Cooling System Development and Optimization for DI Engines

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SAE TECHNICALPAPER SERIES 2000-01-0283

Cooling System Development andOptimization for DI Engines

Franz W. KochFEV Motorentechnik

Frank G. HaubnerVKA

Reprinted From: Vehicle and Engine Systems Models(SP–1527)

SAE 2000 World CongressDetroit, Michigan

March 6-9, 2000

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2000-01-0283

Cooling System Development and Optimization for DI Engines

Franz W. KochFEV Motorentechnik

Frank G. HaubnerVKA

Copyright © 2000 Society of Automotive Engineers, Inc.

ABSTRACT

The reduction of the fuel consumption and the emissionsare the two main goals for the development of currentand future engines. Both consumption and emissions are

highly influenced by the fluid and the material tempera-tures of the engine. This offers potential especially at lowengine speeds and engine loads to reduce the coolingpower and increase the material temperatures to a tribo-logic and thermodynamic optimized level.

The cooling system which is able to control the coolingpower and the material temperatures, the required con-trol devices and the control strategy are designated asintelligent heat management. The definition of therequirements for the control devices and the definition ofthe control strategies requires detailed knowledge aboutthe thermal engine behavior.

To determine the optimized operating of the intelligentheat management system and define the characteristicsof the control devices numerical simulation models havebeen established at FEV and verified by experimentalinvestigations. These models comprehend the combus-tion process, the heat rejection to the different enginecomponents and also the tribologic system.

The paper will present a detailed thermal engine model,describing the heat fluxes within the engine and the inter-action of the material temperatures with the thermody-namic and the tribological system. A control strategy forthe intelligent heat management and the benefits in

terms of fuel consumption and emissions will be pre-sented.

INTRODUCTION

Today more than 75% of all travels in Germany areshorter than 10 km as depicted in figure 1 /IKA/.

Figure 1. Travel distance

This corresponds to a low required vehicle drive powerreaching approx. 5 kW averaged for a typical city cycle asdepicted in figure 2 /IKA/. The maximum drive power isbelow 20 kW for a middle class vehicle corresponding toapprox. 20% of the maximum engine performance.

Out of the high share of part load operation, the reductionof the friction losses is the most promising way to reducethe fuel consumption significantly, as well for the low partload steady state operation as for the engine warm up.For both the main parameter influencing the engine fric-tion losses is the coolant temperature. The intelligentcontrol of the coolant temperature is one way to reducethe fuel consumption and also the emissions.




Figure 2. Required engine performance city cycle

A system comprehending the knowledge to operate thecooling system, the required devices to adjust the coolingpower and the strategy for this control is designated asIntelligent Cooling System (ICS) and will be presented inthe following.

The intelligent heat management system is used to con-trol the coolant temperatures according to the optimizedtribological and thermodynamic temperature. To deter-mine this optimum the influence of the material tempera-tures on the following has to be known:

• gas exchange

• combustion• heat transfer to the combustion chamber walls• engine friction losses• engine wear, especially the piston and piston rings

This realization of the ICS presumes the knowledgeabout the energy losses due to

• heat losses during combustion from the gas to thechamber

• energy losses due to incomplete combustion• energy losses due to the engine friction losses

THERMAL BASICS

The requirement of the cooling system is caused by therestriction of the maximum material temperatures in com-bination with maximum gradients leading to thermalstresses superimposed on the mechanical loads. Tem-peratures and gradients are caused by the heat lossesfrom the combustion to the combustion chamber follow-ing the Newton approach. With respect to the increasingengine performance of the latest and future TDI enginesup to 60 kW/l, the demands to the cooling system arehigher. The detailed knowledge of the thermodynamic

and thermal engine behavior is absolutely necessary forthe development.

The heat losses during the complete cycle leads to theenergy distribution in the combustion chamber. Basicallythe following energy transport mechanism are possible:

• heat rejection to the cylinder head• heat rejection to the piston• heat rejection to the liner• heat rejection to the exhaust ports

The heat rejection to the liner is the combined thermalload due to the direct heat transfer from the combustionchamber, the energy transport to the liner from the pistonvia the piston ring and the piston skirt and the frictionlosses of the piston and the piston rings. It is assumedthat the friction losses of these components are trans-ferred to the liner completely.

The thermodynamic basics are required to determine thenet heat release from the combustion process to thecombustion chamber. The analysis is based on the pres-

sure trace, the valve timing, the crank train parametersand the intake air temperature. Further the fluid proper-ties have to be considered.

To determine the heat transfer coefficient the followingexpression (Woschni-formula) is used.

with

• D liner diameter• p [N/m2], T [K] local averaged pressure and temper-

ature during combustion• p1 [N/m2], T1 [K], V1 [m3] condition at start of com-

pression• p0 [N/m2] pressure w/o combustion (for motored

engine)• Vh [m3] cylinder displacement• c1 [-] = 6,18 + 0,417 c u / c m gas exchange• c1 [-] = 2,28 + 0,308 c u / c m combustion• c2 [m/sK] = 3,24 E-3• cu / c m swirl number

With knowledge of the average pressure and tempera-ture of the gas inside the combustion chamber and alsotaking the wall temperature into account, the heat flux tothe chamber walls can be determined. Especially for theliner and the surface with direct contact to the gasdepends on the crank angle.

The interaction of the heat transfer from the piston to theliner is more complex, depending on the local tempera-tures and the heat transfer by the oil film on the liner. Fur-ther both temperatures influence the oil film temperatureand thereby the friction losses.

( )α i mh

D p T c c cV Tp V

p p= + −− −0 013 0 2 0 8 0 531 2

1

1 10

0

, * * * * * ***

*, , ,




Actually a closed analytical description of these interac-tions is not available, therefore simulation calculationsare very much based on experience.

The piston itself has three different possibilities to transferheat, by

• the piston rings• the piston skirt• the bottom to the oil in the crankcase

Pistons with an oil cooling duct have the possibility toreduce the material temperatures especially near the firstcompression ring by an effective cooling. In recent TDIengines with a specific performance of about 50 kW/l,pistons with cooling duct are used. To form the coolingduct the carrier of the first compression ring can be used

/MTZ/. Usually the cooling duct can be introduced to pis-tons with axissymmetric piston bowl, which is used for 4valve engines.

The distribution of the heat transfer from the piston to theliner depends on many parameters, e.g. heat conductivityof piston and ring material, heat transfer between pistonand ring, liner temperature and heat transfer betweenring and liner. Figure 3 depicts the measured heat flux fora TDI engine /KLA/.

Figure 3. Heat flux to the liner /Kla/

The energy balance of the liner gains special attentionwith respect to the high influence of the liner temperatureon the friction losses. Basically, the heat rejection fromthe combustion chamber and from the piston has to be

carried off, by• forced convection to the coolant• conductivity to the crankcase housing and then to the

ambient and the oil• natural / forced convection via the engine surface

• the engine oil which is splashed to the liner from thecrank train, especially if the piston is near TDC

The energy transport to the coolant can be describedanalytically by the heat transfer coefficients for natural / forced convection determined and verified in many exper-imental investigations.

At full load, the maximum material temperatures near tothe coolant passages are above the critical temperature

at which nucleate boiling starts. This very efficient heattransfer mechanism depends on several parameters.Today no complete analytical description is available, theinfluence of the surface roughness, the pressure and thefluid proper ties was investigated experimentally.

To describe the heat transfer by forced convection in atube, the following expressions /VDI/ can be used, basedon the Nusselt number with Nu = α *l /λ

laminar flow with Re d < 2300

turbulent flow with Re d > 2300, 0,6 < Pr < 500, l/d > 1

• d = tube diameter• l = tube length• Re = Reynolds number• Pr = Prandtl number = η*cp / λ • η = fluid viscosity• ηW = fluid viscosity at wall temperature

Basically, laminar and turbulent flow can occur in a com-bustion engine, depending on the engine speed, the cool-ant flow rate and the local coolant passage design andthe coolant flow velocities. The maximum material tem-peratures observed in TDI engines, especially the cylin-der heads at the usual pressure level to operate coolingsystems, makes it quite sure that also nucleate boilingoccurs. The very high heat transfer coefficients achievedby this effect can protect the engine as long as the vapor

can be carried off the hot spot. As soon as vapor remainsat one position the heat transfer coefficient is reducedsignificantly, usually leading to fatal damage of the engine(burn out).

This effect has a special importance for the operation ofthe ICS at low loads when the engine can be operated w/ o coolant flow to the block. In this case the heat transfer

Nu

dl

d

l

d

d

dW

= +

+

3 660 19

1 0117

0 8

0 467

0 14

,, Re Pr

, Re Pr

,

,

,η

η

( )( )Nudl

d dW

= − − +

0 0235 230 18 0 8 10 8 0 32 3 0 14

, Re , Pr ,, , / ,

ηη




to the coolant is based on natural convection followingthe expression

with Gr the Grashof number

The rapid increase of the heat transfer coefficient has tobe considered especially for the analysis of the materialtemperatures by FEM which have to be done in order todetermine the lowest possible coolant flow rate andthereby the maximum material temperatures.

ENGINE THERMAL BEHAVIOR

Basic for the layout of the cooling system is the maximumengine performance. The cooling system requires suffi-cient capacity to carry off a certain amount of heat inorder to keep all fluid and material temperatures withinthe individual limits. Further the function of all compo-nents, e.g. the plain bearings or the piston group, set lim-its to the maximum temperatures.

The ratio of the cooling power to the engine performanceis an important parameter to assess the cooling system.This ratio is designated as cooling number if the heatrejection to the coolant is divided by the engine perfor-mance and extended cooling number if all heat lossesare divided by the engine performance.

cooling number

extended cooling no.

The (extended) cooling number can be obtained bydetermination of the energy balance, which has beendone by FEV on several series and prototype TDIengines. The extended cooling number versus load forvarious engine speeds of a typical 2,0 l TDI engine isdepicted in figure 4. Obviously, for a conventional systemthe extended cooling number varies between two asymp-totics, which are the y-axis at idle and a parallel to the x-axis. The latter meet the y-axis at approx. 0.6 to 0.8,which corresponds to the 1/3 rule, saying that app. 1/3 ofthe fuel energy is distributed to the engine performance,the exhaust enthalpy and the heat losses. The coolingnumber reaches values of approx. 0.4 - 0.5, so the maxi-mum cooling power reaches approx. half of the maximumengine performance.

Figure 4. Extended cooling number TDI engine

Beneath the heat rejection to the coolant the engine cool-ing is achieved also by the engine oil, especially at thepiston and in the lower part of the liner, which has directcontact to the oil droplets moved by the crank train com-ponents. The heat rejection to the oil is on a level thatmakes the use of an oil cooler necessary. The heat trans-fer from the oil to the coolant is depicted versus speedand load in figure 5.

Figure 5. Heat flow oil to coolant for a TDI engine

Finally, the heat losses from the combustion chamber aretransferred via the material directly to the ambient by nat-ural or forced convection, depending on the local air flowvelocity within the engine compartment. The oil pan hasusually a high share of the convection losses the high oiltemperature and the position.

Another heat sink is the intercooler, which gains impor-tance especially at higher loads and speeds. This energyis also transferred directly to the air flow through theengine compartment. The different energy losses lead tothe energy balance which is depicted for an engine speed

of 2000 rpm in the figure 6.The cooling power shows a nearly linear dependency onthe engine speed and the engine load which is shown infigure 7.

( )Nu Grx x= +

0 508

0 952

1 41 4

, *Pr

, Pr Pr

/ /

( )Gr

g Lwall fluid

= −

β ϑ ϑ

ηρ

3

2

η =&Q

Pcooling

e

η extendedfuel e exhaust

e

Q P HP

= − −& ∆

25 50 75 100 125 150 175 200 225 250 275 3001000

15002000

25003000

35004000

0

2

4

6

8

10

12

14

16

a d d i t i o n a l o

i l c o o

l e r c o o

l i n g p o w e r

[ k W ]

engine load [Nm]

engine speed [rpm]

14-16

12-14

10-12

8-10

6-8

4-6

2-4

0-2




Figure 6. Energy balance at 2000 rpm

With the cooling power a function following the expres-sion

the influence of the load can be isolated. At constantspeed the coolant flow rate and thereby the flow veloci-ties and also the heat transfer coefficients (presumesforced convection) are nearly constant. The coolant wet-ted surface is identical. Out of this can be concluded thatan increase of the difference between the material andthe coolant temperature is required to achieve a highercooling power. Exactly this effect can be used in theopposite way to increase the material temperature byreducing the coolant flow rates.

Figure 7. Cooling power of a TDI engine

The measurement of material temperatures for variousspeeds and loads show that for each operating point a

certain material temperature is reached depending alsoon the distance to the combustion chamber and to thecoolant jacket. The analysis of the temperatures showsthat the position within the engine dominates the differ-ence of the material temperature to the coolant tempera-ture.

INFLUENCE OF THE COOLANT TEMPERATUREON THE MATERIAL TEMPERATURES

The comparison of material temperatures at differentpositions shows nearly identical temperature gradientsversus load and speed, meaning constant temperaturedifferences between any two positions within the engineindependent of load and speed.

The material temperatures within today’s engines areruled by the coolant temperatures, due to the use offorced convection with high heat transfer coefficients fromthe material to the coolant. Further, for heat transfer fromthe combustion chamber to the coolant a certain differ-ence in temperature is required, depending on the heatflux, the transport direction (one or more dimensional),the material properties and the local heat transfer coeffi-cients on the gas and the coolant side.

Figure 8 depicts the material temperatures within a DIengine cylinder head versus coolant temperature. Thetemperatures have been measured in the valve bridgebetween intake and exhaust valve. The material tempera-

tures depend linearly on the coolant temperature, anincrease of 10 K leads to a temperature rise of approx. 8K. This linear dependency was observed nearly identicalon other TDI engines as well as on gasoline engines / KRU/.

Figure 8. Influence of the coolant temperature on headtemperatures

Figure 9 shows an identical behavior within the engineblock. Again the gradient of the material temperatures is8 K in the material for 10 K in the coolant, independent of

the position within the engine.

& * *Q A Tcooling =α ∆




Figure 9. Influence of the coolant temperature on blocktemperatures

This knowledge can be used to reduce the required infor-mation to control the engine material temperatures in thevehicle. If the engine thermal behavior and the interac-tions within the engine are described once, only oneblock temperature and one cylinder head temperature isrequired to control the intelligent management system.

With respect to today engines mostly equipped with acoolant temperature sensor at the engine outlet but nomaterial temperature sensors, a further transfer of theinteraction coolant temperature and material temperatureis required.

INFLUENCE OF THE COOLANT TEMPERATUREON THE THERMODYNAMICS

An important question arise from the influence of thecoolant temperature on the required cooling power. It isexpected that by reduction of the Newton’s temperaturedifference from the gas to the chamber wall, the heattransfer is reduced. Despite this, investigations at the adi-abatic engine have shown that the opposite can beobserved, explained with higher heat transfer coefficientsnear to the wall. /WOS/

Thermal investigations carried out at FEV on severalengines showed that temperatures at different locationswithin the engine are nearly identical influenced by theoperating point. Further it was observed that the temper-ature level in the engine ruled by the coolant temperaturehas only a minor influence on the heat rejection to the

material and thereby to the coolant.Figures 10 and 11 depict the change of the heat flow, thegas temperature within the combustion chamber and thechange of the heat transfer coefficient versus coolanttemperature. The temperature shows the expecteddecrease, compensated by an increase of the heat trans-fer coefficient. The influence of the coolant temperature

on the wall temperature can be seen also in the averagedgas temperature, which is app. lowered by 16 K for acoolant temperature decrease of 20 K, so the averagedgas temperature is ruled by the liner material tempera-ture /KRU/.

The heat transfer coefficient shows a marginal increaseof approx. 1 % in comparison to the basic, which can beexplained by the temperature influence on the heat trans-fer coefficient corresponding to Woschni.

All in all the heat flow increases by approx. 1% if the cool-ant temperature is decreased by 20 K,. Based on thisinvestigations can be stated that the coolant temperaturelevel has a minor influence on the heat rejection to thecombustion chamber walls.

Figure 10. Heat flow versus coolant temperature head

The same influence was observed for the engine block,as depicted in figure 12. Again the heat rejection to theliner changes by 1% if the coolant temperature isdecreased by 20 K.

This is very important for the reduction of the engine fric-tion losses. With the piston group friction losses alsoruled by the liner temperatures, these again by the cool-ant temperature, it is obvious that the coolant tempera-ture has the main influence on the engine friction lossesand is thereby the most promising parameter for optimi-zations.




Figure 11. Heat flow versus coolant temperature block

INFLUENCE OF THE COOLANT TEMPERATUREON THE ENGINE FRICTION

The engine friction losses are highly influenced by theengine speed and the fluid temperatures. Based oninvestigations carried out at a high number of TDIengines, the friction mean effective pressure (FMEP) hasbeen inter- and extrapolated for the complete operating

range and is depicted versus speed and oil temperaturein figure 12.

The complete engine friction losses depicted in figure 12are caused by the different engine components

• crankshaft, meaning the plain bearings and the radialsealing

• piston group meaning the piston, the piston rings andthe con rod bearing

• valve train, comprehends the valve drive, the cam-shaft bearings and the valves

• accessories, comprehend the water pump, the oil

pump and the alternatorThe analysis of the engine friction losses split to the dif-ferent components show a high influence of the fluid tem-peratures on the crankshaft and the piston group. Thesecomponents have a share of approx. 70 - 75% of thecomplete engine friction.

Figure 12. FMEP of the complete engine versus enginespeed and oil temperature

The increase of the fluid temperatures from the conven-tional 90°C to 110°C leads to an engine speed depen-dent reduction of the engine friction losses by approx. 4 -10%.

The friction losses of the piston group can be reducedfurther by an increase of the liner material temperatures.This increase can be achieved by reduced heat transfercoefficients from the liner to the coolant which is possibleby reducing the flow rates and thus the cooling power,which requires a control of the coolant flow rate depend-ing on the required cooling power.

By an increase of the liner temperatures a further reduc-tion of the piston group friction losses up to 10% isexpected for the low part load. With increasing engineload the cooling power has to be increased to keep thematerial temperatures at an acceptable level. Thisreduces the potential of friction reduction by increasedmaterial / liner temperatures.

Another restriction of the increase of the liner tempera-ture arise from the wear of the piston and the piston ringswith reduced oil viscosity due to increased oil film tem-peratures. It is expected that the areas of possible mixedfriction near TDC and BDC increase if the capability ofthe oil film is reduced. The influence on the friction lossesmay be low according to the low piston velocity. Theforces acting on the piston can still reach high values,which have been observed by investigations with FEV-PIFFO (Piston Friction Force Measurement). FEV-PIFFOis a unique tool to measure the axial force of the pistongroup acting on the liner for fired operation.

Figure 13 and 14 depicts the comparison of the mea-sured axial piston group forces at 2000 rpm enginespeed for a coolant temperature of 60°C and 90°C.

The engine was motored (figure 13) and fired (figure 14).




Figure 13. Piston group friction force (motored)

Figure 14. Piston group friction force (full load)

Obviously the forces are lower at 90°C coolant tempera-ture except the range near TDC. This corresponds to areduction of the piston group friction losses by app. 20%

for the temperature rise of 30 K.Due to the lower capability of the oil film to carry the pis-ton rings and the piston by hydrodynamic pressure directcontact to the liner can be expected, meaning mixed fric-tion. This is also expressed by the friction coefficient „µ“describing the oil film shear stress and the ratio of frictionforce (axial Force) and force normal to the liner. Thiscoefficient is usually on a level of app. 0.05. Near TDChigher values can be observed indicating that the pistonand the liner and/or the piston rings and the liner havedirect contact, which in most cases is combined withwear.

Predictions of wear are very difficult, therefore assump-tions have to be made about the wear rates and the influ-ence of the coolant temperature on this rate.

The increase of the tribologic relevant temperatures atthe liner and the plain bearings means an increase of thelocal oil temperatures. The oil film temperature on theliner surface is ruled by the liner material temperature,which itself is ruled by the coolant temperature especiallyfor engines with a 100% water jacket (100 % means thewater jacket is as long as the travel of the 1 st compres-

sion ring meaning the complete stroke and also at thesame height within the block). Again the predominantinfluence of the coolant temperature on the friction lossescan be observed.

INVESTIGATIONS OF THE ENGINE WARM UPBEHAVIOR

The high influence of the fluid temperatures on the fric-

tion losses and thereby on the fuel consumption drawsthe attention to the engine warm up offering very highpotential of fuel consumption and emission reduction.

The high fuel consumption of an engine after cold start isinfluenced by the coolant as well as the material thermalinertia. Experimental and theoretical investigations haveshown that both parameters have nearly the same influ-ence on the warm up behavior. As part of the enginedesign, the engine operation has to be considered, espe-cially the interaction with the passenger compartmentlinked via the heater.

The warm up time for today’s TDI engines show espe-

cially at low ambient that the nominal operating tempera-ture is not reached when the heater fan is turned on. Inthis case approx. 4-5 kW heater performance arereached at the end of the ECE-cycle. Figure 15 depictsthe coolant outlet temperature of a state of the art TDIengine and the ECE-cycle speed profile versus time dur-ing the ECE-cycle. Further the heater performance isdepicted.

Figure 15. Basic TDI engine warm up

Obviously the engine does not reach the usual 90°C

coolant outlet temperature but only approx. 60°C. Furtherthe heater performance of 4,2 kW is reached afterapprox. 12 min, a time after that 70% of all cycles are fin-ished already, so it can be stated that the heat rejection tothe coolant is insufficient for the winter operation of TDIengines.

Obviously the thermal inertia influences the warm up onlyuntil the balance temperature level is reached afterapprox. 800 s. Measures to reduce the warm up timetherefore can only partly contribute to an improvement of




the thermal behavior. Additional measures are requiredto increase the available heat to reach the nominal oper-ating temperature of the engine as well as sufficientheater performance. The possible measures to reducethe warm up time and increase the balance temperaturelevel are presented in the following:

• reduction of the coolant volume, the engine massand improved isolation of the engine

• use of additional heater systems• use of the engine control unit to manipulate the

engine thermodynamics and thereby the thermalbehavior

The use of additional heater systems, especially the fuelburner has already been brought into series production.Detailed descriptions can be found in literature. Basicallyall the additional heater systems lead to an improvedwarm up behavior but also an increased fuel consump-tion. The most promising system for future TDI enginescomprehends the exhaust heat exchanger in combinationwith a higher exhaust backpressure. This concept leadsto an increased gas exchange work, shift the engineoperating point to higher IMEP´s and increases theexhaust temperature. Still as for all additional heater sys-tems the package and the costs for the components aredrawbacks that have to be considered in addition to thefuel consumption.

The use of the engine control unit ECU to increase theheat rejection to the coolant by switching to a cold startmapping is already part of today’s TDI vehicle applica-tion. This solution has the advantage to avoid additionalcomponents but is restricted by the influence on theemissions and the driveability. With respect to the futureemission level EU IV, this solution is too much restrictedto be a real opportunity for the future.

REQUIRED INFORMATION TO CONTROL THESYSTEM

The information required by the intelligent cooling systemare listed in the following. All these information and corre-sponding sensors are already available in today’sengines, allowing an easy access to the data by the ECU.Consequently the integration of the ICS into the ECUshould be the target for the further development. Alterna-tively the CAN bus allows the data transfer, with the ECUanswering the requests of the ICS within a certain time.

Important and in today’s engines already available infor-mation to control the ICS are:

• intake air, coolant and oil temperature, NTC resis-tance

• engine load and speed• air flow meter• EGR-valve sensor

• vehicle velocity• heater control• starter signal• oil temperature, gear box• brake light• idle sensor

The above listed information could be used to control theICS and the electrical components. Based on the inputdata, the coolant flow distribution can be directly adjustedby referencing predefined mappings. The determinationof these mappings has to be done by detailed investiga-tions of the thermal engine behavior.

If the interactions and influences of load and speed andthe coolant temperature on the engine performance isknown, the control of the flow rates overall the engineoperating range can be realized as shown in figure 16.

Based on the input data the ECU/ICS determines therequired coolant flow distribution within the completecooling system and thereby the coolant in- and outlet

temperature at all components by use of predefined map-pings. The cooling power and heat rejection and materialtemperatures result from the conditions on the coolantside as has been presented in chapter 3.

Figure 16. Control procedure of the ICS

OPERATION OF THE ICS / COOLING STRATEGY

The ICS is based on two major types of information. Firstis the measured parameters as listed in the last chapter.Second the description of the cooling power dependingon these information. This description is designated ascooling strategy.

The detailed strategy presumes the knowledge of thethermal behavior of the engine. This can be based onmeasurements as well as on simulation. Obviously, thesimulation model has to be capable of representing allinteractions that con occur overall the engine operatingrange.

Input

ECU / ICS

electricalwater pump

radiator,inner circuit valvesblock valve

flow rates of the pump, to the block, to the radiatoras function of

engine speed and engine load•ambient temperature / air temperature•fluid temperatures / material temperature




The main parameters influencing the material tempera-tures within the engine are

• engine speed• engine load• coolant temperature inlet / outlet• coolant flow• engine friction, at part load and during the engine

warm up, at maximum engine performance or near tofull load the influence diminishes

Minor influence can be observed for the following param-eters

• oil temperature, as function of the material tempera-tures, friction and the coolant temperature, especiallywith respect to the use of the oil cooler

• ambient temperature, especially in the engine com-partment

• engine compartment air flow

Based on the input information the ECU/ICS has toadjust the total coolant flow rate to the engine and withinthe cooling system to control the cooling power of the cyl-inder head and the engine block individually and alsocontrol the cooling inlet and outlet temperatures.

In the following an example is given for the steady stateoperation. The flow rates and the coolant temperature atthe engine outlet is given based on investigations on a2,0l TDI engine.

The figures 17 and 18 depict the required flow rates tothe cylinder head and the engine block. Basically thedependency of the cooling power of the speed and theload is visible also in the flow rates. Further in the low

speed and load range the cooling power especially of theengine block can be achieved by natural convection, sothe coolant flow rate can be reduced to zero.

Figure 17. Mapping of coolant flow rate to cylinder head

Figure 18. Mapping of the coolant flow rate to the engineblock

Figure 19. Mapping of the coolant outlet temperature

Finally, figure 19 depicts the coolant outlet temperatureversus load and speed. The increased coolant tempera-ture at part load is required to take full advantage of thereduced friction losses by increased liner temperatures.The reduction of the flow rate alone leads to a more eventemperature distribution within the engine versus loadand speed and shift all temperatures especially at lowpart load to higher values. Still the required liner tempera-tures can be reached only by increasing the coolant tem-peratures.

REQUIRED CONTROL DEVICES

The intelligent heat management system requiresdevices to control the coolant flow rates within the systemas well as the fluid and thereby the material tempera-tures.

The solution giving the most degrees of freedom to adjustthe flow rates in each line of the cooling system wouldrequire one proportional valve for each line. For the com-plete system with split cooling of block and head and aparallel assembly of the external heat exchangers five(six) lines = radiator, oil cooler, heater, inner circuit, EGR-cooler, (optionally exhaust heat exchanger). With thevalve for the engine block in total six (seven) proportionalvalves would be required.




This would allow the deletion of the conventional thermo-stat and also the use of the conventional water pump. Inthis case a bypass to the system would be required withan additional valve.

A possible assembly is depicted in figure 20. The figureshows two valves, with the one upstream of the pumpcontrolling the heat transfer to the ambient and therebythe coolant inlet temperature and the one downstream ofthe pump controlling the cooling power within the engineblock and the cylinder head by variation of the flow rates.Alternatively, figure 21 shows the system with an electricwater pump instead of the bypass to the pump.

Figure 20. Cooling system with conventional water pump

Figure 21. Cooling system with electrical water pump

The electric water pump allows the engine to operatewithout any flow to the engine, thus no energy losses tothe pump. The flow to reach the required cooling powercan be adjusted independent from the engine speed.This can be used as well for worst case conditions asafter engine shut down.

For future TDI engines with high efficiency, the use ofadditional heater systems gains more importance. The

preheating systems especially the fuel burner, require a2nd water pump to the main pump, which is usually anelectrical one with relative low hydraulic power. In thiscase, the drawback of higher costs for an electrical waterpump can be compensated by use of an electrical pumpas main pump.

The integration of various functions into one housingwould mean for the future concept an assembly consist-ing of the radiator, degas system, all proportional valvesand also the pump into one housing. Here also the elec-tric high end devices would be located.

SIMULATION OF THE ENGINE WARM UP

Based on experimental investigations, a model has beenestablished in order to simulate the function of an ICSys-tem on a state of the art TDI cooling system. Therefore,the engine warm up has been simulated for the ECEcycle (the cycle is repeated after 800 s) with a startingtemperature of -7°C. Special attention was paid to thecoolant outlet temperature and the fuel consumption ver-sus time. The simulation model of the basic cooling sys-tem was set up according to figure 20.

The basic warm up and fuel consumption for the enginestart leads to the warm up curve depicted in figure 21.The warm up of the engine with ICS was started w/ocoolant flow both to block and cylinder head. For the ICS,the coolant flow rate through the engine block and thecylinder head was adjusted individually depending on thematerial temperature. After the tribologic favorable tem-perature level was reached the flow rates were

increased to keep the material temperature nearly on aconstant level. Obviously the warm up time of the mate-

rial can be reduced significantly by the use of the intelli-gent cooling system.

Figure 22. Engine warm up at -7°C

The control of the flow rates is based on the block andhead material temperature which explains the tempera-ture fluctuations. To avoid these fluctuations the ICS hasto foresee and adjust the required flow rates. In order toavoid rapid temperature fluctuations in the block andhead a smooth mixture of the cold coolant of the externaland the hot coolant of the internal system is required.




This means special requirements for the speed to controland adjust the flow rates. The flow rates to the cylinderhead and the engine block that are required to keep thematerial temperatures nearly constant on the leveldepicted in figure 21 are shown in figure 23 and 24.

Figure 23. Flow rate to the block during warm up

The flow rate to the block is kept at zero for the first fiveminutes of the warm up. After the required material tem-perature of the block of 130°C is reached the ICSincreases the flow rapidly but to a very low value ofapprox. 0,45 l/min. After a short period the flow rate isreduced again. The figure demonstrates the highdemands to an intelligent system in terms of controlspeed and accuracy. The situation to control the flowrates within the cylinder head is similar to the block.Again a rapid increase of the flow rate is required alreadyafter two minutes.

Figure 24. Flow rate to the head during warm up

The engine warm up at -7°C was simulated w/o use ofthe heater. The use of the heater means completely dif-

ferent boundary conditions for the ICS. In this case theavailable energy to heat up the system has to be distrib-uted to the engine and the passenger compartment. Forthis purpose the simulation model offers the possibility todefine and test various control strategies and optimizethe system performance with respect to fuel consump-tion, emissions and comfort. To investigate the influenceof the passenger compartment heating on the warm up of

the engine block and cylinder head the heater was usedfor the following simulation. Again the simulation starts at-7°C, the coolant flow to the engine block is controlled toreach the quickest possible warm up and thereby the low-est fuel consumption. The coolant flow to the cylinderhead was controlled based on the coolant outlet tempera-ture instead of the material temperature. The ICS adjuststhe flow rate linear increasing with the coolant tempera-ture.

Finally the air flow through the cabin heater is controlledbased on the air outlet temperature. With increasing airoutlet temperature the air flow is increased and reachesthe maximum value, which corresponds to state of theart, at a temperature of 40°C. The coolant temperaturesof the block and the head are depicted in figure 25.

Figure 25. Engine warm up with cabin heating

Figure 26 depicts the coolant flow rate to the engineblock versus time, which is nearly identical to the basicwarm up.

Figure 26. Flow rate to the block during warm up

The coolant flow to the cylinder head is transferred to thecabin heater downstream of the engine. This means acompromise between reaching the shortest possibleengine warm up leading to a low fuel consumption andalso reaching sufficient cabin heater performance. Theflow rate to the cylinder head is depicted in figure 27 andshows significant higher flow rates in comparison to thebasic warm up.




Figure 27. Coolant flow rate to the head

With the flow rate also the energy transport to the cabinheater is achieved. The heat transfer to the cabindepends also on the heater efficiency and the tempera-ture depending flow rates both of the air and the coolantflow. The controlled air flow through the cabin heater andthe air outlet temperature is depicted in figure 28, show-ing that during the first 100 s the air flow can be reduced

to a minimum value due to the low temperature level ofthe coolant. After 400 s the conditions at the cabin heaterare nearly constant with an air temperature at the outletof approx. 30°C.

Figure 28. Heater air flow and outlet temperature

Based on the air flow at the heater and the temperaturedifference from inlet to outlet the heater performance wasdetermined. The heater performance during the warm uptime is depicted in figure 29. During the first 200 s theheater performance shows a linear increase and reaches2.5 kW. After this time the air flow and air temperature atthe heater outlet is nearly constant. Also the coolant flowrate and the coolant outlet temperature at the cylinderhead is nearly constant. The heater performancereaches a mean value of approx. 3.2 kW. This value islow in comparison to the required heater power of state ofthe art middle class vehicles, reaching up to 6-7 kW, indi-cating the requirement of additional measures to improvethe warm up behavior of the engine and the passengercompartment.

Figure 29. Heater performance during warm up

CONCLUSION

The design of engines in terms of friction reduction waslong since part of the mechanical development. Recentlythe potential of further optimization for the part load oper-ation as it becomes more and more evident, leads to afunctional optimization. Therefore the improved adapta-tion to the tribologic and thermodynamic demands at part

load gains more importance. By implementing the Intelli-gent Cooling System (ICS) the warm up time can bereduced significantly, thereby the fuel consumption andthe emissions are reduced significantly. Also for part loadoperation the increased operating temperatures offer areduction up to 5 % of the fuel consumption.

Beneath the positive influences on the warm up behaviorand the fuel consumption and emissions the demands tothe control components are high, especially with respectto the speed and accuracy necessary to adjust the cool-ant flow rates.

SUMMARY

The reduction of the fuel consumption e.g. by reducedfriction losses, is beneath the reduction of the emissionsthe most important goal of the development of combus-tion engines. FEV has developed CAE-tools to investi-gate the thermal engine behavior and predict theinfluence of optimization measures. These tools havebeen used to establish simulation models, especially forTDI engines. One example representing the state of theart for TDI engines was presented in this paper. The ther-mal, thermodynamic and tribologic basics are presentedand discussed. Based on this description a simulationmodel was set up and used to predict the engine warmup behavior for a conventional and a modified coolingsystem, designated as Intelligent Cooling System (ICS).The comparison shows a significant reduction of thewarm up time leading to an important reduction of thefuel consumption.

The investigations demonstrate that the controlled cool-ing system is a promising way to reduce fuel consump-tion and emissions, especially by reducing the frictionlosses at engine part load and during the engine warmup.




REFERENCES

1. /IKA/ Vorlesungsumdruck Kraftfahrzeuge I, RWTH,Aachen, 1995

2. /KRU/ Krüger, Michael, Untersuchungen zur ther-mischen Bauteilbelastung kleiner schnellaufenderDieselmotoren mit indirekter und direkter Einsprit-zung, Dissertation RWTH Aachen, 1995

3. /VDI/ VDI Wärmeatlas, 5. Auflage

4. /WOS/ Woschni, G.; Kolesa, K.;Spindler, W., Iso-lierung der Brennraumwände - Ein lohnendesEntwicklungsziel bei Verbrennungsmotoren? MTZ 47(1986) 12

5. /KLA/ Klaus, Benedikt, Untersuchung des Wärme-transports vom Kolben über die Ringe und die Zylin-derbüchse zum Kühlmittel, Dissertation TUMünchen, 1996


Cooling System Development and Optimization for DI Engines

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