Experimental Thermal and Fluid Science 27 (2003) 781–787

www.elsevier.com/locate/etfs

New approach for performance testing of air-cooled condensing units

Cristian Cuevas 1, Eric L. Winandy *

Laboratory of Thermodynamics, University of Li�eege, Campus du Sart Tilman-Baat. B49, 4000 Li�eege, Belgium

Accepted 18 November 2002

Abstract

Condensing units are designed to work with the condenser-fan component working in open space. Testing condensing units in

these conditions presents several drawbacks: difficulty to find a really infinite plenum and to avoid air shortcuts around the con-

denser inlet which means temperature variations, impossibility to measure the mass flow rate passing through the condenser. On the

other hand, separating the condenser inlet and outlet plenums means a risk of not controlling the pressure difference between the two

plenums and then overfeeding or underfeeding the condenser.

This paper discusses the advantages and disadvantages to work in one way or in another. A method is presented to avoid

condenser over or underfeeding and guarantee a low air velocity profile.

Finally, an example of a flat-coil condensing unit tested with two different fans, variable refrigerant charge and evaporating

temperature is shown to illustrate the discussion.

� 2003 Elsevier Inc. All rights reserved.

Keywords: Flat-coil condensing units; Condenser; Compressor; Experimental procedure

1. Introduction

Condensing units are designed to work outside and

normally in an open space (Fig. 1a). In practice, this is

not always the case since obstacles may be present closeto the air-cooled condenser and significantly alter the

flow. Nevertheless, laboratory testing requires special

care for the air flow going through the condenser.

ASHRAE standards [1] is not very precise about this

matter as long as ‘‘no abnormal air flow is created over

the condensing unit’’. European Standards [2] is even

less precise except ‘‘the given performance must be ap-

plied to a clean condenser’’.

1.1. Open plenum test

Common practice is to put the condensing unit in a

room where the ambient air is maintained to the set

point temperature by an air-conditioning system (Fig.

1b). But this practice may let some doubt on the real air

*Corresponding author. Tel.: +32-4-366-4800; fax: +32-4-366-4812.

E-mail addresses: crcuevas@udec.cl (C. Cuevas), eric.winandy@

planetmail.com (E.L. Winandy).1 Tel.: +32-4-366-4825.

0894-1777/03/$ - see front matter � 2003 Elsevier Inc. All rights reserved.

doi:10.1016/S0894-1777(02)00315-1

flow to the condenser depending on the external venti-

lation system, position of ducts.

If ventilation rates are too strong, there may be some

kinds of jets on the condenser that would overfeed it

while if ventilation rates are too low, air will shortcutthrough the condenser which will increase temperature

variations and the condensing temperature level. As an

example, Fig. 2 shows temperature records 50 mm in

front of the condenser on the air-supply side in open

plenum mode for a 13 kW cooling capacity condensing

unit. It can be seen that when the external ventilation is

too low, air shortcuts outside the condenser can induce

temperature variations up to 2 K.Another drawback of this practice is that there is no

possibility to make any balance on the air-side of the

condenser.

1.2. Separate plenums method

The second method consists in separating the con-

densing unit inlet and outlet. This method has many

advantages, since a thermal balance on the air can be

performed beside the one on the refrigerant side. A

way to do it is to install the flat coil condenser in

some kind of duct and to control the flow going

Nomenclature

h specific enthalpy, J kg�1

_HH enthalpy flow, W

LMTD logarithmic mean temperature difference, K

M mass, kg_MM mass flow rate, kg s�1

N revolution speed, Hz

Dp pressure drop, Pa_QQ heat flow, W

SCC specific condenser capacity, WK�1

t temperature, �CUA global heat transfer coefficient, WK�1

v specific volume, m3 kg�1

V velocity, m/s

V volume, m3

_VV Volume flow rate, m3 s�1

_WW electric power, W

Greek symbols

D difference

e efficiency

Subscripts

a air

amb ambient

cd condensercor coriolis

cp compressor

eh electric heater

ev evaporator

ex exhaust

f saturated liquid

fan fan

g saturated gasis isentropic

loop water or air loop

nom nominal

pump pump

r refrigerant

R404a refrigerant R404a

room room

s isentropicsc sub-cooling

sh shaft

su supplied

sw swept

v volumetric

w water

Fig. 1. Air-cooled condenser configurations for testing.

782 C. Cuevas, E.L. Winandy / Experimental Thermal and Fluid Science 27 (2003) 781–787

through the condenser (Fig. 1c) as recommended for

an air-cooled condenser specific test [3,4]. The disad-

vantage of this method is that the air flow is far from

the ‘‘in-field’’ air mass flow since it goes straightly

through the condenser. Another disadvantage is that

the duct geometry has to be adapted each time a newcondensing unit is tested.

The method proposed here is presented in Fig. 1d

where two plenums are separated by a wall where the

condensing unit is placed. Low velocities in front of the

condenser guarantee to reproduce field conditions as

close as possible.

This method presents the disadvantage that air

plenum pressure before and after the condensing unit

must be controlled to be the same in order to main-

tain the same conditions as in a real situation. If this

condition is respected, the fan will work at nominal

conditions. Otherwise, overfeed or underfeed of thecondenser is possible if no precaution is taken. The

precautions to take care are given in the following

part.

2. Description of the test apparatus

2.1. Air network

The air flows through a closed loop driven by a fan

placed in the air channel outside the room. The air flows

from the hot side of the climatic room into the air

channel through cooling coils supplied with tap-water as

coolant. The cold side temperature is kept constant by

tuning the cooling water flow rate thanks to a regulating

valve driven by a PID controller. The controller tem-perature sensor is placed in the cold side of the room in

front of the supply air grill. This air grill provides a

uniform air velocity distribution at condenser inlet. The

air network is shown in Fig. 3.

Fig. 3. Air circulation in the test bench.

Cold side Hot side

cdap ,

roomap ,

Slope of plastic tape

Fig. 4. Pressure difference measurements in the climatic room.

Fig. 2. Inlet air temperature to the condensing unit.

∆

Fig. 5. Pressure difference and plastic-tape angle.

C. Cuevas, E.L. Winandy / Experimental Thermal and Fluid Science 27 (2003) 781–787 783

Air mass flow rate in the loop was adjusted by means

of a simple visual pressure-difference measuring element.

This measuring device consists in a strip or plastic tapeinstalled in a vertical hole in the wall separating the two

sides of the room. When the strip stays vertical, it means

there is no over-pressure in one of the two rooms.

Tests were performed to determine the sensitivity of

this simple visual element, which is used during the tests.

Thanks to a variable-speed fan, it was possible to adjust

precisely air mass flow rate. During the test, the con-

denser fan was switched on and the air mass flow ratethrough the room was tuned to generate positive and

negative pressure differences between the two sides of

the room. The pressure-differences are measured using

inclined-tube manometers with an accuracy of �1 Pa,

and the slope of the plastic-tape is estimated by ocular

inspection. Two pressure differences are measured: be-

tween the hot and the cold side of the room and between

the cold side of the room and the inside of the condensercasing (just after the coils). Details are shown in Fig. 4.

Fig. 5 shows the pressure difference between the two

rooms as a function of the slope. In the angle sensitivity

range of the hanging plastic tape, the pressure variation

is maximum �2 Pa. In this range the pressure drop

change on the condenser coils is about 2 Pa, as shown in

Fig. 5.

Assuming air mass flow rate through condenser is

proportional to the square root of pressure drop in the

condenser coils, then the ratio between actual and

784 C. Cuevas, E.L. Winandy / Experimental Thermal and Fluid Science 27 (2003) 781–787

nominal mass flow rate can be computed. This ratio is

shown on Fig. 5 as a function of strip slope. It can be

seen that air mass flow rate does not vary more than 1%

in the range )90� to 90� of slope. In conclusion, this tool

is enough to regulate air flow through the loop.

2.2. Water network

The evaporator is connected to a closed water circuit,

which includes an electric water heater outside the room

with a maximum heating capacity of 15 kW, a pump and

plastic connection tubes. The water loop is thermally

insulated. Electrical power can be adjusted in all therange.

2.3. The refrigerant network

The refrigerant network is shown in Fig. 6. It consists

in the condensing unit, a coaxial water heated evapo-

rator, a Coriolis mass flow meter, a thermostatic ex-

pansion valve and a sight glass. Refrigerant used isR404a.

The condensing unit is composed of an air-cooled

condenser, a scroll compressor and a refrigerant vessel.

It was tested with two different fans with different ro-

tation speeds. The compressor and condensing units

characteristics are detailed in Table 1.

Fig. 6. Test bench con

2.4. The measuring system

Temperatures are measured with type T thermocou-

ples (copper–copper–nickel) with an accuracy of 0.3 K.

Refrigerant mass flow rate is measured with a Coriolisflow meter with an accuracy of ð�0:25% of readingÞ þð�0:025% of _MMnomÞ, where _MMnom ¼ 0:27 kg/s. The flow

meter was installed in the liquid zone, so it was necessary

to ensure a good sub-cooling in order to avoid erroneous

measurement [6].

Pressures are measured with an accuracy of 0.2%.

Temperatures, pressures, mass flow rate and powers are

measured and recorded by a data acquisition systemwith a time period of 5 s.

3. Test analysis

3.1. Evaporator

The cooling capacity is obtained from the followingevaporator heat balance on water side:

_HHev ¼ _QQeh þ _WWpump þ _QQamb;ev ð1Þwhere _HHev is the cooling capacity (refrigerant enthalpy

flow rate extracted from the evaporator), _QQeh is the

electric power given to water by means of electric re-

nection diagram.

Table 1

Condensing unit characteristics

Compressor type Scroll

Compressor swept volume (m3/h) 14.5

Compressor RPM 2900

Nominal capacity (kW)� 13.09

Nominal power input (kW)� 3.95

Fans 1 & 2 diameter (cm) 42

Fans 1 & 2 N� vanes 4

Fan 1 RPM 1380

Fan 2 RPM 940

Fan 1 power (W) 240

Fan 2 power (W) 105

*Nominal conditions: tev ¼ 0 �C, tcd ¼ 40 �C and no sub-cooling.

C. Cuevas, E.L. Winandy / Experimental Thermal and Fluid Science 27 (2003) 781–787 785

sistances, _WWpump is the pumping power given to the

evaporator water network (93 W) and _QQamb;ev is the heat

flow rate transmitted from ambient to the evaporator.

This last term is given by

_QQamb;ev ¼ UAw;loop tamb;ev

�� ðtw;su;eh � tw;ex;ehÞ

2

�ð2Þ

A calibration of the water loop permitted to identify

heat transfer coefficient of the loop UAw;loop. It was

found:

UAw;loop ¼ 6 W=K ð3ÞThe refrigerant mass flow rate can be computed from:

_MMr ¼_HHev

ðhex;ev � hsu;evÞð4Þ

It is compared to the Coriolis mass flow rate for secu-

rity.

3.2. Compressor

The compressor balance in steady state is given by

_WWcp þ _HHcp þ _QQamb;cp ¼ 0 ð5ÞThe ambient losses of the compressor are determined

from this heat balance. It is necessary to evaluate them

since they are injected in the hot plenum and are part of

the condenser heat balance (Eq. (10)).The lubricant circulation is neglected here. Indeed,

when working with scroll compressors, oil concentration

is expected to be less than 1%.

In order to analyze compressor performances at dif-

ferent operating conditions, the volumetric and com-

pressor isentropic efficiencies are computed:

Volumetric efficiency is given by

em ¼_VVsu

NVswð6Þ

where Vsw is the swept volume and _VVsu is the volume flow

in the suction inlet:

_VVsu ¼ _MMmsu ð7Þ

Compressor global isentropic efficiency is given by

ecp;s ¼ws

wsh

ð8Þ

where ws is the isentropic work of compression and wsh

is the specific shaft work:

wsh ¼_WWsh

_MMð9Þ

3.3. Condenser

Calorimeter air balance allows us to calculate the air

flow rate passing through the condenser. Indeed we

have:

_MMaðha;su � ha;exÞ þ _HHr;cd þ _WWfan þ _QQamb;cp ¼ 0 ð10Þwhere ha;su is the air enthalpy in the cold room and ta;ex isthe air enthalpy in the hot room. _HHcd is the condenser

capacity (refrigerant enthalpy flow rate extracted from

the condenser) and _WWfan is the fan power given to the

ambient.

Treating the condenser as semi-isothermal heat ex-

changer (where the refrigerant has an infinite capacityflow rate), we can define an overall heat transfer coeffi-

cient UAcd by

_QQcd ¼ UAcd LMTDcd ð11Þwhere LMTDcd is the log-mean temperature difference

computed as [7]:

LMTDcd ¼ðta;ex;cd � ta;su;cdÞln

tcd�ta;su;cdtcd�ta;ex;cd

� � ð12Þ

where tcd is the condensing temperature computed by

tcd ¼tf;cd þ tg;cd

2ð13Þ

and ta;ex;cd is a corrected air temperature, which is re-

calculated by thermal balance on the condenser.

Since the condenser is also characterized by its

specific condenser capacity (SCC), this value is also

interesting for the manufacturer, even if the discussion

here will be based on the UAcd values. The SCC is

defined by

SCC ¼_QQcd

ðtg;cd � ta;su;cdÞsð14Þ

The refrigerant properties and the equations are solved

using EES [5].

4. Case study: condensing unit performance with two

different fans

In order to illustrate the method, the results obtained

on a condensing unit working with two different fans is

786 C. Cuevas, E.L. Winandy / Experimental Thermal and Fluid Science 27 (2003) 781–787

analyzed in this part. Furthermore, as a sensibility

analysis, the condensing unit was running at two dif-

ferent evaporating temperatures and for each one, the

refrigerant charge was varied in order to change the

condenser sub-cooling.The general results are presented in Table 2 and the

condenser specific results in Table 3.

4.1. Compressor

Fig. 7 shows the variation of the isentropic and vol-

umetric efficiencies. As it can be seen, the volumetric

efficiency decreases as the pressure ratio increases. It hasbeen shown that this trend can be explained by internal

heat transfers: the scroll chamber inlet density decreases

as the pressure ratio increases since the suction heating

up increases [9].

The isentropic efficiency presents the same variation

with the pressure ratio but the reason is quite different.

It is well known that, as soon as the external pressure

ratio becomes higher than the internal built-in one, theinternal efficiency start decreasing, as the pressure ratio

increases [8]. This effect is visible at the level of the

overall efficiency.

Table 2

General results

Test MR404A (kg) tev (�C) tcd (�C) D

1 (Fan 1) 9.358 )21.0 44.8

2 (Fan 1) 9.208 )21.4 44.1

3 (Fan 1) 9.678 )21.6 45.7

4 (Fan 1) 9.358 )0.5 55.6

5 (Fan 1) 9.208 )0.7 54.9

6 (Fan 1) 9.678 )1.2 57.1

7 (Fan 2) 9.204 )21.4 47.9

8 (Fan 2) 9.052 )20.6 47.8

9 (Fan 2) 9.424 )20.7 49.3

10 (Fan 2) 9.204 )0.9 61.6

11 (Fan 2) 9.052 0.2 61.5

12 (Fan 2) 9.424 )0.1 63.9

Table 3

Summary of the main values for the condenser

Test ta;su;cd (�C) ta;ex;cd (�C) Dtsc (�C) _QQcd (W)

1 (Fan 1) 32.0 38.0 6.0 8367

2 (Fan 1) 32.0 37.8 3.9 8197

3 (Fan 1) 32.0 38.2 9.1 8348

4 (Fan 1) 32.2 44.2 9.1 15576

5 (Fan 1) 32.1 44.0 6.9 15320

6 (Fan 1) 32.2 44.6 13.5 15647

7 (Fan 2) 32.2 41.4 3.3 8094

8 (Fan 2) 32.1 41.0 2.9 8310

9 (Fan 2) 32.2 41.5 8.1 8394

10 (Fan 2) 32.2 48.6 9.6 14805

11 (Fan 2) 32.4 49.7 5.4 15254

12 (Fan 2) 32.4 49.8 13.6 15456

4.2. Condenser

Table 3 shows a summary of the main values com-

puted for the condenser. In this Table, are given the

results obtained with two different fans.Quite constant air mass flow rates are obtained for

each fan: this is a good trace for confirming that the tests

were well carried out. Another information is given by

the order of magnitude: fan 1 has 46% more mass flow

rate than fan 2. Fan 1 has a mean air flow rate of 1.319

kg/s and fan 2 a mean airflow of 0.890 kg/s. For each

case, the difference between the mean values for the air

mass flow rate and the maximum and minimum value islower than �5%.

Supplementary measurements of the air mass flow

rate were carried out in order to confirm the results

obtained. The air velocity was measured at 4� 10

(rows� columns) points, 5 mm in front of the condenser

grid with a hot wire probe (10% accuracy). The sum-

mary is given in Table 4.

As it can be seen, the air mass flow rate measured viathe air velocity for the fan 1 and 2 agree well (about 10%

of difference) with the air mass flow rate computed via

the condenser balance.

tsc (�C) _MMcor (Kg/s) _QQeh (W) _WWcp (W)

6.0 0.0427 5629 3872

3.9 0.0420 5475 3854

9.1 0.0421 5584 3843

9.1 0.0962 11179 5644

6.9 0.0957 10946 5628

13.5 0.0961 11076 5615

3.3 0.0431 5210 3882

2.9 0.0425 5527 4001

8.1 0.0433 5483 3993

9.6 0.0946 10012 5990

5.4 0.1046 10089 5992

13.6 0.0963 10490 6253

_MMa (kg/s) LMTDcd (K) UAcd (W/K) SCC (W/K)

1.374 9.4 886 637

1.385 8.9 925 656

1.336 10.2 816 597

1.285 16.7 932 648

1.280 16.1 951 655

1.255 18.0 870 616

0.870 10.5 773 500

0.927 10.6 781 518

0.889 11.8 711 481

0.900 20.1 737 496

0.874 19.2 796 515

0.881 21.6 714 484

Fig. 8. Influence of the sub-cooling on the UAcd.

Fig. 7. Compressor isentropic and volumetric efficiencies.

Table 4

Measurement of the airflow rate

Fan 1 Fan 2

N (rpm) 1380 940

V a (m/s) 2.46 1.72_MMa (kg/s) 1.18 0.82

C. Cuevas, E.L. Winandy / Experimental Thermal and Fluid Science 27 (2003) 781–787 787

With the results obtained in these tests, the evolution

of the global heat transfer coefficient can be analyzed,

Fig. 8 shows the results. Of course the global UAcd

(assumed as one zone heat exchanger) depends on the

degree of sub-cooling. In general, when the sub-cooling

increases the global heat transfer coefficient of the con-denser decreases. This is observed for both fans.

The conclusion is that for the conditions reached

during the tests, the UAcd values decrease about 14 W/K

for each increasing of the sub-cooling of 1 K.

5. Conclusions

Two methods for testing condensing units are pre-

sented in this paper. Advantages and drawbacks of each

method are discussed and the precautions necessary totake into account for separate plenums methodology.

The main component of the condensing unit to study

here was the condenser. For this component, thermal

balance was developed in order to calculate the air mass

flow rate that passes through the condenser. With the

method proposed, the air mass flow rate can be pre-

dicted with a discrepancy lower than 5%. The air mass

flow rate computed by balance is compared with the airmass flow rate measured. There exists a good agreement

between both values, the difference being lower than

10%.

As an example, with the results presented in this

paper an analysis of the UAcd is shown in function of

the degree of sub-cooling. This parameter decreases with

the degree of sub-cooling of about 14 W/K per K of sub-

cooling.

Acknowledgement

This study has been supported by COPELAND Eu-

rope-Welkenraedt S.A.

References

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circulation air-cooling and air-heating coils–test procedures for

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