A Methodology for the Design of Engine Cooling Systems in Standalone Applications

ABSTRACTIn this paper, a methodology for the design process of enginecooling systems is presented, which is based on theinteraction among three programs: a code developed forradiator sizing and rating, a 3D commercial code used for theair circuit modeling, and a 1D commercial code used for themodeling and simulation of the complete engine coolingsystem. The aim of the developed methodology, in additionto ensure the system thermal balance, is the improvement ofthe design process of the cooling system itself, whileshortening the development times, in non-automotiveapplications. An application to the design of a locomotiveengine cooling system is presented. The system designed hasbeen assembled and tested, showing the validity of themethodology, as well as the compliance of the designedsystem with the initially specified thermo-hydraulicconstraints and requirements.

INTRODUCTIONWhile in automotive applications the cooling system isdesigned for the majority of operational conditions andcoolant temperatures are managed the by actively controllingcoolant flow and engine output [1, 2], in specific purposeapplications, such as the case of locomotive engines, thetrend is still to size the cooling package so that sufficient heatis rejected at extreme operating conditions (full engine

power, high ambient temperature). Nevertheless, many of thetechnological practices matured in automotive thermalmanagement are now being implemented in specificapplications [1]. Along with this, there is also an increasingdemand to reduce product development times, to respond tomarket trends and to reduce investments. In this context,development engineers are encouraged to develop conceptualdesign strategies for the rapid assessment of the wholethermal management system at the component, system, andapplication levels, based on the use of CAE tools, in order toachieve design objectives and specifications for differentsystem aspects. The models have to be kept simple,especially compared with state of the art radiator calculationmethods, if acceptable calculation times are to be achieved[3].

The computational cost of commercial 1D simulationprograms is small when compared to complete system 3Dsimulations. Thus, 1D codes are useful to conduct simulationof large systems; however, mechanical components and heatexchangers are considered only as momentum and energysources or sinks, so that the detailed behavior of the heatexchanger must be studied in a 3D environment. This meansthat both 1D and 3D simulations are necessary in order toachieve both requirements (simulation of large systems anddetailed evaluation of each component). The internal flow ina component, simulated in 3D, can be incorporated into aglobal 1D cooling system network. With this, the details of

A Methodology for the Design of Engine CoolingSystems in Standalone Applications

2010-01-0325Published

04/12/2010

Antonio J. Torregrosa and Pablo OlmedaUniversidad Politecnica de Valencia

Antonio Garcia-RicosVossloh España

Jordi NatividadVossloh

Carlos A. RomeroUniversidad Tecnologica de Pereira

Copyright © 2010 SAE International

Downloaded from SAE International by Automotive Research Association of India, Friday, August 01, 2014

the internal flow are taken into account while conservingoverall mass flow in the system, thus reducing uncertaintiesin boundary conditions prescribed in the 3D model andreducing the overall simulation time.

In this work, an integrated engineering approach for thedesign of special application engine cooling systems ispresented, with emphasis put on the role of simulation fromthe conceptual phase onwards. Theoretical data was used forthe development of a radiator calculation program;commercial CFD PowerFlow code [4] was used for thedesign of the air circuits, while the flow models of both thecooling air and the coolant were constructed undercommercial 1D FlowMaster simulation environment [5]. Theiterative coupling of these computational tools resulted in aco-simulation methodology that was applied to the design ofthe locomotive engine cooling system, and allowed a rapiddefinition of a solution, avoiding the need for extensiveprototype level testing.

The paper is organized as follows. First, an overview of theparticularities of the general design process is given, togetherwith a description of design methodologies for engine coolingsystems. Then, the preliminary design procedure is outlined.This step should provide the basis for the detailed calculationor selection of components such as coolant pumps, fans andradiators; also, an outline of the particular air ventilation path,as well as coolant piping topology in accordance with theapplication, should be produced. Subsequently, a descriptionof the program developed for the selection and rating of heatexchangers, with application to engine radiators, is presented.Then, the application of the methodology to the design of thecooling system of the locomotive engine is described, withspecial emphasis put on the air flow and global thermo-hydraulic cooling system modeling. Finally, the mostimportant conclusions are presented, including somesuggestions for future work.

DESCRIPTION OF THE DESIGNPROCESSThe design and fabrication of mass produced engine coolingsystems follow the flow diagram sketched in figure 1 [6].

Figure 1. Global flow diagram

Figure 2. Cooling system design methodology

The process starts with the identification of the applicationobject to be designed, after which the physical limitations areset and the performance characteristics are dimensioned.During this stage it is likely that technical information onmain system components comes from technical literature andsuppliers' catalogs. The following phase comprises thepreliminary activities intended to size the main components(pumps, fans, and radiators), as well as to providespecifications for the selection and purchasing of originalequipment. Based on the results of the preliminary design,several sketches of the system under design must be drawn,modeled and simulated, with the aim of analyzing andselecting the most suitable variant that optimally matches thevarious components necessary for overall cooling systemperformance. In the general design process a certain numberof prototypes must be developed and tested, depending on thefield of application.

As far as the locomotive engine application is concerned,prototypes are not practical, because of the small lotproduction, and therefore computational modeling andsimulation represent a great advantage in terms of time and


resource savings, and the general scheme of figure 1 isreduced to the CAE development methodology, based on thepreliminary design (there is no allowance for the prototypemanufacture and testing). In essence, the handling of thecooling system design methodology is described by figure 2.

In this methodology, the radiator design and modelingprogram, as well as the 3D air flow modeling are coupled tothe global 1D thermo-hydraulic model of the engine coolingsystem. The 1D model can be linked to a detailed air flowmodel of the ventilation duct, which in turn provides theboundary for the 1D model. In this way, a reasonableapproach to the study of the thermal management system isachieved. The results are coolant and air pressure,temperature and flow distribution as average values for allsystem nodes and branches. The better the knowledge of theparts of the cooling system, the more accurate are the results.

The 3D CFD program provides local steady flow results,from which the momentum drop and thus the pressure dropcharacteristic of the component, required as an input by the1D program, can be calculated. Once the 1D computation hasfinished, new pressure and mass flow rate on the boundariesare solved and fed back to the 3D program and the radiatorcalculation program. The final result of this coupled iterativecomputation is achieved when values of design parametersand thermo-fluid properties at the coupling boundariesconverge satisfactorily.

COOLING SYSTEM DESIGNMETHODOLOGYOnce the application, the system specifications and theperformance requirements are known, the design processstarts by making a number of assumptions allowing for thepreliminary design of a coarse engine cooling system layout.Typically, as the quantity of air and its outlet temperature arefunctions of the radiator core geometry, trial an errorprocedures have been used, which may be avoided with theuse of the radiator sizing program, which has the ability toperform parametric studies, including core type sweeps. Afterselecting initial radiators and fans, simulations are performedin which preliminary sizes and components are modified untilall operating conditions are satisfied. First, a fan is selectedand the 3D simulation of the ventilation circuit is carried out.By selecting a radiator core, the air velocity and pressuredrop distributions along the ventilation path are obtained andfed to both the radiator calculation program and the globalthermo-hydraulic 1D model of the cooling system. Thethermal dissipation characteristics of the radiators are used intheir corresponding 1D sub-models, together with otherthermo-hydraulic information and then the whole system issimulated, providing the coolant flows and temperatures,which are then compared to the specified values.

This iterative process not only allows assessing the heatsource-sink balance, but also gives relevant insight into thehomogeneity of the distribution of air temperatures and flowsas the air crosses the radiators. After this, issues on thereconfiguration of the cooling tract, the arrangement andnumber of radiators and fans, among other technicalmeasures, can be considered. In the end, the bestconfiguration of the radiators and the geometry of theventilation duct can be obtained. Of course, the final designmust account for other aspects, such as space requirements,weight, easiness of installation, accessibility, noise emission,etc.

MODELLING PROGRAMSA brief account of the programs used is given in thefollowing.

3D CFD programThe airflow system is necessary to remove waste heat fromthe engine and any heat exchangers present, such as thecharge air cooler (aftercooler), engine oil cooler, transmissionoil cooler (where available), etc. Thermal interactionsbetween the air side system and oil, transmission, and coolantloops are supported by the 1D code.

Because of the complex nature of the flow field, the variationof the drag and heat transfer over the whole surface of theradiator must be obtained from CFD simulations. In thiswork, commercial software [4] was used for the simulation ofthe air circuit in the engine cooling system, which solves thethree dimensional flow equations governing conservation ofmass, momentum and energy and the k-ε turbulence model.The radiators are modeled as porous media, with the flowrestricted to one direction through the radiator cores. The corepermeability is represented by specifying a quadraticrelationship between pressure drop across, and velocitythrough, the core. A similar relationship between heat transfercoefficient and velocity can also be specified.

1D thermo-hydraulic programAir and coolant parts of the cooling system are notindependent. The 3D computation must use the global flowconditions as boundary conditions for the detailed local flowcalculations and the local pressure drops must be used in theglobal flow calculations. Therefore, an iterative process isrequired for a solution of the overall system. In the presentwork, a computational 1D program designed to predictpressure, temperature and fluid flow distribution, amongother important parameters in thermo-hydraulic systems, wasused [5]. Its physical basis is built on empirical correlationsas well as on analytical equations for mass, momentum andenergy. A large database of elements, expandable by the user,is available to describe relevant components of thermo-


hydraulic systems such as pipes, pumps, heat exchangers,valves, etc.

Radiator Calculation programThe radiator is designed for the required heat flows inconjunction with the mass flows of the coolant and thecooling air, as well as their temperature. Radiator heattransfer equations are available from many sources[7,8,9,10,11], in numerous forms; discussion of the details ofthermal and hydrodynamic correlations upon which radiatorsizing is based is beyond the scope of this work. Followingthe methodology described by Kays and London [7] for thedesign of compact heat exchangers, and taking into accountother relevant literature [12,14], a computational program forthe design of engine radiators has been developed. Theradiator calculation program includes correlations for thecalculation of all main cross flow heat exchangers, includingautomotive radiators, and is intended for the sizing and ratingcalculation, the calculation of thermo-physical properties, andthe execution of different parametric and sensitivity studies,together with pressure loss dependencies. The flow diagramin figure 3 illustrates the calculation procedure.

Input data - During the analysis of a heat exchanger it isnecessary to specify: coolant type, coolant and air massflows, coolant and air temperatures, allowable coolant and airpressure losses, required heat transfer rate (it does not have tobe set when the size is a constraint), radiator type,configuration and geometry of the core, and overalldimensions (height, width, depth).

Output results - The output protocol of the radiator displaysthe used input values along with the calculated output values.Among the data provided by the program, the most relevantare:

• Input parameters: input temperatures and volumetric flowrates, maximum pressure drops, and geometric parameters ofthe radiator.

• Operation parameters: temperatures of input and outputprocess fluids, heat duty (thermal characteristics), overallheat transfer coefficient, total heat transfer area, mean fluidvelocities, and pressure losses.

• Construction parameters of the cross flow radiator: ductlengths, cross sectional area of ducts, tube pitch, fin densityand others.

Figure 3. Simplified scheme of design and ratingcalculation of the radiator

PRELIMINARY DESIGN OF THEENGINE COOLING SYSTEMThe purpose of the preliminary design of the engine coolingsystem is the determination of the radiator total heat transfersurface area, the pump performance curve, and the selectionof the fan. Since, depending on the application, the systemcan incorporate other heat exchangers besides the radiator(aftercooler, EGR cooler, oil cooler, etc.), during thepreliminary design more than one of these components mayneed to be selected in accordance with the arrangement of thesystem, and also different compromises can becomenecessary when arranging the fluid flow architecture.

HEAT DISSIPATED TO THE COOLANTThe amount of heat to be removed from the engine by thecoolant results from the balance between the various heatflows that characterize the internal combustion processes,with the particularities associated with the thermal control ofthe air intake process. This heat amount is usually unknownfor the designer. Hence, its estimation is the first step in thepreliminary design procedure.

The heat rejected by the engine can be estimated by using adetailed thermal model (using nodal finite difference of finiteelement techniques) [15,16]. If sufficient information is notavailable, then some correlations are available: Taylor andToong [17], as well as Lahvic [18] proposed some of themost frequently cited correlations in literature. Also,correlations based on nominal power or fuel consumption canbe established. The possibilities analyzed in this work were:

• Lahvic's empirical correlation [18], which can be applied tospark ignition and Diesel engines; here, Vd (l), n (rpm), MT


(Nm), Ne (kW) are the engine displacement, speed, torqueand brake power of the engine, respectively:

(1)

• The model based correlation of Parish [19], which requiresmore detailed information of the engine operation, withconsideration of the heat generated by the friction. In itsnormalized presentation related to the charge inducted intothe cylinders, (Vd·/2), the correlation for the specific heat

rejected to the coolant, , has the formgiven in equation (2), where σ represents the Stefan-Boltzmann constant; D is the cylinder diameter; kg, Pr are theconductivity and Prandtl number of the gas in the cylinder,respectively; Re is the Reynolds number characteristic of the

flow in the cylinder; and are the mean values of the

gas and flame temperatures, respectively; and is thecombustion chamber wall temperature.

(2)

• Based on commercial literature related to enginespecifications, a relationship between effective power andheat rejected to the coolant was obtained, as it is depicted infigures 4 and 5, for the heat rejected to the coolant, throughthe engine water jacket, and the heat rejected to the coolantby the engine aftercooler, respectively.

Figure 4. Heat dissipated by the engine to the coolantflowing through its water jacket.

Figure 5. Heat dissipated by the aftercooler to thecoolant.

RADIATORIn general the preliminary design of the radiator can becarried out in the following sequence:

• Calculate the heat to be dissipated in the radiator makingallowance for the dirtying process in the radiator surfaces, as:

(3)

• The heat carried off by the air is assumed to be equal to the

energy dissipated by the coolant ( )

• The volumetric air flow through the radiator core, after

fixing the air temperature increase, °C,

is given by equation (4) where and are the airspecific heat and density, calculated for an assumed mean air

inlet temperature of 40 °C:

(4)

• The volumetric coolant flow rate through radiator is given

by equation (5), where and are the coolantspecific heat and density calculated for an assumed meancoolant temperature through the water jacket;

is the coolant temperature drop acrossthe radiator, which ranges usually from 6 to 12 °C.

(5)

• The mean coolant temperature in the radiator is computedas:


(6)

• The mean air temperature as it passes through the radiator isgiven by:

(7)

• The required radiator surface area is found as:

(8)

Here, is the overall heat transfer coefficient, based onthe wetted surface [20]:

(9)

where is the convective heat transfer coefficientbetween the coolant and radiator surfaces, δ is the wall

thickness, k is the thermal conductivity of the material, isthe convective heat transfer coefficient between the air andradiator surfaces; ψ is a fin coefficient. For automotive

radiators the values for range from 140 to 180 W/(m2K) [20].

• The frontal surface area of the radiator can be found asgiven by equation (10), where is the air speed at theradiator front (typical values range from 6 to18 m/s [20]), nottaking into account the vehicle speed.

(10)

• The radiator width is now given by equation (11), where is a volumetric factor ranging from 0,6 to 1,8 mm−1

[20].

(11)

• The heat dissipated by the radiator per inlet air-coolant

temperature difference, , is nowexpressed as:

(12)

• The thermal performance of the radiator can becharacterized by the heat flow per inlet temperaturedifference per unit frontal area, as a function of coolant massflow rate and air mass flow rate per unit frontal area:

This sequence is used to calculate the initial designparameters either for a detailed design of the engine radiatoror for its selection from a manufacturer's catalog, restrictingalso the allowable pressure losses. The methodology couldstart from the global dimensions allowed for the radiator, independence of the available space.

FAN - The information required for the selection of fans isthe static pressure rise, (Pa) and the volumetric air flow

rate (m3/s) at the nominal operating point, in addition tothe operating curve relating the change in pressure across thefan to the volume flow through it, the tip and hub diametersand the fan blade shape: straight or twisted. The axial flowpropeller-type fans are those most used in engineapplications. The data needed is easily available from fansuppliers.

The volumetric air flow rate has been calculated in theprevious paragraph, while the pressure must be found afteranalyzing a preliminary air flow layout, and considering theprinciple of minimum power consumption. In any case,pressure drop across the radiator core must be less than 25mm water and most usually in the 12 to 18 mm water range[20]. Also, in the preliminary stage of the cooling systemdesign, correlations for pressure drop for discrete componentsof the air flow path, as well as tabulated information availablein the specialized literature, can be used to estimate per-component and total pressure drop across the air duct.Depending on the application of the cooling systems thestatic pressure developed by fans ranges from 600 to 1450 Pa[21]. The fan tip speed is determined as [20]:

(13)

where ψa is a blade shape coefficient, ranging from 2,2 to 2,9for curved blades, and from 2,8 to 3,5 for straight blades. Fornoise considerations u ≤ 110 m/s [22].

The preliminary fan diameter can be determined as:

(14)

Here, is the calculated air flow velocity through therotor. This depends on the relationship between the effective


area swept by the fan blades, and the radiator frontal area. Forthe systems in use, this relationship is in the range from 0,45

to 0,6, whence ranges from 13 to 40 m/s [22].

The rotational speed of the fan is determined as:

(15)

The power consumed by the fan can be now assessed byassuming the fan efficiency (typical values range from0,6 to 0,7):

(16)

The performance of the fan may be given in a non-

dimensional form, by the flow, , and pressure, ,coefficients defined by:

(17)

Equations (13),(14),(15),(16),(17) define the basic parametersfor either the detailed design or selection of the fan for theengine cooling system, with the possibility of using thespecialized selection software provided by fan manufacturers(DOE-2 simulation program, ASHRAE simulation toolkit,commercial simulation programs such as Trace and HAP)[22, 23]. Finally, some authors have developed models forfan performance selection and assessment [24]. The selectionmust be made in such a way that the fan operates in the bestregion for the static pressure drop curve of the air inductionduct.

COOLANT PUMPIn most engine applications the coolant flow rate is obtainedby a centrifugal open impeller type pump driven by theengine with an appropriate pulley ratio. The pressure requiredfor the pump depends on the pressure drop characteristics ofthe water jacket as well as on the cooling system externalresistance. The values of the pressure developed by thecoolant pumps at the engine nominal speeds are in the rangeof 50 … 100 kPa. The pressure developed by the pump is

generally expressed in terms of head of water ( ).The power consumed by the pump, taken into account itsmechanical efficiency, (typical values ranging from 0,7to 0,9 [18]):

(18)

The pump handbook [25] provides estimates for hydraulicefficiency, volumetric efficiency, rotor friction power, andseal/bearing losses as a function of flow, specific speed, androtor diameter at the design point. With the design pointestablished, the pump specific speed, ns, is calculated as:

(19)

where is the impeller speed, and is the discharge head.Once the pump specific speed is known, the guidelines forcalculating the impeller geometry available in the specializedliterature can be used [26]. The total capacity of the enginecooling system (in liters) can now be correlated with theengine power by [20]:

APPLICATION OF THEMETHODOLOGYThis part of the work is devoted to the design of the coolingsystem for a locomotive engine. The engine requires twodifferent coolant loops for engine water jacket andaftercooler. Performance targets are being definedconsistently with the engine manufacturer, althoughinformation for the estimation of the size of the coolingsystem, except for the pump performance characteristics andtheir desirable operational point, has not been provided. Thesystem must meet thermal requirements for high temperatureoperation (ambient condition = 35 °C), under maximum loadat regulated speed: maximum allowable engine coolant outlettemperature = 85 °C, and maximum allowable aftercooler airoutlet temperature = 65 °C.

APPLICATION PARTICULARITIESThe locomotive application demands a cooling system designthat includes the “normal” engine cooling components forcoolant and air in two separated circuits for engine andaftercooler. The engine circuit can consist of the engine waterjacket, the pump, the oil cooler, the compressor, theexpansion tank, and the radiators. Except for the compressorand the oil cooler, the aftercooler circuit could be analogousto the engine one. The project is conducted under severalconstraints: the engine cooling system components mustaccommodate to a given space and dimensions, performancetargets must be consistent with the locomotive application,and the time for the development is short.

In accordance to classical locomotive practices, thearrangement of the locomotive traction wagon is as follows: aDiesel engine with lateral radiator banks, where the air flowis driven through lateral grills in the intake towards theradiators, sucked by two centrifugal roof-mounted fans.


Tubes and fittings connecting different components can belocated in the middle space between radiators and the engine.

Table 1. Initial parameters in the preliminary design

PRELIMINARY DESIGN - As was described in theprecedent paragraphs, the design process of the enginecooling system starts with the determination of the nominalparameters of the core system components, and the definitionof a preliminary layout of the whole cooling systemarrangement, with the estimated dimensions for pipes andfittings accordingly with the topology of the system. Withthis in mind, in this paragraph the preliminary calculation andselection of radiators and fans are described, starting by theassessment of the heat rejected to the coolant.

Table 2. Basic parameters for radiator and fan sizingand selection

Due to the absence of detailed engine information allowingusing a thermal model, Lahvic's correlation was used in orderto find the total heat rejected to the coolant. This heat was inturn split into a part rejected through the engine (includingthat part dissipated in the oil cooler), and the remaining partdissipated through the aftercooler. Then, making someassumptions described in table 1, some important data tocalculate and select the radiators and fans were found, aspresented in table 2. The value of the heat flow to beexchanged in the radiator was used as an input to all theprograms used.

Cooling system packagingBased on the data obtained after the preliminary design, andtaking into account the locomotive manufacturer'srecommendation, various core sizes and configurations wereanalyzed for the engine and after-cooler radiators. Theirthermal and hydraulic characteristics are presented in figures6 and 7. The performance characteristics of the selected fansare defined by the head rise versus volumetric flow rateperformance curve drawn in figure 8.

Figure 6. Heat transfer performance surfaces: a) engineradiator; b) after-cooler radiator.

Figure 7. Radiator pressure loss characteristics: a)coolant side; b) air side.

The conceived layout of the coolant part of the system isillustrated in figure 9. The coolant flows through the enginewater jacket, the radiators and the oil cooler loop in parallelwith the compressor loop. In the second after-cooler circuit,the coolant is driven by a separate pump, in a way analogousto that in the engine circuit. The two circuits can beconnected via a link valve in case of necessity. Space andgeometry allowed for the cooling system determine thecoolant piping packaging. Any deficiencies in piping and


their resulting coolant flow restrictions would negativelyimpact engine performance and durability.

Figure 8. Fan performance curve.

Figure 9. Proposed engine coolant system layout.

MODELLING AND SIMULATIONAs the coolant pumps are part of the engine module, the maincomponent selection was reduced only to the radiators andfans. For the first ones, the radiator calculation program wasused, considering the data of the preliminary design. Makinguse of commercial literature two centrifugal (54 inchesdiameter, 11 blades), fans were selected. CFD and 1Dthermo-hydraulic modeling were then carried out usingcommercial programs PowerFlow and FlowMaster,respectively.

Figure 10 represents a cutaway view of the ventilationpassage to be modeled with the CFD program. By simulatingthe forced air flow through lateral grills and radiators, under apressure difference created by the fans installed on the roof,the detailed air pressure drops along the ventilation path, forthe volume air flows through upper (engine) and lower(aftercooler) paths were obtained, as presented in table 3. Themodeled pressure drops along the duct were used to build thehydraulic impedance characteristics, required by the 1Dmodel.

Figure 10. View of the ventilation duct.

Table 3. Pressure drops in the ventilation path for avolume air flow.

Although it is possible, in principle, to incorporate thedetailed CFD model of the ventilation part into the 1Dthermo-hydraulic model, a much simpler model has beenbuilt, with the radiator cores modeled as porous media. Thecomputed air velocity and pressure drop distributions acrosstheir surfaces are shown in figures 11 and 12, respectively.

<figure 11 here>

<figure 12 here>


The pressure and velocity plots of figures 11 and 12 showthat the air flow behavior is better across engine radiatorsthan across the aftercooler ones. Inflow disturbances, bothsteady distortions and turbulence, in particular, can increasethe pressure losses and air velocity misdistribution. The airflow distribution across the radiator frontal area is acceptable.

Now, the thermal and hydraulic information gathered duringthe preliminary design, the calculation of radiators, and theCFD simulation results, are used to build the model of thecomplete engine cooling system.

The model developed is a collection of the different enginecooling system components that make up both the coolingand the air loops. The entire air duct has been represented bythe fans and the radiators with the required aerodynamicimpedances defined by the total pressure loss vs. air flowventilation duct characteristics lumped in the radiator sub-models. With such a network it is possible to predict thetemperature levels reached by the coolant and the air flows(thermal power dissipation capacity).

After providing the initialization data required for all thecomponents and fluids, the response of the system under fullload conditions and extreme ambient temperature wassimulated. The heat dissipated by the engine, aftercooler, oil

cooler and compressor were input as a function of time asshown in figure 13. The coolant and air flows through theradiators are represented in figure 14.

Figure 13. Input of heat flow rate through engine waterjacket, after-cooler jacket, oil cooler and compressor.

Figure 11. Air velocity field on the radiators frontal (left) and rear (right) sides.

Figure 12. Air pressure distribution before (left) and after radiators (right).


Figure 14. Coolant and air flow rates trough engine andafter-cooler radiators.

The coolant and air temperatures through engine andaftercooler radiators are plotted in figures 15 and 16,respectively. The temperatures are expressed in adimensionless form, as the ratio of the respective temperatureto the difference between inlet coolant and air temperatures.Under extreme conditions, the coolant and air temperaturesare below the threshold values defined in the designconditions, so the system designed meets the specifiedrequirements. Although the simulation model allowedassessing the behavior of different thermal and hydraulicquantities, only the thermal and flow rate responses related toradiators fall within the scope of this work. With the modeldeveloped different thermal operating and ambient conditionscan be simulated.

Figure 15. Inlet and outlet coolant and air temperaturesthrough one of the engine radiators.

Figure 16. Inlet and outlet coolant and air temperaturesthrough one of the after-cooler radiators.

EXPERIMENTAL RESULTSAfter assembling the cooling system in accordance with thecomponents selected, and the general arrangement used in themodelling process, the traction wagon was placed in thetesting room, were the load was created by a controlledelectrical load box. The coolant and air parts wereinstrumented to measure fluid pressures, temperatures andflows at the locations of interest in order to estimate thepressure losses, the heat flow rates, and the general energybalance.

Even though the cooling system is being designed forextreme conditions, the field test had to accommodate to theactual ambient conditions, with the normal ambienttemperature of 14 - 15 °C. Table 3 shows the dimensionless


temperatures in percentage (defined as the ratio of thedifference between inlet coolant temperature and airtemperature at the considered zone to the difference betweeninlet coolant temperature and inlet air temperature) at eachradiator surface. There is an appreciable misdistribution ofthe air velocity across the radiator face. The main problemappears in the lower region, where the velocity profile is verycomplicated, as was evidenced from the CFD simulation; it isdifficult to eliminate the air side misdistribution along theradiators as the flow is very non-uniform and the air duct hasinsufficient room to smooth out the flow.

Table 3. Dimensionless temperatures at each radiatorsurface

By using the measured coolant flow rates and temperaturesthrough radiators, compressor and oil cooler, the heat flowrates in these heat exchangers were calculated and are shownin table 4.

Table 4. Heat flow rates measured in the system heatexchangers, as percentage of the total heat exchanged

Calculating the air flow rates required to close the heatenergy balance in the radiators, considering the inlet andoutlet air temperatures, a great discrepancy was observed inthese magnitudes in relation to the measured values, as isnoticed in table 5.

Table 5. Heat flow rates measured in the system heatexchangers, as percentage of the total heat exchanged

In comparing the experimental and CFD results (theseobtained after running the model with the revised conditions,the real field test ambient conditions), the mean air flow ratesobtained from the CFD modeling are closer to the valuesobtained with the heat energy balance (the heat dissipated bythe coolant). The discrepancies in measured air flows can beassociated with the location of the measurement probe andthe stability of the measurement, which suggestsreconsidering the measuring method, Comparisons between

the experimental and predicted parameters for the engine andaftercooler radiators are presented in table 6.

From the model and experimental results, it follows that thewater jacket and aftercooler radiators are well dimensionedfor the actual thermal demand under full load conditions. Theactual thermal load dissipated by the four radiators amountsto a magnitude very close to the thermal load calculatedduring the preliminary design. If this value of heat rejected isextrapolated to the conditions prescribed during the radiatordesign, the installed dissipation power of the radiators isabout 11 % higher than the actual heat dissipation demand,which means that there is certain safety margin for heatdissipation capacity.

Table 6. Percentage differences between experiments andpredictions for the two radiators.

In general, the model predictions and experimental values arein good agreement, the major differences being observed inthe air flow circuits. This exercise has evidenced thestraightforwardness of the methodology followed in thecooling design process presented in this work.

CONCLUSIONSA methodology has been developed for the design of enginecooling systems, which enables information from threeprograms to be combined in order to get a fast design ofcooling systems: a home-made program for sizing and ratingof heat exchangers, a 3D CFD program (PowerFlow) used forair flow modeling, and a 1D thermo-hydraulic program(FlowMaster) for the global modeling of the cooling system.Previous to the methodology, a classical preliminary designprocedure is used to start the design of the engine coolingsystem, provided that the engine heat rejection to the coolantis known.

The methodology has been applied to the design of thecooling system for a locomotive Diesel engine, based onlocomotive manufacturer specifications, as well as onphysical characteristics of the system and on informationfrom the literature. The comparison of experimental andmodeled results confirmed the suitability of the methodology.


In particular, the performance of the cooling system designobtained, as far as the heat rejection under full loadconditions is concerned, showed that the preliminaryassumptions were accurate and that the models developedreproduced well the real system behavior.

The air flows rates obtained with the energy balance frommeasurements on the coolant side of the system were used tovalidate the air flows predicted with the CFD model, giventhe lack of consistency in the experimental measurements ofthe air flow across the ventilation passages. The measurementof the air flow distribution and the actual pressure losses hasbeen especially difficult, one reason being that the ventilationduct has non-regular forms and shapes, which give rise tocomplex turbulent flow phenomena, as has been noticedduring the 3D CFD simulation of the ventilation part of thesystem. Thus, an important step in the modeling of thecooling system has been the application of numerical tools tothe cooling air flow.

Finally, this work has shown the usefulness of the developedmethodology for single application engine cooling systemdesigns.

For confidentiality reasons most of the results have beenpresented in non dimensional form.

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http://www.sae.org/technical/papers/2001-01-1732

http://www.sae.org/technical/papers/942270

http://www.sae.org/technical/papers/960073

Prediction of Water Pumps,” SAE Technical Paper2001-01-1715, 2001.

CONTACT INFORMATIONPablo César Olmeda Gonzá[email protected]

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A Methodology for the Design of Engine Cooling Systems in Standalone Applications

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