Diesel thermal management optimization for
eective eciency 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
This work focuses on the cooling of diesel engines. Facing heavy constraints such as emissions 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 𝑔 𝐶𝑂𝑘𝑚2 by 2012, namely 25 % reduction from today's level.
Such objectives can only be reached with an optimization of all engines components from injection strategies to power steering. A classical energy balance of an internal 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 temperature, the friction losses by reducing the oil viscosity and the coolant pump power consumption.
A model is rst built to simulate the engine thermal behavior from the combustion itself to the temperatures of thedierent engine components. It is composed by two models with dierent time scales. First, a thermodynamic model computes the in- cylinder pressure and temperature as well as the heat ows for each crank angle.
These heat ows are the main input parameters for the second model: the nodal one. This last model computes all the engine components temperatures according to the nodal model theory. The cylinder walls temperature is then given back to the thermodynamic model to compute the heat ows.
The models are then validated through test bench measurements giving excellent results for both Mean Eective Pressure and uids (coolant and oil) temperatures.
The used engine is a 1.9l displacement turbocharged 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 ow rate as a function of the engine temperature level. Simulations have been done for both stationary conditions with eciency improvement up to 7% for specic 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.
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 an de limiter les émissions polluantes (principalement le𝐶𝑂2) 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 de𝐶𝑂2 de 25% pour 2012.
L'objectif est ambitieux et nécessite des eorts à tous les niveaux: de la combus- tion elle-même à l'optimisation de chaque composant an 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 inuençant son rendement. En eet, toute chaleur excédentaire doit être évacuée vers le milieu ambiant an 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 déteriorer 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 engendrant des pertes par frottement importantes. Enn, 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 diérentes.
Tout d'abord, un modèle thermodynamique permet de calculer la température et
du moteur ainsi que de son circuit de refroidissement et de lubrication. 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 an 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 Eective et les températures des uides (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 chaue 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 xés, à savoir un modèle ther- mique global d'un moteur diesel permettant de prédire l'eet du refroidissement sur son rendement.
To Caroline and Téo whose smiles supported me during these last years
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 bénécier 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 conance an que je gravisse un à un les échelons pour nir 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.
Enn, je tiens à remercier les diérents lecteurs pour l'intérêt qu'ils ont marqué à ce travail.
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 Recent improvement in engine cooling architecture . . . 10
2.2.1 Valve control . . . 10
2.2.2 Electrical coolant pump & 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 Eciency 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 temperature uctuation . . . 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 Temperatures . . . 35
3.3.2 Fuel mass ow rate . . . 35
3.3.3 In-cylinder pressure . . . 36
3.3.4 Coolant ow 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 ow rate . . . 50
3.6.2 Temperature . . . 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 ow 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
CONTENTS
4.4.2 Droplet Vaporization . . . 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 implementation . . . 84
4.5.1 Numerical 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 coecients . . . 97
5.3.5 Nodal subdivision . . . 102
5.3.6 Implementation . . . 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 & metal properties . . . 114
5.4.3 Hydraulic circuit . . . 115
5.4.4 Transfer coecients . . . 118
5.4.5 Nodal subdivision . . . 120
5.4.6 Dimension of the dierent parts . . . 135
5.4.7 Implementation . . . 135
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 coolant circuit . . . 156
6.2.3 Heat uxes 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 eciency increase . . . 178
7.1.3 Friction losses . . . 179
7.2 Simulation results . . . 180
7.2.1 Strategy . . . 180
7.2.2 Implementation . . . 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
CONTENTS
7.3.3 Extra-Urban Driving Cycle . . . 198
7.3.4 Fuel economy as a function of engine speed, engine load and cycle length . . . 199
7.4 Summary . . . 201
8 Conclusions & Perspectives 203 8.1 Main contribution and conclusions . . . 203
8.2 Perspectives . . . 206
A Geometrical relations 207 B Droplet Vaporization 209 B.1 Mass transfer . . . 209
B.2 Energy transfer . . . 211
C Thermodynamic and physical properties of the used species 213 C.1 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.1 Oil properties . . . 217
D.1.1 Thermal conductivity . . . 217
D.1.2 Heat capacity . . . 217
D.1.3 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 Heat capacity . . . 223
D.2.4 Thermal conductivity . . . 223
D.3 Metal properties . . . 227
E Measurements for validation process 229
F Measurements for optimization process 233
References 237
List of Figures
2.1 Mercedes 35 hp . . . 9
2.2 Split cooling architecture (1) . . . 13
2.3 Valve controlled cooling 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 heat exchanger . . . 30
3.4 Coolant circuit scheme . . . 31
3.5 Foucault currents brake . . . 33
3.6 Test Bench Scheme . . . 34
3.7 LabView interface for temperature monitoring . . . 35
3.8 Fuel mass ow rate measuring device . . . 36
3.9 Thermodynamic loss angle denition (3) . . . 39
3.10 Polytropic coecient vs. crank angle for dierent indications (4) . . . . 40
3.11 Polytropic coecient 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 ow rate vs. PWM . . . 50
4.1 Burnt fuel mass fraction as a function 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 coecient variation as a function of the crank angle (8) . . 64
4.5 Heat ux and emissivity for three dierent crank angles (9) . . . 65
4.6 Valve pressure drop coecient as a function of the valve opening (9) . . 68
4.7 Valve lift law (10) . . . 69
4.8 Mass ow rate during the intake stroke as a function of the crank angle degree for dierent engine rotation speeds . . . 71
4.9 Primary atomization of a water jet (11) . . . 73
4.10 Jet penetration as a function of time (11) . . . 74
4.11 Spalding parameters as a function of the droplet temperature . . . 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 equation validation (9) . . . 82
4.16 Combustion kinetic (7) . . . 82
4.17 Calculated temperature 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 Simplied scheme of the oil hydraulic circuit . . . 96
5.5 Heat transfer coecient 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 temperature for the dierent nodes . . . 109
5.11 Crankcase oil temperature as a function of the coolant mass ow . . . . 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 decomposition (13) . . . 121
5.18 Nodal decomposition for the "Diagonal circulation" . . . 122
5.19 Nodal decomposition for the "Pin circulation" . . . 124
LIST OF FIGURES
5.21 Simulink node conguration . . . 137
5.22 Simulink cycle model conguration . . . 138
5.23 Coolant temperature history at the engine inlet and outlet . . . 139
5.24 Node temperature history . . . 142
5.25 Node mass inuence 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-𝜃 diagram (4) . . . 151
6.3 Setting points for global model validation (14) . . . 153
6.4 Outlet coolant temperatures with initial model (14) . . . 155
6.5 Inlet coolant temperatures with initial model (14) . . . 155
6.6 Oil temperatures with initial model (14) . . . 156
6.7 Heat transfer coecients between piston and oil for initial and modied correlations (14) . . . 160
6.8 Modeling of the power evacuated by the radiator (14) . . . 163
6.9 Final results for inlet temperature (14) . . . 164
6.10 Final results for outlet temperature (14) . . . 165
6.11 Final results for oil temperature (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 ow with and without control (14) . . . . 185
7.4 Temperatures for dierent nodes at 40 kW and 3130 rpm with coolant ow control (14) . . . 185
7.5 Predicted indicated eciency increase (14) . . . 186
7.6 Predicted eective eciency increase (14) . . . 187 7.7 Oil temperature after cold start for dierent initial coolant ows (14) . 188
7.8 Eective power benet with coolant ow control after a cold start (14) 189 7.9 Comparison between classical coolant ow rate and optimized coolant
ow rate . . . 191 7.10 Fuel mass ow rate improvement at 30 kW of eective power . . . 192 7.11 Eciency improvement vs. temperature increase . . . 194 7.12 Comparison between optimized and classical coolant ow rate for an
European Driving Cycle . . . 195 7.13 Coolant temperature evolution for four consecutive European Driving
Cycles with and without optimized coolant ow rate . . . 196 7.14 Comparison between optimized and classical coolant ow rate for an
Extra-urban Driving Cycle . . . 199 7.15 Coolant temperature evolution for three consecutive Extra-urban Driving
Cycles with and without optimized coolant ow rate . . . 200 A.1 Combustion chamber volume (9) . . . 207 D.1 Engine oil heat capacity as a function of the temperature . . . 218 D.2 Engine oil kinematic viscosity as a function of temperature for 15W40 oil 219 D.3 Density of water/glycol mixture as a function of the temperature . . . . 222 D.4 Kinematic viscosity of water/glycol mixture as a function of the temper-
ature . . . 224 D.5 Heat capacity of water/glycol mixture as a function of the temperature . 225 D.6 Thermal conductivity of water/glycol mixture as a function of the tem-
perature . . . 226
List of Tables
3.1 Engine main characteristic . . . 26
3.3 Gear ratio values . . . 44
3.4 Laplace transform 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 Coecients for the Woschni equation (9) . . . 64
4.3 Intake and exhaust valve characteristics . . . 69
4.4 Variable engine rotation speed results . . . 70
5.1 Equivalent conductance for the dierent transfer modes . . . 92
5.2 NTU heat 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 temperatures for an engine rotation speed of 2000 rpm . 108 5.8 Engine - pump pressure coecients . . . 116
5.9 Radiator parameters . . . 116
5.10 Typical coecients for convective heat transfer between cylinder walls and coolant ow . . . 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 ow rate distribution in the cylinder head for a "diagonal circulation"133 5.15 Mass of the dierent mechanical components . . . 135
5.16 Coolant volume repartition . . . 136
5.17 Coolant circuit power balance at a given steady state condition . . . 140
5.18 Coecients⟨𝛼𝑖⟩ for this engine geometry . . . 145
6.1 Experimental results for thermodynamic model validation (4) . . . 149
6.2 Comparison between simulated and measured values for thermodynamic
model validation . . . 152
6.9 Cylinder wall & coolant temperatures for dierent in-cylinder heat trans- fer coecients . . . 173
6.10 Indicated powers & optimized coolant ow rates for dierent in-cylinder heat transfer coecients . . . 174
6.11 Indicated powers for dierent discharge coecients . . . 174
7.2 Fuel economy expectation . . . 190
7.5 Fuel economy expectation . . . 201
C.1 Coecients giving the enthalpy as a temperature function for the used species . . . 214
C.2 Air properties . . . 215
C.3 Fuel physical properties . . . 216
D.1 Coecients for oil heat capacity correlation . . . 217
D.2 Coecients for oil kinematic viscosity correlation . . . 218
D.3 Coecients for oil density correlation . . . 219
D.4 Coecients to compute the density as a function of the mixture temper- ature for dierent glycol concentrations . . . 222
D.5 Coecients to compute the viscosity as a function of mixture temperature for dierent glycol concentrations . . . 224
D.6 Coecients to compute the heat capacity as a function of the mixture temperature for dierent glycol concentrations . . . 225
D.7 Coecients to compute the thermal conductivity as a function of the mixture temperature for dierent glycol concentrations . . . 226
D.8 Metal Properties . . . 227
E.1 smallcaption . . . 231
F.1 smallcaption . . . 235
Glossary
˙
𝑚 Mass ow [𝑘𝑔𝑠]
𝜆 Regular pressure drop coecient for smooth ducts [-]
𝐶𝑓 Heat capacity mass ow [𝑊𝐾] 𝐻𝑙 Engine heat loss [J]
𝑛𝑝 Polytropic coecient [-]
𝑃𝑚𝑓 Mean friction pressure [Pa]
𝑅𝑚 Clankpin radius [m]
𝛼 Thermal diusivity [𝑚𝑠2] 𝛽 Dilatation coecient [1/K]
Δ𝐻𝑟 Pressure drop [m]
Δ𝑡𝑖𝑛𝑗 Injection time [s]
𝛿 Droplet diameter [m]
Δ𝑥 Node thickness [m]
𝜖 Emittivity[-]
𝜂 eciency [-]
𝛾 Ratio between heat capacity at con- stant pressure and specic volume [-]
𝜇 Dynamic viscosity [𝑚𝑠𝑘𝑔] 𝜈 Kinematic viscosity [𝑚𝑠2] 𝜔 pulsation [𝑟𝑎𝑑𝑠 ]
Φ Heat ow [W]
𝜙 Jet angle [radian]
Ψ Radiative heat transfer coecient [𝐾𝑊4]
𝜌 Density [𝑚𝑘𝑔3]
𝜎 Stefan-Boltzman constant [𝑚𝑊2𝐾4] 𝜏 Ignition delay [ms]
𝜃 Crank angle [radian]
𝜉 Volume fraction [-]
𝜉𝑟 Pressure drop coecient [-]
a Crankshaft length [m]
AFR Air to Fuel Ratio [-]
B Bore [m]
B𝑇 Thermal Spalding parameter [-]
B𝑌 Species Spalding parameter [-]
C𝐷 Discharge coecient [-]
c𝑝 Heating capacity at constant pres- sure [𝑚𝑜𝑙𝐾𝐽 ]
c𝑣 Heating capacity at constant specic volume [𝑚𝑜𝑙𝐾𝐽 ]
C𝑥 Aerodynamic coecient [-]
D Diameter [m]
D∗ Species diusivity coecient[𝑚𝑠2] E Heat exchanger eciency [-]
e Thickness [m]
E𝑎 Activation energy [𝑚𝑜𝑙𝐽 ] F Heat ux [𝑚𝑊2]
F𝑟 Force [N]
f𝑟 rolling resistance coecient [-]
g Gravity [9.81 𝑚𝑠2 ]
G𝑖𝑗 Equivalent conductance between node i and j [𝑊𝐾]
Gr Grasho number [-]
H Enthalpy [J]
h Molar enthalpy [𝑚𝑜𝑙𝐽 ] h𝑐 Convection coecient [𝑚𝑊2𝐾] h𝑒 Height [m]
h𝑚 Mass enthalpy [𝑘𝑔𝐽]
IMEP Indicated Mean Eective Pres- sure[bar]
k Thermal conductivity [𝑚𝐾𝑊 ]
KS Global heat exchange coecient [𝑊𝐾]
L Length [m]
l Connecting rod length [m]
L𝑚 Wet perimeter length [m]
L𝑣 Vaporization Heat [𝑘𝑔𝐽] LHV Low Heating Value [𝑘𝑔𝐾𝐽 ] M Molar mass [𝑚𝑜𝑙𝑒𝑘𝑔 ]
m Mass [kg]
m𝑖𝑛𝑗 Injected fuel mass [kg]
N Engine rotation speed [rpm]
n Droplets number [-]
n𝑐𝑦𝑙𝑖𝑛𝑑𝑒𝑟𝑠 Number of cylinders [-]
NTU Number of Transfer Units [-]
Nu Nusselt number [-]
P Power[W]
p Pressure [Pascal]
p𝑥 Partial pressure of species x [Pascal]
Pr Prandtl number [-]
Q Heat [J]
q𝑣 Volumetric ow [𝑚𝑠3]
R Perfect gas constant [8.31𝑚𝑜𝑙𝐾𝐽 ]
r Radius [m]
Ra Rayleigh number [-]
Re Reynolds number [-]
RR Reaction Rate [∘𝑘𝑔𝐶𝐴] S Area [𝑚2]
s Stroke [m]
T Temperature [K]
T Torque[Nm]
t Time [s]
U Internal energy [J]
u Molar internal energy [𝑚𝑜𝑙𝐽 ] u𝑚 Mass internal energy [𝑘𝑔𝐽] V Volume [𝑚3]
v Speed [𝑚𝑠] v𝑟 Radial speed [𝑚𝑠]
W Work[J]
x Jet penetration [m]
Y𝑥 Molar fraction of species x [-]
Le temps d'un monde ni commence.
Paul Valéry
Introduction 1
Currently the world is facing major challenges leading political world to reduce drasti- cally𝐶𝑂2 emissions and fuel consumption.
∙ more and more people are living in urban areas;
∙ more and more people will have access to personal vehicles (especially in emerging countries like China or India);
∙ 𝐶𝑂2 emissions aect 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 nd other ways to produce long term, greener, renewable energy sources.
Regarding these worldwide trends, the dierent political actors have taken mea- sures to decrease both 𝐶𝑂2 emissions and oil consumption. The most famous one is the Kyoto Protocol. In order to achieve the dierent objectives set by the Protocol, re- gional political decision makers have published laws and regulations for the automotive industry.
The objectives set in the Kyoto Protocol are 20% reduction of total𝐶𝑂2emissions by 2020 compared to the emissions of the year of reference - 1990.. In the European Union, private cars are responsible for 12% of the total 𝐶𝑂2 emissions. The environmental commission of the European Parliament (16) has then xed emissions objectives for cars with a reference mass of less than 2610 kg. These objectives are: 120 g 𝐶𝑂2/km
by 2012 and 95 g 𝐶𝑂2/km by 2020. The current level is 160 gr 𝐶𝑂2/km. That is to say 25% reduction by 2012 and 41% reduction by 2020. In terms of fuel consumption 160 gr 𝐶𝑂2/km corresponds to 6.03 l/100 km (on the base of classical diesel fuels) or , 120 g 𝐶𝑂2/km to 4.52 l/100 km and 95 g 𝐶𝑂2/km to 3.58 l/100 km. This huge reduction 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 Economy (or CAFE) standards for passenger cars.
These standards have not changed since 1990 and are still set at a minimum of 27.5 miles/gallon (8.54 l/100 km). These regulations have in fact more economical reasons (control of the energy balance) than ecological ones as the US have not signed the Kyoto Protocol.
In Japan, the government has dened objectives for gasoline engines of 21.4% con- sumption reduction by 2010 in comparison with the levels of 1995. For diesel engines, the objectives are set at 13.1% consumption reduction. The dierence of objectives between gasoline and diesel comes from the fuel consumption dierence between these two engines. It is then expected that gasoline engines and diesel ones will have a fuel consumption of respectively 6.5 l/100 km and 8.3 l/100 km (18).
As it can be seen, environmental, energy and economical constraints have forced political decision makers to put heavy requirements on car manufacturers. The objec- tives of 20% consumption reduction in the European Union and Japan are particularly ambitious. New technologies in every vehicle eld 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 eective 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 resistance and 30% in the rolling resistance (tire deformation). These values are based on a S𝐶𝑥 of 0.64 (actualSedan car) and a tire coecient of 6 kg/t (actuel low energy consumption tire). This repartition is highly speed dependent, a
driver's behavior or the use of the vehicle (urban or rural) is also determining for the vehicle consumption. Indeed, the inertia eect can represent up to 50% of the energy consumption for urban driving.
The possibilities of eciency improvement can be drawn from these two energy balances:
∙ the indicated torque can be increased with a higher engine compression ratio or better injection technologies;
∙ heat losses can be reduced with an optimized cooling (with a cooling ow 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 resistance can be lowered thanks to an optimal vehicle design;
∙ the rolling resistance can be reduced with low hysteresis tires;
∙ the driver's behavior through his average speed or his speed variations (inertia eect) can greatly inuence the fuel consumption;
∙ the use of the vehicle can be improved with the implementation of a "Stop & Start"
strategy, the use of biofuel blends, automatic gear boxes and cruise control.
All 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 temperature dependent) and mechanical losses by reducing the power consumption of the water pump.
The architecture of this study is as follows. First of all, in the chapter two, a bibli- ographic study will present the history of engine cooling, the current main technologies
and the recent improvements. This work follows the Phd Thesis of Frederic Pirotais (10) who has presented a model of the same engine without studying the link between the engine thermal state and its eciency.
To assess the impact of new cooling strategies, a model has been built. This model takes into account the interaction between the combustion, the wall heat uxes 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: rst 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 dierent engine.
As the results of the model will be compared to measurements, the model parameters are set for a specic 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 implementation 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 ve. In this chapter, the interaction between the two models through the cylinder sleeves heat uxes and tem- peratures is also presented as well as its implementation 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- metric eciency and the combustion law are adapted to match the measured pressure history for a whole range of engine rotation speed and eective power. Concerning the nodal models, global measurements such as coolant temperature, oil temperature and fuel consumption are performed. The model architecture and heat ux laws are then modied in order to match the measurements for steady state conditions; the nodes weights are modied to match the measurements for transients conditions. The
The model exploitation is presented in chapter seven. A new cooling strategy is pre- sented and implemented on the developed model. The expected benets for the dierent 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 benets 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.
I think cars today are almost the exact equivalent of the great Gothic cathedrals. I mean the supreme creation of an era, conceived with passion by unknown artists, and consumed in image if not in usage by a whole population which appropriates them as a purely magical object.
Roland Barthes
2
Engine cooling state of the art
This chapter will present the history of the combustion engine and how cooling tech- nologies have evolved in parallel. It is particularly interesting to note that cooling exists since the very rst development of the internal 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 heat 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 modern computer era that the rst 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 ow control. This current research topic is of rst importance since the engine eciency has become the main requirement for the engine design.
Finally, a section is dedicated to the approach of Frederic Pirotais, its limitation and how the present work adds the eciency prediction to that category of models.
2.1 History
The development of cooling circuit started just after the development of the rst internal combustion engine. The rst engines in the modern 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 Edward Somerset with a vapor cooling system in 1663.
A milestone in the engine history is the invention of the piston by Denis Papin with a rst commercial model in 1712. The principle of cooling 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 developments brought the engine history one step further. The works of the Niepce brothers led to an engine called Pyréolophore (19) based on heat transfer only. The water vapor was not the working uid 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 rst engine based on the combustion of an inammable gas.
The principles of thermodynamic and thermal engines have then been theorized in the book of Sadi Carnot published in 1824: "Réexions sur la puissance motrice du feu et sur les machines propres à développer cette puissance". From this moment, the experimental development of gas red 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 rst water cooling system based on the thermosiphon eect was created on the piston engine of Otto and Langen (20). The development of cooling technologies based on either air or water heat transfer then continued during the following years with the works of Gottlieb Daimler and Willem Maybach which led in 1884 (20) to:
∙ the creation of a cylinder with blades,
∙ the invention of the rst ventilator based on blades mounted on the crankshaft,
∙ the rst engine with dierentiate cooling: cylinder head with water and cylinder walls with air.
In 1892 with a growing eective power, the amount of heat release led to the rst heat exchangers: the rst water coil appeared on a Daimler engine. This invention preceded by a few years the rst radiator patented in 1897 and installed in 1901 in
2.1 History
had another breakthrough feature: the water was provided by a pump driven by the crankshaft. The cooling architecture has almost not changed since then. This car achieved 35hp at 950 rpm with a maximum speed of 75 km/h (21).
Figure 2.1: Mercedes 35 hp
Some new developments appeared along the XXth century leading to the current classic cooling circuit, among them the rst 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 sucient 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 lled as soon as possible especially in critical area like mountain passes.
Many manufacturers have 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 internal combustion en- gines. Even if the problem of boiling was solved, the air-cooled engine continued to be known as the most reliable cooling solution (as with the Volkswagen Beetle).
The two dierent cooling methods (water and air) continued to be developed during
the following years. In 1961, the Renault 4 was equipped with the rst pressurized water cooling circuit with a water-coolant mixture. The boiling point was then delayed with a higher temperature. In 1969, Porsche introduced an air turbine for engine cooling on its 917 model. 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 rst 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 eciency 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
All the recent improvements aim to increase the engine eciency, to improve the passen- ger thermal comfort, to reduce the vehicle emissions, to decrease the warm-up time and to reduce the engine temperature uctuations. This goals can be achieved by several devices which are detailed hereunder.
2.2.1 Valve control
J.R. Wagner associates 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 temperatures according setting points determined by a two- dimensional lookup table based on the engine speed and load.
In (27), an electric temperature control valve was designed. This control valve is controlled to minimize the temperature variations depending on load and trac condi- tions. The management of this valve is based on temperature sensors and the vehicle speed measurement.
In the same approach as the one presented in (24), predictions are made in (28) on the eect of an electronically controlled proportional valve aiming at keeping metal temperature under a specic value. The conclusion stated that this valve should produce an higher engine eciency and reduce thermal stress and fatigue.
Pursuing the goal of combustion performance promotion, the potential of electronic
2.2 Recent improvement in engine cooling architecture
reference, an electronic control valve is associated to an electrical water pump in an experimental study.
As there is a potential for the heater core to be insucient at low pump 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 coolant ow and to allow coolant temperature control.
Finally, as during the rst 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-lled 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), dierent innovative engine cooling systems are compared. Among these, the THEMIS System and Coolmaster both use an electrical coolant pump to achieve coolant ow 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 ow rate. However, some other techniques are studied and presented hereafter.
Nucleate boiling The nucleate boiling show great potential 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
Departure from Nucleate boiling. In such a case, the heat transfer coecient decreases abruptly. Therefore the cylinder wall temperature can increase above its thermal limit.
Heat storage In order to reduce the engine thermal uctuations, the possibility of a heat load averaging system has been studied in (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 coolant inventory can be decreased and permits faster warm-up time.
Variable ow It has been shown that the coolant ow rate can 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 eciency of the cooling system by delivering appropriate ow 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. volumetric ow)
Heat exchangers The Ultimate cooling System presented in (33) proposes to use engine coolant exclusively to cool all the uids in the vehicle i.e. refrigerant, oil, exhaust gas and fuel. The main benets are a packaging reduction 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 eciency increase lies between 4 and 6 % according to reference (36).
The split cooling rst goals were: to reduce warm-up, pollutant emissions and to increase the compression ratio in spark ignition engine (1). Figure 2.2 shows a typical split cooling architecture.
2.3 Modeling history
Figure 2.2: Split cooling architecture (1)
2.3 Modeling history
The history of engine heat transfer modeling started in 1939 with the work of Nusselt and Eichelberg (37). These rst 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 temperature was assumed and the heat transfer was expressed as:
𝜙=ℎ(
𝑇𝑔𝑎𝑠−𝑇𝑤𝑎𝑙𝑙𝑠)
(2.1) where the convection coecientℎ was computed through semi-empirical correlations in order to match the global measurement of heating power.
As the computation capacity increased, the rst attempts to solve the heat equation (Equation 2.2) started.
∇(𝑘Δ𝑇) =𝜌𝑐𝑝∂𝑇
∂𝑡 (2.2)
But the high frequency of the boundary conditions and the high spatial discretization led to huge computing resources and time request and was thus not developed as a global approach but limited to local components modeling.
These two approaches were reconciled (38) by using the Eichelberg approach in the components where steady state thermal behavior is achieved and by using equation 2.2 in components aected by the transient heat transfer. This last zone is called the penetration depth. This method gave satisfactory results for local heat transfer through surfaces with irregularities.
Beside these specic 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 complete 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, all the devices where modeled with thermal equations such as the ones described in Chapter 5. The main dierence 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 uxes based on a mean wall temperature which is xed for all the simulations. There is no feedback between the thermodynamic and the nodal model. Fuel consumption cannot be assessed, as the indicated eciency is not inuenced by the nodal model results.
Finally, must be cited the 1-D black box model developed for the engine designers.
These models are based on a global energy balance. These softwares provide libraries where the dierent components of the cooling circuit can be chosen and their parameters designed. These models have a limited physical description but give reliable results and are cost eective. 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 dierent 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
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 heat transfer correlation can be very dierent.
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 various paths upon which energy is exchanged between nodes. The main criteria chosen for this 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 inuence of local temperatures on the engine eciency) or global (e.g. the study aims at determining the minimum coolant ow rate to evacuate enough heat 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 metal masses are divided in two:
one separating the burning gases from the coolant owing 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 nally 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 correlations are still used with good simulation results. For instance, the correlation of Woschni is used in (41) and (28). On the other side, the
discrepancy between all the semi-experimental correlations is high as shown in (42).
2.3.3 Combustion model
To assess the eect of the thermal management on the engine eciency, one of the crucial part is the combustion model.
This model can be as simple as an algebraic correlation giving the heat release.
These models like the one of Taylor and Toong are still used in recent studies (24).
This simplied approach is also used in (32) with good results on global value like warm up time.
Single zone models which compute pressure and temperature history in the com- bustion chamber can be used as in (28). However, a heat release law has still to be dened.
The use of such heat 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 heat transfer correlation.
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 function of the engine rotational speed and load.
Otherwise, a dynamic coupling can be achieved. This approach, chosen in (27), demonstrates all the benet obtained with a fully coupled approach even if the com- plexity of the model is increased
2.4 Current cooling strategies
2.4 Current cooling strategies
The current cooling strategies consist in designing the pump for strenuous cycles such as maximum power or low load at low speed. The pump characteristic is then chosen to provide sucient coolant ow rate in order to keep the coolant temperature below its boiling point. These cycles can be for instance race conditions, the engine is then used at its maximum power, sand hole1 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 ow 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 have a lower eciency. 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 inecient compared to the current control devices. The system does not respond properly to environ- mental changes such as ambient temperature.
1The vehicle has to go out of a sand hole by riding in circles
2.5 Cooling control technologies
The rst developments on cooling control appeared in the second half of the eighties.
Three dierent controls can be distinguished:
∙ valve controlled cooling,
∙ air ow controlled cooling,
∙ pump controlled cooling.
The valve controlled strategy is the easiest to put into practice. The two expected benets are the indicated eciency increase by lowering the thermal losses and the eective eciency increase by lowering the friction losses. A study performed in 1988 (2) used this technology to raise the coolant temperature up to 150 ∘C. A high boiling point (180 ∘C) coolant uid 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 temperature. The thermal losses are smaller as the thermal gradient between the combustion gas and coolant is lower. The measured benets 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
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 rst complete 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 ux through the radiator. The system manages all the heat ows to improve eciency and passenger comfort. All the actuators are driven by a micro-controller which receives all the sensors information. Again the main dierence with a classical cooling system is the coolant temperature 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)
It can be seen that eective benets in fuel consumption can be achieved (around 3
%). The rst 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 demand of the coolant pump,
∙ reduce the warm-up time,
∙ maintain the engine at high temperature to improve indicated and eective e- 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 consequences of this strat- egy are: a lower pump energy demand (lower pressure drops) and a lower warm-up time;
2. maintain the coolant ow over a minimum value until the coolant temperature reaches its reference value in order to avoid hot points in the engine;
3. avoid large temperature dierences between engine outlet and inlet which leads to large temperature gradients in the engine. These gradients can lead to non uniform expansion;
4. control the outlet temperature in case of high loads or demanding external con- ditions (high ambient temperature).
The controlled parameter can be either the coolant temperature (24), (28) or the temperature of one metallic part (28), (29).
The coolant pump can even be stopped during the engine warm up as shown in (28) and (29).However this strategy must be avoided in order to prevent hot points which
