieved 35hp at 950 rpiii with a majcimum speed of 75 km/h (21).

Figure 2.1: Mercedes 35 hp

Soine new developmeiits appeared along the XXth century leadiiig to the curreiit classic cooling circuit, amoiig them the first thermostat was introduced on the Renault Reinstella in 1933 (22).

The engine cooling faced many problems before World War II. Indeed, the technology of water circulation and cooling was inherited from the vapor engines where water loss was accepted. As the water was heated, it expanded and caused leakage through the pump Seal. When sufficient water had been drained out the engine, the pump could not send water to the radiator anymore, leading to overheating and water boiling. The engine had to be water filled as soon as possible especially in critical area like mountain passes.

Many manufacturers hâve then continued to produce air-cooled engines to avoid these troubles like Volkswagen and Porsche. During World War II, the need for reliable vehicles led to research and developments on boiling water in internai combustion en­

gines. Even if the problem of boiling was solved, the air-cooled engine continued to be known cis the most reliable cooling solution (as with the Volkswagen Beetle).

The two different cooling methods (water and air) continued to be developed during

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2. ENGINE COOLING STATE OF THE ART

the following years. In 1961, the Renault 4 was equipped with the first pressurized water cooling circuit with a water-coolant mixture. The boiling point was then delayed with a higher température. In 1969, Porsche introduced an air turbine for engine cooling on its 917 niodel. This turbine provided 2400 liters of air per second at a speed of 8400 rpm with a power consumption of 17 hp (23).

In 1979, the first electric water pump was tested on a Renault 14 prototype and in 1988 the pressure of the cooling circuit reached 1.4 bar on the Peugeot 605 (23).

As the need of engines with high efhciency was growing, the area of engine cooling underwent lot of research and developments from the mid eighties.

2.2 Recent improvement in engine cooling architecture

Ail the recent improvements aim to increase the engine efficiency, to improve the passen- ger thermal comfort, to reduce the vehicle émissions, to decrease the warm-up tinie and to reduce the engine température fluctuations. This goals can be achieved by several devices which are detailed hereunder.

2.2.1 Valve control

J.R. Wagner associâtes a mathematical model of a Servo-Motor driven thermostatic valve and a variable speed electric water pump to a lumped parameter engine thermal model (24), (25), (26). This association is used in order to control the engine coolant or cylinder head and wall températures according setting points determined by a two- dimensional lookup table based on the engine speed and load.

In (27), an electric température control valve was designed. This control valve is controlled to minimize the température variations depending on load and traffic condi­

tions. The management of this valve is based on température sensors and the vehicle speed measurement.

In the same approach as the one presented in (24), prédictions are made in (28) on the effect of an electronically controlled proportional valve aiming at keeping métal température under a spécifie value. The conclusion stated that this valve should produce an higher engine efficiency and reduce thermal stress and fatigue.

Pursuing the goal of combustion performance promotion, the potential of electronic equipments to control coolant température and flow rate are studied in (29). In this

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2.2 Recent improvement in engine cooling architecture

reference, an electronic control valve is associated to an electrical water puinp in an experimental study.

As tliere is a potential for the heater core to be insufficient at low puinp speeds, Advanced Thermal Management Systems are studied in (30) in order to increase engine warm-up, cabin warm-up and heater performance. These Systems include Electrical Flow Control Valve and Dual Electric Fans.

In

(31),

a Smart thermostat valve is presented. Its design features a

DC

gear motor and rotational potentiometer to control valve position. This advanced System is intended to provide the ability to reduce the overall codant flow and to allow coolant température control.

Finally, as during the first 195 seconds of an European Drive Cycle 60% of the fuel energy is used to warm-up the engine and transmission, the PITSTOP project was initiated in 2005 (32) to reduce fuel consumption during the power train warm up period. In this framework, the wax-filled thermostat was exchanged for an electrically controlled diverter valve.

2.2.2 Electrical coolant pump Fan

In most of the approaches, the electrically controlled thermostat is associated to an electrical coolant pump (24), (27), (25), (29), (30), (26), (32) and a electrical fan (27), (30) and (26)

In (33), different innovative engine cooling Systems are compared. Among these, the THEMIS System and Coolmaster both use an electrical coolant pump to achieve coolant flow rate decrease of about 30 to 50%. The pump power consumption lies between 200 and 600 W.

2.2.3 Others

As it has been shown, electrical devices are the most popular to control the coolant flow rate. However, some other techniques are studied and presented hereafter.

Nucleate boiling The nucleate boiling show great potentid to achieve low power coolant pump (of about 50 W) (33). The working principle is based on the convective heat transfer between coolant and cylinder wall which can be drastically increased in case of coolant nucleate boiling. The drawback of such an approach is the risk of

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2. ENGINE COOLING STATE OF THE ART

Departure from Nucleate boiling. Iii such a case, the beat transfer coefficient decreases abruptly. Tlierefore the cylinder wall température can increase above its thermal limit.

Heat storage

In order to reduce the engine thermal fluctuations, the possibility of a heat load averaging System lias been studied iii (34). This System is based on a heat accumulator which contains a phase change material in thermal contact with the coolant. With such a System, the codant inventory can be decreased and permits faster warm-up time.

Variable flow

It has been shown that the coolant flow rate Ccm be controlled with a variable speed water pump. Other techniques can however be used to achieve the same goal. For instance, the variable pre-swirl has been tested in (35). This technique shares the same objective of increasing the total efficiency of the cooling System by delivering appropriate flow in response to transient thermal conditions, and by reducing the parasitic load of the coolant pumping System.

The variable pre-swirl control technique uses actuated inlet guide vanes in order to modify the pump characteristic curve (Head vs. volumétrie flow)

Heat exchangers The Ultimate cooling System presented in (33) proposes to use engine coolant exclusively to cool ail the fluids in the vehicle i.e. réfrigérant, oil, exhaust gas and fuel. The main benefits are a packaging réduction in front end and a better charger air cooling resulting in better engine performances.

Split cooling This technique proposes two cooling circuits: one for the cylinder head and one for the engine block. Indeed, the cylinder head is more thermal stressed (heat power from the exhaust gas) leading to engine block over-cooling if only one circuit is used. The expected efficiency increase lies between 4 and 6 % according to reference (36).

The split cooling first goals were: to reduce warm-up, pollutant émissions and to increase the compression ratio in spark ignition engine (1). Figure 2.2 shows a typical split cooling architecture.

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2.3 Modeling history

l.l’. Water putnp 6.6:

2.2'. Heat exchanger 7.

3.3*. Electro-«tiagnetic flow meter 8

.

8

'.

4

.

4

'.

Flow control valve for coolant 9.

5.5’. Thermocouple

Mixing box

Cyl. head gasket without water holes Flow control valve for beat exchanger DC-OY Dynamometer

Figure 2.2: Split cooling architecture (1)

2.3 Modeling history

The history of engine beat transfer modeling started in 1939 with the work of Nusselt and Eichelberg (37). These first works studied the convective heat transfer between the combustion gas and the cylinder walls. The main assumptions were: steady State and one-dimensional heat transfer. A constant mean gas température was assumed and the heat transfer was expressed as:

(t> = h {Tgas - Tyjalls) (

2

.

1

)

where the convection coefficient h was computed through semi-empirical corrélations in order to match the global measurement of heating power.

As the computation capacity increased, the first attempts to solve the heat équation (Equation 2.2) started.

V(fcAT) = (2.2)

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2. ENGINE COOLING STATE OF THE ART

But tlie high frequency of the boundary conditions and tlie high spatial discretization led to huge computing resources and tirae request and was thus not developed as a global approach but limited to local coinponents modeling.

These two approaches were reconciled (38) by using the Eiclielberg approach in the components where steady State thermal behavior is achieved and by using équation 2.2 in components afFected by the transient heat transfer. This last zone is called the pénétration depth. This method gave satisfactory results for local heat transfer through surfaces with irregularities.

Beside these spécifie studies on the heat transfer through the cylinder walls, the global cooling circuit was also described using more and more complex model. One of the most complété models is the Vehicle Engine Cooling System Simulation program developed by the Michigan Technological University ((39) and (40)) where the nodal ap­

proach was chosen. In this model, ail the devices where modeled with thermal équations such as the ones described in Chapter 5. The main différence with the work presented here is the engine model. Indeed, in these models, as the one presented in (10), the thermodynamic model gives the heat fluxes based on a mean wall température which is fixed for ail the simulations. There is no feedback between the thermodynamic and the nodal model. Fuel consumption cannot be assessed, as the indicated efficiency is not influenced by the nodal model results.

Finally, must be cited the 1-D hlack box model developed for the engine designers.

These models are based on a global energy balance. These softwares provide libraries where the different components of the cooling circuit can be chosen and their parameters designed. These models hâve a limited physical description but give reliable results and are cost effective. The current cooling circuits are developed on this base without any interaction with the fuel consumption or thermal discretization of the engine itself.

Besides these general trends, a screening of different thermal models is presented here after. This study is articulated around four main points: the lumped capacity model (also called nodal model), the heat losses (representing the main heat source to the cooling circuit), the combustion model and the coupling between the coolant and combustion model.

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2.3 Modeling history

2.3.1 Lumped capacity model

The lumped capacity models are widely used to describe engine thermal behavior. How- ever, the degree of description, the node masses or used beat transfer corrélation can be very different.

In (41), a particular focus is reserved to the combustion chamber which is split into four nodes: one for the head, one for the piston and two for the liner.

In (24), the cylinder-head alone is described by seventeen nodes. The corresponding network has been constructed to represent the varions paths upon which energy is exchanged between nodes. The main criteria chosen for tins nodalisation is the Biot number which helps to assure that the validity of the thermal block behavior (the thermal gradients in the node are negligible) is correct.

On the contrary the description can be very low, as in (25) with only 3 nodes (one for the heater (engine), one for the coolant pump and one for the cooler (radiator).

The detail level of the engine description is determined by the approach: either local (e.g. the study aims at determining the influence of local températures on the engine efficiency) or global (e.g. the study aims at determining the minimum coolant flow rate to evacuate enough beat from the engine).

The ability of these models to model transients behavior is also one of their main advantages. This ability is exploited in (28) where the métal masses are divided in two:

one separating the burning gases from the coolant flowing inside engine head and block, the other separating the coolant from the external air.

2.3.2 Heat losses

In steady State, the engine cooling can be resumed to a heat transfer between two points (the combustion chamber hot point and the outside air cold point). Between these two points, there is a global conductance which summarizes the heat transfer between hot gases and cylinder wall, between cylinder walls and coolant and finally between coolant and outside air through the radiator.

As these conductances are placed in a serial way, the most limiting conductance(s) will pilot the heat transfer.

For that reason, old corrélations are still used with good simulation results. For instance, the corrélation of Woschni is used in (41) and (28). On the other si de, the

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2. ENGINE COOLING STATE OF THE ART

discrepancy between ail the semi-experimental corrélations is high as shown in (42).

2.3.3 Combustion model

To assess the efFect of the thermal management on the engine efficiency, one of the crucial part is the combustion model.

Tins model can be as simple as an algebraic corrélation giving the beat release.

These models like the one of Taylor and Toong are still used in recent studies (24).

Tins simplified approach is also used in (32) with good results on global value like warm up time.

Single zone models which compute pressure and température history in the com­

bustion chamber can be used as in (28). However, a beat release law has still to be defined.

The use of such beat release law leads to a poor description of the impact of the engine thermal State on the combustion process. To overcome this lack, separate multi- zone combustion model can be used as in (41). This modeling strategy give good results for thermal engine state/combustion process interaction

2.3.4 Model coupling

As it has been shown, new cooling strategies could be assessed with a lumped model associated with a multi-zone combustion model using standard beat transfer corrélation.

The two models can be associated using look-up tables like in (26) or (41) where the thermal power from the combustion chamber is a fonction of the engine rotational speed and load.

Otherwise, a dynamic coupling can be achieved. This approach, chosen in (27), demonstrates ail the benefit obtained with a fully coupled approach even if the com- plexity of the model is increased

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2.4 Current cooling strategies

2.4 Current cooling strategies

The current cooling strategies consist in designing the puinp for strenuous cycles such as maximum power or low load at low speed. The puinp characteristic is then chosen to provide sufRcient coolant fiow rate in order to keep the coolant température below its boiling point. These cycles can be for instance race conditions, the engine is then used at its maximum power, sand hole ^ or mountain driving, the engine is then used at high load and low engine rotation speed. (43).

With such strategies, the engine is properly cooled for every conditions. But, for low load and high engine rotation speed, the engine is unnecessarily overcooled. Indeed, the pump is, for the current strategies, directly driven by the crankshaft leading to a coolant flow rate which is increasing with the engine rotation speed.

The engine cooling thus increases also with the engine rotation speed while the thermal losses do not follow this trend. The engine thermal state is then lower than the maximum tolerated regarding the mechanical strength.

Moreover, according (44), the main disadvantages of such a design are as follow:

• the use of centrifugal impellers that hâve a lower efhciency. These impellers are used to produce relatively high pressure which is not necessary;

• the pumps are driven by either a belt or a gear generating side loading which considerably lowers the pump bearings lifetime;

• limited placement leading to an expensive maintenance;

• the "melting wax" based thermostat; this type of control is inefficient compared to the current control devices. The System does not respond properly to environ- mental changes such as ambient température.

' The vehicle has to go out of a saind hole by riding in circles

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2. ENGINE COOLING STATE OF THE ART

2.5 Cooling control technologies

Tlie first developments on cooling control appeared in tlie second half of tlie eiglities.

Three different Controls can be distinguished:

• valve controlled cooling,

• air flow controlled cooling,

• puinp controlled cooling.

The valve controlled strategy is tlie easiest to put into practice. The two expected benefits are the indicated efficiency increase by lowering the thermal losses and the effective efficiency increase by lowering the friction losses. A study performed in 1988 (2) used this technology to raise the coolant température up to 150 °C. A high boiling point (180 °C) coolant fluid was used. Figure 2.3 shows the used architecture.

Figure 2.3: Valve controlled cooling architecture (2)

Figure 2.4 shows the obtained fuel economy results with such a device. The results are close to the results obtained and presented in the last chapters. The fuel economy grows with the coolant température. The thermal losses are smaller as the thermal gradient between the combustion gas and coolant is lower. The measured benefits are also higher at part load. Indeed, the coolant circuit is designed to provide enough cooling at the critical points as at full load. The cooling at these setting points is then close to the optimal one.

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2.5 Cooling control technologies

Figure 2.4: Fuel economy with a valve controlled cooling (2)

In the early nineties, Valeo and Renault developed the first complété cooling System control dedicated to automotive engines (45). This System was composed by an electric pump, electronic driven valves, electric fan and shutters, and blades that control the air flux through the radiator. The System manages ail the beat flows to improve efïiciency and passenger comfort. Ail the actuators are driven by a micro-controller which receives ail the sensors information. Again the main différence with a classical cooling System is the coolant température which is raised by approximatively 20 °C.

The results obtained with this cooling System are presented in Table 2.1.

Cycle Fuel consumption Mechanical Pump [L/100 km]

Fuel consumption Electrical Pump

[L/100 km]

1 18.01 22.39

2 10.31 10.19

3 8.89 9.21

4 8.69 8.87

5 8.75 8.69

6 8.44 8.2

7 8.39 8.48

8 8.95 8.38

Average 8.63 8.37

Table 2.1: Consumption results with the Valeo

System

(45)

(41)

2. ENGINE COOLING STATE OF THE ART

It caii be seen that effective benefits in fuel consumption can be achieved (around 3

%). The first cycle (cold start conditions) shows an increase in fuel consumption which is due to the System starting procedure where each device (fan, pump...) is tested at full power.

The main goals of such a technology are (46):

• control the energy deinand of the coolant pump,

• reduce the warm-up time,

• maintain the engine at high température to improve indicated and effective effi- ciency. This last point is more important at part load.

In order to reach such objectives, the following strategies can be applied (46):

1. bypass the heat exchanger (thermostat job). The main conséquences of this strat- egy are: a lower pump energy deinand (lower pressure drops) and a lower warm-up time;

2. maintain the coolant flow over a minimum value until the coolant température reaches its reference value in order to avoid hot points in the engine;

3. avoid large température différences between engine outlet and inlet which leads to large température gradients in the engine. These gradients can lead to non uniform expansion;

4. control the outlet température in case of high loads or demanding external con­

ditions (high ambient température).

The controlled parameter can be either the coolant température (24), (28) or the température of one rnetallic part (28), (29).

The coolant pump can even be stopped during the engine warm up 2is shown in (28) and (29).However this strategy must be avoided in order to prevent hot points which can damage the engine.

20

Figure

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