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Applied Thermal Engineering 31 (2011) 2474e2484

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Applied Thermal Engineering

journal homepage: www.elsevier .com/locate/apthermeng

Thermo-hydraulic characterization of an automotive intercooler for a lowpressure EGR application

Cristian Cuevas a,*, Danielle Makaire b,1, Philippe Ngendakumana b,2

aDepartamento de Ingeniería Mecánica, Facultad de Ingeniería, Universidad de Concepción, Casilla 160-C, Concepción, Chileb Thermodynamics Laboratory, University of Liège, Campus du Sart Tilman e Bâtiment B49, Parking P33, B-4000 Liège, Belgium

a r t i c l e i n f o

Article history:Received 22 January 2010Accepted 10 April 2011Available online 19 April 2011

Keywords:Automotive intercoolerCharge air coolerLow pressure EGRNOx emissionsAir-cooled heat exchanger

* Corresponding author. Tel.: þ56 41 220 3550; faxE-mail addresses: crcuevas@udec.cl (C. Cuevas)

(D. Makaire), pngendakumana@ulg.ac.be (P. Ngendak1 Tel.: þ32 4 366 4821; fax: þ32 4 366 4812.2 Tel.: þ32 4 366 4803; fax: þ32 4 366 4812.

1359-4311/$ e see front matter � 2011 Elsevier Ltd.doi:10.1016/j.applthermaleng.2011.04.013

a b s t r a c t

In this work an experimental study is carried out to determine the thermo-hydraulic performance of anintercooler (IC) with flat tubes provided with triangular plain internal fins and louvered external finswhen it is used on a car equipped with a low pressure EGR. The main unknowns to be answered are thethermo-hydraulic characteristics of the IC working under humid conditions induced by EGR, theconditions under which the water content in the mixture of air and exhaust gases begins to condenseand the conditions under which the condensed water will be retained inside the IC. The exhaust gasesare here replaced by a mixture of dry air and water vapour which are mixed upstream of the IC. The IC issubmitted at the following testing conditions: on the ambient air side, the air temperature is fixed ataround 20 �C and the air velocity is settled at 1, 2 and 4 m s�1; on the exhaust gases side, the supplytemperature was varied between 90 and 150 �C, with dry gas flow rates of 20, 50 and 100 g s�1, and watercontents varying between almost 0 and 0.09 kgw kgg�1. At these conditions the IC transfers between1 and 13 kW with overall heat transfer conductance varying between 0.05 and 0.38 kWK�1 and effec-tivenesses between 0.3 and 0.97. A water accumulation inside the IC was detected for the tests carriedout at low dry gas flow rate of 20 g s�1 and for all the water content explored here (higher than0.02 kgw kgg

�1).� 2011 Elsevier Ltd. All rights reserved.

1. Introduction

Engine pollutants emissions have been of high concern for thepast decades so that automotive manufacturers had to developsystems to meet increasingly strict emission regulation. With theaim of meeting NOx emissions standards, manufacturers havestudied the possibilities of re-circulating a part of exhaust gases inthe combustion air, which is called Exhaust Gas Recirculation(EGR). In fact, this recirculation lowers the in-cylinder combustiontemperature and thus, reduces thermal NOx formation. Differentarchitectures have been proposed for EGR and the exhaust gasescan be re-circulated with or without cooling at low or high pressure[1e3].

In this study, a cooled EGR located in the low pressure line aftera depollution system is considered (EGRlp in Fig. 1). In this configu-ration the exhaust gases are taken after the turbocharger turbine

: þ56 41 225 1142., danielle.makaire@ulg.ac.beumana).

All rights reserved.

through an EGR cooler and injected upstream of the compressor. Asoot cleaning system is usually installed prior to the cooler asresearchers have shown that EGR coolers can be fouled because ofthermophoretic particle deposition and condensation of acids andheavier hydrocarbons [4,5]. After the turbocharger compressor, thecharge is cooled before the engine supply bymeans of an intercooler(IC). The water vapour that is present in the ambient air and in theexhaust gases can condense in the intercooler as its surfacetemperature drops below the dew point of the water vapour in thecharge. The resulting condensate can accumulate onto the surfaceuntil it is removedbygravitational orflowforces. It canaffect theheattransfer and pressure drop performance. Thus retained condensatescause degradation of the heat exchanger overall thermalehydraulicperformance. One has to note that acidic condensation can alsoappear in the intercooler like in theEGR cooler as the charge containscombustion gases. In fact, different chemicals can contribute to theacidity of the condensates from combustion gases. Combustion offuels containing sulphur leads to the formation of sulphuric acid [6].Nitric acid canalso formdue to thepresenceofNOxcompounds in theexhaust andfinally, organic acids can be formed due to the unburnedhydrocarbons in the exhaust. Moroz et al. [3] have tested highpressure and lowpressure EGR circuitswith low sulphur fuelswhere

Nomenclature

UA overall heat transfer conductance, WK�1

C coefficient of discharge or air velocity, m s�1

c specific heat, J kg�1 K�1

d diameter of orifice, mE velocity of approach factor_H enthalpy flow, WI current, AK factorM mass, kg_M mass flow rate, kg s�1

n number of testsP pressure, barR resistance, U_R residuals, WRe Reynolds numbert temperature, �C_U internal energy variation, WV voltage, VZ factor

Subscriptsa ambient airex exhaustg exhaust gasesIC intercoolerinj injectedmeas measurednoz nozzleop orifice plateref referencesep water separatorsu supplyv vapourw waterwb wet bulb

Greek symbolsb diameter ratioe effectiveness or expansibility factor or errork isentropic exponentr density, kgm3

s standard deviations pressure ratio or time, s

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e2484 2475

the charge was cooled by a high efficiency water-cooled charge aircooler. They collected the condensates from the charge cooler andanalyzed them. It was found that all the condensates had pH above 3and almost above 3.5. By regarding the acidic dew point and waterdew point temperatures, the first one is always higher than thesecond one, so acidic condensationwill appear always first. Thus, themore unfavourable case is to consider thewater condensation, whenboth acidic and water condensation appear. Despite the possibleacidic condensation in the intercooler, only humid condensation isconsidered in this work.

Tests were performed in dry and humid conditions on anintercooler, with flat tubes provided with triangular plain internalfins and louvered external fins. The heat exchanger was placed ina wind tunnel where combustion exhaust gases were replaced by

Internal Combustion

Engine

IC

EGRlp

EGRhp

Turbocharger

Fig. 1. EGR configuration.

a mixture of dry compressed air and steam. There are two objec-tives to be covered in this paper: the first one concerns theconditions under which the flue-gases water content begins tocondense in the IC and the second one concerns the conditionsunder which the condensed water will be retained inside the IC.This experimental characterization of the heat exchanger is the firststep before developing and validating, in a further study, a predic-tive model of the intercooler performance.

The structure of the paper is as follows. First a literature reviewis performed on air-cooled heat exchangers, in particular on heatexchangers with triangular fins enhancement systems inside thetubes. Then, we thoroughly describe the experimental apparatusand test method. The calculation method used for the analysis isalso described. Finally the results and conclusions are presented.

2. Relevant studies

Fin and tube heat exchangers are commonly used in the auto-motive industry where compactness is desired (intercoolers, radi-ators, air-conditioning condensers and evaporators). In air-cooledheat exchangers, extended surfaces are often used to improve thethermal performance of the heat exchanger by increasing thesurface and inducing turbulent mixing of the air. Moreover, inter-rupted surfaces as louvered fins are also used to break the thermalboundary layer and increase the thermal performance. Experi-ments for the overall air side heat transfer coefficient have beenwidely performed over the past years in wet and dry conditions[7e11]. However, the analysis is usually performed on the air side(external surface).

In charge air coolers, the compressed air flows inside the tubes,which also have extended surfaces in order to increase the tube-side thermal conductance. Webb et al. [12] give a review ofrecent technology concepts applied to air-cooled heat exchangers.Some inserts are generally placed into the tubes to enhance theheat transfer. Dewan et al. [13] review the passive heat transferaugmentation techniques. They report that twisted tapes, wirecoils, ribs, fins and dimples are the most common passive methods

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e24842476

to increase the heat transfer. However, they focus their review ontwisted tapes and wire coils. Sachdeva et al. [14] numerically studythe heat transfer enhancement in a plate fin heat exchangerproduced by triangular inserts with delta-wing vortex generatorson the surface of the triangle insert. The passage for the flow istriangular. It is shown that the vortex generators increase the heattransfer.

In many air cooling applications, it has been shown that vapourcondensates can accumulate on the external fins of fin and tubeheat exchangers. On plain fins, the retained condensate bridges thefin gaps and in louvered fins, it blocks the inter-louver gaps. Thusthe quantity and location of retained condensate has an impact onthe air-side heat transfer and pressure drop [15e18].

Moroz et al. [3] have measured water condensation in a water-cooled charged air cooler for different coolant temperatures anddifferent burning fuels. The geometry of the charge air cooler is,however, not given. The influence of the EGR rate and the air fuelequivalence ratio is shown. If the EGR rate is very high, therewill bemore condensate but, when more fresh air is introduced into thecharge, the mixture of air and EGR is more diluted and water partialpressure decreases and there will be less condensate.

This water accumulation has not been studied inside tubes withtriangular fins. Kundu et al. [19] present an analytical model thatanalyzes the performance of a triangular fin working under dehu-midifying conditions. The results are compared to previous finmodels but are not analyzed experimentally.

To conclude, very little experimental work has been done onplate and fin heat exchangers with flat tubes, which have trian-gular fin inserts working under dry or wet conditions inside thetubes. The aim of this work is to characterize the heat exchangerbefore developing and validating, in a further study, a predictivemodel.

3. Test bench description

The experimental set-up is composed of four circuits:compressed air and steam circuits, which, once they are mixed, gives

RAS

Temperatures No

Co

Water separator

Orifice

Compressed air circui

Steam circuit

Electric boiler

Electric boiler

Stea

m

boile

r

Nozzle

Flowv

Pressure control valve

Ambient air c

Load cell

Reservoir

IC

Fig. 2. Experime

place to the exhaust gases circuit that enters the IC, and the ambientair circuit, which is used to cool the IC. These circuits are condi-tioned independently to achieve the conditions required to char-acterize the IC. Fig. 2 presents a diagram of the installation.

3.1. Compressed air, steam and exhaust gases circuits

The compressed air circuit begins with a non-lubricated screwcompressor, which provides an air flow (at a temperature near tothe ambient one and around 40% of relative humidity) of 200 g/s ata pressure between 4 and 7 bar. Then, part of this flow is senttowards the IC and the exceeding flow is rejected to the environ-ment. In this circuit, a gate valve was installed to reduce the pres-sure to that required. The air flow derived to the IC is regulatedthanks to a gate valve installed at the circuit exhaust and it ismeasured by means of an orifice plate installed upstream of the IC.The temperature of the air is regulated by using two electric boilers.Near to the IC, a steam flow is injected to regulate the moistureconditions at the IC supply.

To set the compressed air temperature to the required one, twoelectric boilers were installed: one of 12 kW (not isolated) and oneof 3 kW (isolated). The first one is connected by sections of 3 kW toapproach the desired temperature and the second one is equippedwith a manually regulated voltage device to regulate more finely itspower between 0 and 3 kW and thus the air temperature.

The vapour is produced by a 27.5 kW electrical boiler. Thisvapour, initially in a saturated state, passes through a pressurecontrol valve and then through a 500 W electrical boiler where it isoverheated and dried before being injected into the air stream.After this boiler, a nozzle and a gate valve were installed tomeasureand to regulate respectively the steam flow.

To collect the eventual condensed water, two water separatorswere installed downstream of the IC. The collected water is thenbrought into a tank to determine by weighing the condensed massflow rate. The water flow is also accompanied by a little air leakageto make sure that all the water flow leaves the separators and thatthere is no water accumulation inside them.

zzles

Electricmotor

Centrifugalfan

Damper

oling coils

Air compressor

Flow control valve

t

Pressure control valve

Air leakage

Electric boiler

control alve

Flue gases circuit

ircuit

ntal set-up.

Fig. 3. Intercooler.

Table 1Heat exchanger characteristics.

Number of tubes 7Internal free-flow cross-sectional area per tube 497 mm2

Ratio internal free-flow to internal tube area 0.97Internal finned area 2.032 m2

Internal total area 2.817 m2

External frontal area 0.0702 m2

External frontal free-flow area 0.0316 m2

Ratio external free-flow to external frontal area 0.45External finned area 4.408 m2

External total area 5.228 m2

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e2484 2477

3.2. Ambient air circuit

The IC was placed in a closed air loop which includes a centrif-ugal fan, a damper, a series of cooling coils and four nozzles tomeasure the ambient air flow. The cooling coils are used to set thespecified temperature at the IC supply and the damper to set thespecified air flow rate.

3.3. Heat exchanger

The IC is a fin and flat tubes heat exchanger, provided with flatinternal fins and louvered external fins. The dimensions of the heatexchanger are 6560 mm� 107 mm� 80 mm (width, height,depth). Fig. 3 shows a general view of the heat exchanger and Fig. 4gives some idea about the internal and external fins. Moregeometrical details of the IC are summarized in Table 1. Concerningthe fin characteristics, these are given in Fig. 5 and Table 2 (for theexternal ones), and in Table 3 (for the internal ones).

3.4. Measuring system

3.4.1. Ambient air flowFour nozzles are available in the air stream: two of 190 mm and

two of 80 mm in diameter. For each test, the nozzle combination is

Fig. 4. Details of the inter

selected on the basis of the air flow to be measured and byfollowing the recommendation of the ASHRAE Standard 41.2 [20],which recommends a speed in the nozzle throat ranging between15 and 35 m/s.

3.4.2. Compressed air flowThe air flow was measured by means of an orifice plate. This is

determined by using the methodology and equations recom-mended by the Standard ISO 5167 [21e23]. The pipe diameter is51 mm and the diameters of the different orifice plates used hereare of 13, 22 and 32 mm. The orifice plate is combined witha differential pressure transmitter to determine the air flow.

3.4.3. Steam flowTo measure the steam flow, two different nozzles were used

according to the flow to be injected. The pipe diameter where thenozzle was installed has a diameter of 13 mm and the nozzles usedhave a nominal throat diameter of 1.5 and 3.5 mm, respectively. Themanufacture and the installation of these nozzles do not entirelysatisfy the Standard ISO-5167; for these reason these had to becalibrated. For this calibration, the exit of the regulation valve wasconnected to a plate heat exchanger where the vapour is cooled andcondensed by network water. The condensed vapour is then sent toa reservoir to determine the mass flow rate by weighing. Thecalibration of these nozzles is compared to the results given by theStandard ISO-5167. The difference is here attributed primarily tothe uncertainty on the nozzle diameters. According to this cali-bration, to fit the measured results with those given by the ISOStandard, a throat diameter of 1.605 mm and 3.57 mm respectivelyis necessary for the nozzles that are supposed to have a diameter

nal and external fins.

Fig. 5. Louvered fin and tube geometry.

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e24842478

(according to the tool used in the manufacturing of these nozzles)of 1.5 and 3.5 mm. After correction of the diameters, the steam flowrate is measured with an uncertainty lower than �1.5%.

3.4.4. Condensed flow rateThe condensed water is measured by weighing the water

leaving the water separators installed downstream the IC. Thismeasurement is carried out with a load cell which was calibratedprevious to the tests. This calibration allows to measure this flowwith an accuracy of �4%.

3.4.5. TemperatureTemperatures are measured with T-type thermocouples. The

reference temperature of 0 �C is given by awater/ice mixture. In thecase of the thermocouples installed inside the tubes, these areinstalled using glove fingers, as shown in Fig. 6. This configurationallows minimizing the errors due to the thermal conduction.

Concerning the location of the thermocouples on the exhaustgases side, thesewere installed far from the heat exchanger in orderto avoid errors due to the conduction in piping and to havehomogeneous conditions in the flow. The measured values werethen corrected to bring them back to the conditions at the heatexchanger supply and exhaust. Fig. 7 shows the location of thethermocouples in the flue gas circuit as well as the dimensions ofpiping to perform this correction.

24T-type thermocouples (12 upstream and 12 downstream)were installed on the heat exchanger tubes to measure the evolu-tions of the surface temperature throughout the heat exchangerduring the entire testing campaign (Fig. 8). The main purpose ofthese measurements is to identify the heat exchanger surfacetemperature profile.

On the ambient air side, 3T-type thermocouples were installedupstream and 6 downstream the IC to determine the heat flowtransferred to the ambient air. These measurements are shown inFig. 9.

3.4.6. Data acquisition systemThe signals provided by all of the sensors weremeasured using 4

data acquisition cards. This type of card can measure voltages andcurrents, in D.C. current, and temperatures.

Table 2External fin characteristics.

Tp: 14 mm Ft: n.a.Td: 80 mm Lh: n.a.Th: 7 mm Ll: 5 mmFp: 2.6 mm Lp: 2 mmFd: 80 mm q: n.a.Fh: 7 mm

All the transmitters used here, except the thermocouples, givea current signal as output, which is transformed to a voltage signalby using a 100 U (� 0.1%) resistance. The uncertainties introducedby the acquisition system are given in Table 4.

3.5. Uncertainty analysis

Before beginning with the experimental analysis, some issuesconcerning the measures, sensors and data acquisition systemuncertainties are analyzed.

According to the ANSI/ASHRAE Guideline 2 [24], instrumentalaccuracies are given for a confidence level of 95%. The methodproposed here is based on the same reference.

Measurement uncertainties can be considered as coming fromfour sources:

Measurement system: related to the sensor and data acquisitionsystem inaccuracies,Measures: related to the evolution of measurement, mainly tothe standard deviation and number of measures,Mounting: errors due to the installation, andPhysical: errors related to physical phenomena.

Thus, the measurement uncertainty is given by Eq. (1):

e ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffie2ms þ e2meas þ e2mount þ e2phys

q(1)

The error due to the measures is given by Eq. (2):

emeas ¼ Z$sffiffiffin

p (2)

where Z¼ 1.96 for a confidence level of 95%, s is the standarddeviation and n is the number of readings. In this study this term isneglected, because of the number of readings and the data acqui-sition rate (1 acquisition every 5 s) are very high, and the testingperiod (between 20 and 30 min) is considerably long. Moreover,average values are taken in steady-state regime, so the standarddeviation is also very low.

The error due to themounting can be also neglected here thanksto the care taken in installing the sensors. Concerning the physicalphenomena, these are also neglected due to the difficulty to eval-uate them theoretically. In short, in this study, only the errors

Table 3Internal fin characteristics.

Fd,i: 667 mm Ft,i: 0.07 mmFh,i: 6.4 mm Fp,i: 1.3 mm

Fig. 6. Glove fingers used to measure the temperatures.

230 mm25 mm 25 mm

230 mm

Fig. 8. Location of the surface thermocouples.

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e2484 2479

introduced by the sensor and the data acquisition system areconsidered.

3.5.1. Sensor and data acquisition system uncertaintiesPerformance specifications of sensors are: operating conditions,

output signals, and accuracy. Tables 5 and 6 give the manufactureuncertainties of the pressure and differential pressure devices.

For the accuracy, sensor manufacturers give the “referenceaccuracy”, which defines the limit that errors will not exceed whenthe device is used under reference operating conditions. It includesthe combined linearity, hysteresis and repeatability errors.

3.5.2. Overall uncertaintiesTo define the measurement uncertainty all the uncertainties on

the measurement chain must be taken into account, thus theoverall uncertainty is calculated by Eq. (3):

ems ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffie21 þ e22 þ e22 þ.þ e2n

q(3)

where n is the number of error sources.

tg,ex,IC

tg,su,IC

Tu

Ins

Tube: Insulati

Fig. 7. Location of the thermocoup

edas ¼ �½0:01%rdgþ 0:01%fs� ¼ �½0:01% 0:41 Vþ 0:01% 2:2 V� ¼ �edas ¼ �½0:01%rdgþ 0:01%fs� ¼ �½0:01% 2 Vþ 0:01% 2:2 V� ¼ �4:

In the case of temperatures two uncertainty sources areconsidered: one from the thermocouple tolerance and anotherfrom the data acquisition system. Thus the overall uncertainty isgiven by Eq. (4).

ems;T ¼ �ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi0:52 þ 0:32

p¼ �0:6 K (4)

No doubt this uncertainty is overestimated. In practice therelative uncertainty is lower than �0.3 K. But the value of �0.6 Kwill be retained for further analysis.

For the pressure sensor (and differential pressure sensor), thereare three error sources: the sensor itself, the 100 U resistance andthe data acquisition system.

The pressure is calculated using Eq. (5):

P ¼�

I � I0Imax � I0

�$Pfs ¼

�V=R� I0Imax � I0

�$Pfs ¼

�V=R� 0:0040:02� 0:004

�$Pfs

(5)

According to Table 4, the data acquisition system error is�[0.01%rdgþ 0.01%fs]. In this case, the error is calculated for twoextreme conditions, when the sensor is close to its minimalpressure and when it is at full scale, i.e. for Isensor¼ 4.1 mA andIsensor¼ 20 mA, which gives the following results:

be: Internal diameter = 50 mm. External diameter = 60 mm. Long = 560 mm.

ulation: Thickness: 20 mm. (cond. 0,045 W m K at 40°C)

Internal diameter = 50 mm. External diameter = 60 mm. Long = 1170 mm.

on: Thickness = 20 mm. (cond. 0,045 W m K at 40°C)

les in the exhaust gases side.

2:61� 10�4 V2� 10�4 V

(6)

130 mm

60 mm

15 mm210 mm

210 mm

~ 6500 mm

115 mm

30 mm

Fig. 9. Location of the air thermocouples on the ambient air side.

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e24842480

Thus, the uncertainty introduced by the data acquisition systemand by the resistance for a pressure transmitter of 10 bar is:

Low scale : edasþR ¼ �ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi�vPvR

eR

�2

þ�vPvV

eV

�2s

¼ �0:003 bar for a measured pressure

of 0:063 bar (7)

Full scale : edasþR ¼ �ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi�vPvR

eR

�2þ�vPvV

eV

�2s

¼ �0:013 bar for a measured pressure

of 10 bar (8)

For the low scale, 71% of the error comes from the resistanceuncertainty and 29% from the data acquisition system. For the fullscale 95.8% comes from the resistance uncertainty and 4.2% fromthe data acquisition system. According to this analysis, uncertaintyintroduced by the data acquisition system is very low compared tothe uncertainty introduced by the resistance. On the other hand,the uncertainty of the sensor itself varies depending on the sensortype. In this case, a pressure sensor with an uncertainty of 0.5% of

Table 4Uncertainties of the data acquisition system.

Measurement Range Uncertainties

Temperature �100 to 400 �C <0.3 KPressure 0e2 V �[0.01%rdgþ 0.01%fs]Differential pressure 0e2 V �[0.01%rdgþ 0.01%fs]

full scale is considered. Thus, the total uncertainty on the pressuremeasurement is given by Eqs. (9) and (10):

Low scale : ems;P ¼ �ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi0:0032 þ 0:052

p¼ �0:050 bar

¼ 0:50% of full scale (9)

Full scale : ems;P ¼ �ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi0:0132 þ 0:052

p¼ �0:052 bar

¼ 0:52% of full scale (10)

According to this analysis, the contribution of the data acquisi-tion system to the error on pressure transmitters, and also on thedifferential pressure transducers, is negligible due to the highprecision of the data acquisition cards. Table 7 summarizes themeasuring uncertainties involved directly on the thermo-hydrauliccharacterization of the IC. These uncertainties will be considered inthe experimental analysis.

3.6. Data processing

All the measured values were recorded every 5 s by using a dataacquisition system and the software RTM 3500. The observation ofthe temporal evolutions of the temperatures, pressures and

Table 5Characteristics of the pressure transmitters.

Pa,su,op Pg,su,IC and Pv,su,inj Pv,su,noz

Range 0e7 bar 0e4 bar 0e10 barOutput 4e20 mA 4e20 mA 4e20 mAUncertainty 0.15% reading 0.15% full scale 0.5% full scale

Table 6Characteristics of the differential pressure transmitters.

DPa,op DPa,nz DPv,noz DPa,IC DPg,IC

Range 0e150 mbar 0e12.5 mbar 0e150 mbar,0e1000 mbar

0e5 mbar 0e150 mbar

Output 4e20 mA 4e20 mA 4e20 mA 4e20 mA 4e20 mAUncertainty 0.5% full scale 0.5% full scale 0.5% full scale 0.5% full scale 0.5% full scale

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e2484 2481

differential pressures makes it possible to check the stability of thetest and to choose the steady period, over which the averages willcarry out to be then used in the experimental analysis.

4. Calculation

In this analysis the mass balance is very important to determinethe moisture content at the IC exhaust, as well as the energybalance to determine the heat flow. Hereafter the equations used todetermine them as well as the results are presented.

4.1. Mass balance

The most important mass balance is that related to the watercontent in the exhaust gases. By assuming steady-state conditions,the water mass balance is given by Eq. (11):

�_Mw;a þ _Mw;inj

�� �_Mw;cond þ _Mw;ex

� ¼ 0 (11)

The first two terms correspond to the water entering the IC:the water contained in the compressed air and the injected water.The first one is practically negligible, but it will be considered in theanalysis. The two second terms correspond to the water leaving theIC: the condensedwater and thewater that leaves the IC in a vapourstate. The first one is measured by weighing and the second one isthe unknown variable of this equation.

This mass balance allows to determine the water content at theIC exhaust for the test wherewater condensation occurs. In practicethis value is impossible to measure with traditional humiditytransducers, because these are not designed towork with saturatedair.

4.2. Energy balance

The energy balance on the IC is given by Eq. (12):

_Hg;IC � _Ha;IC ¼ _UIC (12)

where _Hg;IC is enthalpy flow supplied by the exhaust gases, _Ha;IC isthe enthalpy flow to the ambient air and _UIC is the internal energyvariation.

The enthalpy flows are calculated by Eqs. (13) and (14).

Table 7Testing ranges and measuring uncertainties.

Variable Measurement range Uncertainty

Temperatures 20e150 �C �0.6 KPressures Pg,su,IC 1.99e3.0 bar Lower than �0.32% reading

Pa,su,op 2.14e3.24 bar Lower than �0.25% readingFlows _Mw;sep 0.06e2.34 g/s Lower than �4% reading

_Mv 0.43e5.56 g/s Lower than �1.5% readingDifferential

pressuresDPa,nz 131e537 Pa Lower than �0.51% full scaleDPg,IC 204e2779 Pa Lower than �0.50% full scaleDPa,IC 71e523 Pa Lower than �0.52% full scaleDPg,op 5552e12493 Pa Lower than �0.51% full scale

_Hg;IC ¼ _Mg;IC$�hg;su;IC � hg;ex;IC

�� _Mw;sep$cw$ tw;ex;sep � tref

� �

(13)

_Ha;IC ¼ _Ma;IC$�ha;ex;IC � ha;su;IC

�(14)

The reference temperature tref is equal to 0 �C. The enthalpyflows on the exhaust gases side and on the ambient air side aremeasured; so that a third term is introduced: the residual of theenergy balance. The equation of the energy balance is finally givenby Eq. (15) with _RIC the residual of the energy balance. This residualgives an idea about the energy balance accuracy.

_Hg;IC � _Ha;IC � _UIC ¼ _RIC (15)

The internal energy variation is determined using Eq. (16) whereMIC is the heat exchangermass and cIC is the heat exchanger specificheat.

_UIC ¼ MIC$cIC$dTICds

(16)

The mass of the empty heat exchanger is 2.555 kg. In thiscalculation only this mass is considered with a specific heat(aluminium) of 0.88 kJ kg�1 K�1.

4.3. Thermalehydraulic performance

The IC is also characterized through its overall heat transferconductance. This is calculated by using the e-NTU method andconsidering a crossflow configuration. The exhaust gases flowcapacity used to determine it is defined by Eqs. (17) and (18).

Dry regime : _Hg;IC ¼ _Cg;IC;dry$�tg;su;IC � tg;ex;IC

�(17)

Wet regime : _Hg;IC ¼ _Cg;IC;wet$�twb;su;IC � twb;ex;IC

�(18)

Fig. 10. Comparison between the measured flow enthalpies on the air and exhaustgases side for the tests with vapour injection.

Fig. 11. Residuals of the energy balance for the tests with vapour injection.Fig. 13. Exhaust gas side supply temperature versus supply humidity ratio.

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e24842482

For the pressure drops, these are presented by defining the Kfactor as the ratio between the actual pressure drop and thedynamic pressure (Eq. (19)).

KIC ¼ DPIC;meas

rIC$C2IC2

(19)

5. Experimental results and analysis

5.1. Mass and energy balances

Concerning the testing conditions, on the air side all the testswere carried out at a supply ambient temperature of 20 �C and airvelocities of 1, 2 and 4 m s�1. On the compressed gases side, fourparameters were explored: supply gas temperature, between 90and 150 �C, supply humidity ratio, between 0 and 0.09 kgw kgg�1,supply pressure of 2 and 3 bar, and dry gas flow rate varyingbetween 20 and 100 g s�1.

As mentioned previously, the mass balance, given in Eq. (11),allows to determine the exhaust gases humidity. According to theresults, the exhaust gases leave the IC in a saturated state in allcases when there is water condensation inside the IC.

Concerning the enthalpy flows, a systematic difference isobserved in both non-vapour injected and vapour injected tests. Inorder to clarify this difference an uncertainty analysis is carried outto determine if this difference is due to the uncertainty propagation

Fig. 12. IC power versus the supply air velocity.

or to a systematic error. Figs. 10 and 11 show, respectively,a comparison between the flow enthalpies on the exhaust gas andair side, and the residuals of the energy balance for the tests withvapour injection with their uncertainty bars. According to theseresults, the difference is due to a systematic error of the order of5.6% for high heat transferred. As this uncertainty appears in thevapour injected as well as in the non-vapour injected tests, it is notdue to a systematic error on the vapour injection or condensedwater measurement. The systematic error is attributed to ambientair exhaust temperature, which is measured with six thermocou-ples. The temperature profile is not uniform due to the heatexchange in the IC and, eventually, due to a flow maldistributioninside and outside the IC. In addition, the air can pass through thenon-finned frontal surface located on the borders of the IC, whichcreates a bypass. All this can explain the difference (lower than4 �C) found for the measured systematic error. In further analysisthe corrected temperature is considered.

Fig. 12 shows the experimental results of the intercooler heattransfer and the influence of the gas mass flow rate, the supply fluegas temperature and the air velocity on the heat transfer, whichvaries between 1 and 13 kW for the conditions covered in thisstudy. This figure shows that, as expected, the heat transferincreases when the supply gas temperature and/or the gas massflow rate increase and also that the heat transfer increases whenthere is water condensation. The effect of the ambient air velocityand working pressure on the heat transfer could not be observeddue to the limited number of tests carried out and the rangecovered.

Fig. 14. Condensed water versus supply dew point temperature.

Fig. 15. IC overall heat transfer conductance. Fig. 17. Air side pressure drop.

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e2484 2483

The heat transfer increases almost linearly with the supplyhumidity ratio in the exhaust gases, for a given exhaust gas flowrate and exhaust gas supply temperature (Fig. 13). For humidityratios higher than 0.02 kgw kgg�1 condensation appears even atsupply gas temperatures of 150 �C and gas flow rate of 50 g s�1. Thecondensed water flow rate, shown in Fig. 14, increases with thesupply dew point temperature for a given gas flow rate and supplygas temperature. As the supply gas temperature decreases thecondensed flow rate increases for the same humidity ratio (dewpoint temperature) and the same exhaust gas flow rate.

The IC overall heat transfer conductance, shown in Fig. 15, variesbetween 0.05 and 0.38 kWK�1 and its corresponding effectivenessbetween 0.3 and 0.97. Almost a linear relationship is observed forthe heat transfer conductance against the gas flow rate. It isobserved that in wet regime (with vapour condensation) the globalheat transfer conductance is almost twice than the one obtained indry regime (without vapour condensation). The highest effective-ness is obtained at lower gas mass flow rate and in dry regime andthe lower effectiveness in wet regime and at high gas flow rates(Fig.16). Themost important here is to observe that most tests weredeveloped at effectiveness lower than 90%, which guaranties a gooddomain to determine the global heat transfer conductance.

Figs. 17 and 18 show the parameter K, defined in Eq. (19), for theexhaust gas side and the ambient side both plotted against theReynolds number, which is determined by using the averageconditions (temperature, pressure and humidity ratio) to deter-mine the fluid properties and two times the fin separation for thehydraulic diameter. A good correlation is obtained on the air side,which guaranties that measurements were carried out in a goodway. On the other hand, on the gas side a higher scattering is

Fig. 16. IC effectiveness versus NTU.

obtained. Here it is observed that the pressure drop rises due towater accumulation inside the IC, which reduces the tubes surfaceavailable to the gas flow.

5.2. Water accumulation inside the IC

A second study was carried out to determine the conditions atwhich there is water accumulation inside the IC. To perform that, 24thermocouples were installed on the tubes of the IC.

In five tests an accumulation of water was detected: tests 1, 2, 3,4 and 5, shown in Table 8. In these cases 3 tubes were inundated intests 1 and 4, and 4 in tests 2, 3 and 5. The term “inundated tubes”means that there is stagnated water that is accumulated in thebottom of the intercooler, blocking these tubes, and means that thecirculating gas flows through the unblocked tubes. This transientcondition must be taken into account because when the exhaustgas flow rate increases, these blocked tubes disturb the intercoolerperformance: increasing its gas side pressure drop and decreasingits heat transfer area. The water accumulation was detected for thetests at dry gas flow rate of 20 g s�1 for the water content exploredhere (higher than 0.02 kgw kgg�1). Others testing conditions aregiven in Table 8.

According to these results, in case of using this type of inter-cooler on a vehicle equipped with a low pressure EGR, liquidaccumulation could be avoided by choosing a low side exhaustcollector or by installing a purge at the bottom of the exhaustcollector. It is also mandatory to install a water separator at theintercooler exhaust. In case of acidic condensation care must bealso taken in the choice of the materials used for the intercooler,water separator and for the line connecting the intercooler and thecombustion engine.

Fig. 18. Exhaust gas side pressure drop.

Table 8Test conditions where there is water accumulation inside the IC.

Test tg,su,IC, �C tg,ex,IC, �C wsu,IC, kg kg�1 RHg,ex,IC Pg,su,IC, bar Mg,IC, kg s�1 Mw,sep, g s�1 Hg,IC, kW ta,su,IC, �C ta,ex,IC, �C Ma,IC, kg s�1

1 88.8 43.4 0.046 1.11 3.00 0.020 0.48 2.139 21.3 47.0 0.0822 89.1 34.0 0.024 1.17 2.97 0.020 0.20 1.630 19.9 38.3 0.0873 88.5 24.1 0.022 1.33 3.03 0.020 0.28 2.040 19.8 25.9 0.3294 89.5 27.5 0.045 1.21 3.01 0.020 0.70 3.068 20.7 29.9 0.3245 88.5 24.3 0.023 1.64 3.03 0.020 0.25 1.946 20.0 25.8 0.329

C. Cuevas et al. / Applied Thermal Engineering 31 (2011) 2474e24842484

6. Conclusions

An exhaustive testing methodology was applied to the deter-mination of the thermo-hydraulic performance of an automotiveintercooler. In the case of a vehicle equipped with a low pressureexhaust gas recirculation to reduce NOx emissions, a mixture offresh air and humid exhaust gases flows through the intercooler. Inthat case, water condensation can appear and the amount ofcondensed water will be influenced by the engine operatingconditions (EGR rate and air fuel equivalence ratio). A high EGR rateinvolves high moisture content in the intercooler supply andpromotes water condensation. However, a high air equivalenceratio can partially balance the EGR effect as it introduces more freshair: the water vapour concentration is decreased in the exhaustgases and the exhaust gases re-circulated are also more diluted bythe fresh air.

Thus, the intercooler had to be characterized under very wideconditions on the exhaust gas side. The exhaust gases were simu-lated by a mixture of dry air and water vapour.

The mass balance allowed to determine the water content at theIC exhaust. According to the results in all the tests where there waswater condensation, the exhaust gases leave the IC in saturatedconditions. The IC has been characterized through its overall heattransfer conductance which was determined from the e-NTUmethod. It varied between 0.05 and 0.38 kWK�1 with IC effec-tivenesses varying between 0.3 and 0.97. It was also noticed thatthe overall heat transfer conductance in wet regime was almosttwice than the one obtained in dry regime. The effect of theambient air velocity and working pressure on the heat transfercould not be observed due to the limited number of tests carriedout and the range covered.

The pressure drops have been evaluated by a parameter thatcompares the actual pressure drop and thedynamic pressure. On theexhaust gas side, the tests where there was condensate retentionhave beenwell identified through this parameter. This accumulationappeared at exhaust gas flow rates of 20 g s�1 for all the watercontents explored (higher than0.02 kgw kgg�1).Water retention intothe IC can degrade thermo-hydraulic performance by increasing itsexhaust gas side pressure drop and decreasing its heat transfer. Thisshould be taken into account when designing the system.

This experimental characterization performed on that inter-cooler is the first step before developing and validating, in a furtherstudy, a predictive model of the intercooler performance.

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