lntJ. ReJ)'ig. Vol. 21, No. 4, pp.

273-284, 1998
© 1998 ElsevierScience Ltd and IIR
Printed in Great Britain. All rights reserved
ELSEVIER PII:S0140-7007(97)00087-X 0140-7007/98/$19.00+00

Modelling of plate finned tube evaporators and

condensers working with R134A
Jos~ M. Corberfin and M6nica Garcia Mel6n
Applied Thermodynamics Department, Universidad Polit6cnica de Valencia,
E.T.S.I.I, Cno de Vera s/n, 46022 Valencia, Spain
Received 13 June 1997; revised 6 November 1997; accepted 12 November 1997

A model to predict the behaviour of finned tube evaporators and condensers working with
R134a has been developed. For modelling of the refrigerant phase change, evaporation or
condensation, the heat transfer and the pressure drop for the two-phase flow have to be
calculated. Therefore, a number of correlations, the most recommended ones in the reviewed
literature, have been analysed and compared. The results of this comparison are presented for
the evaporation and condensation heat transfer coefficients and for the evaporator frictional
pressure drop. Once the correlations have been implemented in the model and compared with
the experimental results, the ones that work best for the studied heat exchangers have been
ultimately selected.
The experimental study to validate the model has been carried out in a small air-
conditioning unit with cross-flow air-refrigerant type heat exchangers. The results are
compared with model predictions for thermal capacity, refrigerant superheat or subcooling,
and tube-side pressure drop. © 1998 Elsevier Science Ltd and IIR. All rights reserved.
(Keywords: evaporator; condenser; modelling; heat transfer; frictional pressure drop)

Modelisation d'6vaporateurs et de condensateurs

plaques-ailettes utilisant le R-134a
Un modble qui permet de connaftre d'avance le comportement des ~vaporateurs et des
condenseurs employant le R134a a ~t~ d~velopp~. Pour la mod6lisation du changement de
phase du frigorigbne (l'~vaporation ou la condensation), le transfert de chaleur et la chute de
pression pour l'~coulement biphasique doivent ~tre calculus. Donc, un certain nombre de
correlations parmi celles les plus souvent conseillges dans la litt~rature ont ~t~ analys~es et
compar~es. Les r~sultats de cette comparaison sont prgsent~s: lcs coefficients de transfert de
chaleur lors de l'~vaporation et de la condensation et pour la chute de pression par frottement
dans l'dvaporateur sont prdsent~s. Une fois que ces correlations ont ~t~ int~grdes dans le
modble et les comparaisons avec les rdsultats exp~rimentaux ont dt~ effectu~es, les
correlations qui conviennent le mieux pour les dchangeurs de chaleur 6tudi~s ont ~t~
s~lectionn~es.
L'~tude expgrimentale destin~e ~ valider le modble a ~t6 effectu~e sur un petit
conditionneur d'air employant des ~changeurs thermiques ~ courants crois~s (air/
frigorigbne). Les rdsultats obtenus sont compares aux prgdictions du modble concernant la
capacitg thermique, la surchauffe ou le sous-refroidissement du frigorigOne et la chute de
pression dans les conduites. © 1998 Elsevier Science Ltd and IIR. All rights reserved.
(Mots el~s: evaporation;condensation;mod~lisation;transfert de chaleur; perte de pression)

Introduction the motor of change for refrigeration systems towards

improving design systems, planning operation and
Computer-aided modelling and design techniques control strategies. This evolution will help to improve
have enabled significant enhancements in a large system operation and to reduce the electric consumption
number of engineering disciplines over the past few and will bring resource savings during the equipment
years. In the near future, modelling techniques will be development stages.

273
274 J. M. Corberan and M. Garcia Melon

Nomenclature
A heat exchange area Subscripts
U overall heat transfer coefficient for dry
coils i inlet
U' overall heat transfer coefficient for wet o outlet
coils a air
Q rate of heat transfer r refrigerant
m flow rate rfic fictitious property of the saturated air
h heat transfer coefficient calculated at the refrigerant's
i enthalpy temperature
Cr specific heat tot total
T temperature fric frictional
AP pressure drop mom momentum
z length of the tube grav gravitational

Validation of these models with experimental data is of the properties along an element is quite small. This
an important part of the overall process, and this paper procedure is preferred to the one taking longer parts of
sets out how to evaluate one particular model for each the heat exchanger because it allows a more precise
heat exchanger, choosing the most appropriate correla- calculation of the refrigerant heat transfer coefficient,
tion to define the refrigerant side heat transfer and the since it strongly varies with the vapour quality and
frictional pressure drop by comparing predictions with therefore along the heat exchanger.
experimental measurements. A number of papers con- The rate of heat exchanged in each discretised element
cerning heat exchanger models have been published ~-5, is given by the following equation:
but they do not include a comparison of the different
Qtot = Uo'A0"ATlm ( 1)
correlations. The authors have already worked on this
subject and published a paper, 6, where a preliminary being:
analysis on the existing correlations for the condenser is ~rai- T.o
presented. ATom = In( Tr - Tao) (2)
The present paper is concerned only with cases where
(Tr - Tai)
the refrigerant flows inside horizontal finned tubes and
the air flows outside and across them (cross-flow). The and U0 the global heat transfer coefficient evaluated at
heat transfer is due to boiling and condensation. The flow the mean properties of the element, which incorporates
regime has been studied and, after having analysed the influences of the air-side and the refrigerant-side
several flow conditions during condensation and eva- coefficients, and the tube and the finned tube thermal
poration inside the exchangers and calculated the resistance. In order to calculate the finned tube thermal
required characteristic parameters, the Taitel and resistance the procedure recommended by ASHRAE 7
Duckler diagram (see Ref. 2J) showed that the flow has been employed. In this case the ATIm coincides
could be considered as annular. A wider objective of the with the temperature difference because a constant air
present work is to develop and validate a model capable temperature is used.
of dealing with any kind of heat exchanger design and The methodology used to define the heat transfer when
type and also with different refrigerant flow regimes. condensation of the air on the tubes occurs, wet coil, is
based in an analogy of the expression used in dry coils.
While for dry coils the mean logarithmic temperature
Heat exchanger model difference is used (Equation (1)), for wet coils the mean
For the development of this model the heat exchanger logarithmic enthalpy difference is used instead. The rate
geometry has been subdivided into a number of elements of heat exchanged in each discretised element for this
in which the local values of the pressure gradient, heat case is:
flux, etc., are assumed to remain constant. For each
element the thermodynamic variables at the downstream
Qtot = U'o'Ao'Ailm (3)
end of the element are initially calculated from the inlet being
state properties, the heat transfer rate and the pressure iai - iao
gradient at the inlet of the element. The mean of the inlet Ailm = (4)
In (ir_fic -- iao)
and the outlet of these variables is used to recalculate the
(/r_tic -- iai
state at the outlet of the element. This is repeated until
convergence is obtained for the outlet values. Conver- and U'0 the global heat transfer coefficient for wet coils
gence is normally very fast because the rate of variation analogous to the one for dry coils.
 Modelling of plate finned tube evaporators and condensers 275

12 ..~ ....

o ,,'~:~.. -~, - -x - -x - -x - -x---x --xx

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1

vapour quallty

~:~-~ Kandlildar _,t Shah - ,1[] - C h a d d o c k e l al.

- ,IK - K l i m e n k o X Gungor - - -A- - -Collier-Pulling
- - CI. - - S c h r o c k - G r o s s m a n n ----O---Bennet et al. - - • - -VDI
• one-phase

Figure 1 Dimensionless boiling heat transfer coefficient against quality

Figure 1 Coefficient adimensionnel d'dchange thermique it l'~bullition en fonction du titre

The air-side heat transfer coefficient is computed with The local pressure drops considered in this study are
equations from Gray and Webb 8. The refrigerant-side the ones taking place at the 180° bends, which are usually
one-phase (pure liquid or pure vapour) heat transfer incorporated by this kind of heat exchanger. They have
coefficient is calculated using the Dittus-Boelter been calculated following the homogeneous flow
equation 9. assumption.
Concerning the refrigerant-side heat transfer coeffi-
cient in boiling and condensation conditions, it is the
Boiling heat transfer coefficient
objective of this study to determine, among the number
of semi-empirical correlations existing in the technical A number of correlations have been found in the
literature, which of them agree best with the studied heat reviewed literature and for the present work, a
exchangers, and use them to evaluate the predictive comparative study of the most frequently cited ones
capabilities of the model. has been carried out, with the objective of choosing the
An analysis of the axial conduction along the copper most appropriate one for the evaporator studied.
tubes has been made, and after analysing the results, it Figure 1 shows the heat transfer coefficients obtained
seems that for the case studied, this type of conduction with the different correlations studied for a tube with an
does not significantly affect the heat transfer, probably inner diameter of 10 ram, an evaporation pressure of
because the variation of the wall temperatures is very 3 bar and a refrigerant mass flow rate of 0.011 kg/s.
small due to high heat transfer coefficients characteristic These conditions were chosen because they are repre-
of the phase change. However, the axial conduction sentative of the midpoint of the test range used for the
influence could be high at places where only vapour experimental validation of the model: refrigerant flow
flows, therefore, a detailed study dealing with this rate between 0.0097 and 0.0130 kg/s and evaporator
phenomenon is planned in the near future. pressure between 2.46 and 3.59 bar. The evolution of the
The pressure drop of the two-phase flow has been ratio between the two-phase flow heat transfer coeffi-
studied using the one-dimensional separated flow cient (hbif) and the one corresponding to only liquid (hliq)
approach. It can be expressed by the sum of frictional, flow, vs the vapour quality is represented.
momentum, gravitational and local pressure drops. The correlations studied here are from a number of
references ~°-19. The first three have been chosen because
APtot = APfric + APmom+ Apgrav + APIoc (5) they have been proposed in the past few years and the
remainder because they are widely recommended in the
The gravitational and momentum terms can be calculated reviewed literature. Each correlation has been presented
with analytical expressions once the outlet conditions of only for the vapour quality range recommended by its
the element are known. An empirical correlation must be author. Most are valid for qualities between 10 and 90%.
used to evaluate the frictional drop. This study also An interpolation between 10 and 0% (liquid only) and
intended to identify which correlation associated with a between 90 and 100% (vapour only) has been imple-
specific frictional drop, among the many referred to in mented in the model to cover the whole range of qualities
the Literature, predicts best the measured pressure drop because only the correlation recommended by the VDI
of the heat exchangers studied. Atlas 15 covers it all.
276 J. M. Corberan and M. Garcia Mel6n

800O

7OOO

600O
Q.

Q.
300O

200O

1000

I I I ~ I I I I I
0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9
vapour quality

Chisholm [] Soliman ~ Mart.end Nelson

- - 0 - - VDI ~ Homogeneous • one-phase

Figure 2 Frictional pressure drop coefficient against quality

Figure 2 Coefficient de chute de pression par frottement en fonction du titre

The correlation recommended by Kandlikar j°'~ 'Homogeneous model'. In the second group, four
clearly shows the highest results for the convection different correlations, recommended in the reviewed
coefficient, while the one recommended by Klimenko ~2 literature have been studied: Chisholm et al. 23, Soliman
shows the lowest, by an important factor. The remainder et a/. 24, Mueller-Steinhagen 25, Martinelli and Nelson 26
show similar results along all the vapour qualities, it and the one recommended by VDI ~5.
being necessary to implement them in the model to The homogeneous correlations give an almost linear
facilitate comparisons and decide which one works best behaviour increasing with quality, showing a much
for the evaporator studied. Two groups may be lower level than the others. With regard to the separated
distinguished. The first one contains the correlations flow correlations, the one recommended by Martinelli
of Gungor and Winterton 13, Collier and Pulling 14, and Nelson 26 shows the highest level and that by
Chaddock and Noerager 17, VD115 and Shahlr; and the Mueller-Steinhagen 25 the lowest one. The correlations
second contains those by Schrock and Grossman ~8 and by Soliman 24 and that recommended by the VDI Atlas ~5
Bennet et al. 19. The first group shows greater values, give almost identical results, and as will be seen below,
almost double, and as will be seen below in this paper, these two are the ones that best predict the experimental
these are the values that agree best with the experimental results. The correlation by Chisholm z3 gives higher
results obtained. values at small qualities and lower ones at high qualities,
but the results of the total frictional pressure loss across
the evaporator are very similar to those provided by the
Frictional pressure drop coefficient
Soliman 24 and VD115 correlations.
A number of models and correlations for the frictional
pressure drop have been found in the literature., for the
Condensation heat transfer coefficient
present work a comparative study of some of the most
recommended ones is presented, with the objective of A number of correlations have been found in the
choosing the most appropriate one for the model reviewed literature and have been compared with the
developed. Figure 2 shows the results of the comparison objective of choosing the most appropriate one for
of the correlations analysed. the condenser studied. Figure 3 shows the results of the
The correlations can be divided into two main groups: comparison between the analysed correlations. The
evolution of the ratio between the two-phase convection
(1) Homogeneous flow: both phases flow with the same
coefficient and the liquid-phase convection coefficient
velocity.
versus the vapour quality is presented. The results have
(2) Separated flow: each phase has its characteristic
been obtained for a tube inner diameter of 8 mm, a
velocity.
condensation pressure of 9 bar and a refrigerant flow of
In the first group three correlations have been studied 0.011 kg/s. These conditions were chosen because they
by the authors: Mills 2°, Stephan 21 and Hewitt 22. All are representative of the midpoint of the test range used
showed almost identical results for the studied case, for the experimental validation of the model: refrigerant
therefore only the results of one of them have been flow rate between 0.0097 and 0.0130 kg/s and condenser
included in the figure which has been referred to as the pressure between 8.20 and 10.96 bar.
 Modelling of plate finned tube evaporators and condensers 277

lO

,a • 8

¢:
"'O-.i
!,=1 II
=~e Ss

"'"° X X

i
l:
8
"¢=

2
i
t" X xX xX X
X X

o + I ~ I I I I I I
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9
vapour quality

I - - - - - D ~ Cavallini et al. - - .0. - - Traviss et al. ~-, Shah ~, Azer et al.

X Ackers et al 0 - VDI • one-phase

Figure 3 Dimensionless condensation heat transfer coefficient against quality

Figure 3 Coefficient adimensionnel de transfert de challeur & la condensation en fonction du titre

2.5
2.4 I & experime~al X,~l I
2.3

2.2

~ 2.1
x t •x
~ 1.9 X x&x
1.8 &x•
1.7 • t
x
1.6
1.5 I I I I I I
0.0095 0.01 0.0105 0.011 0.0115 0.012 0.0125 0.013
refrlge~nt now r=te (kg/=l

Figure 4 Evaporator thermal capacity against refrigerant flow rate

Figure 4 Capacitd therrnique de l'dvaporateur en fonction du ddbit du frigorigkne

The correlations studied here are from a number of Experimental apparatus and procedure
references27-31,15,7 They have been chosen because they
are widely recommended in most of the papers reviewed. An experimental study to validate the model has been
Each correlation has been presented only for the vapour carried out in a small air-conditioning unit. It is composed
quality range recommended by its author. As in the of cross-flow plate finned tube heat exchangers, a
evaporator case and for calculation purposes, an hermetic compressor of 26 cm 3 displacement and 1 kW
interpolation is needed to cover the whole range of of nominal power, and a thermostatic expansion valve.
qualities. The condenser has two rows with 24 copper horizontal
This figure shows that the correlation recommended finned tubes each, with an inner diameter of 8 mm. The
by Ackers 7 gives much lower values than the other ones. evaporator consists of two rows with 10 copper
The correlation by Aze ad Abis 3°, shows a stronger horizontal finned tubes each, with an inner diameter of
increasing trend vs the quality. The Cavallini 27, VDI ~5, 10mm.
Shah 29 and Traviss 28 correlations are in good agreement The refrigerant temperature at the inlet and outlet
for most of the qualities, and are the ones that work best points of the heat exchangers have been measured with
compared with experimental results. miniature K thermocouples. Additionally, the outer tube
278 J. M. Corberan and M. Garcia Mel6n

10
8
6
4
2

0 I I • I i &l
0
-2

-4

-6

-8

-10
O._rc(E~_
_ _ 0.01 0.0105 0.011 0.0115 0.012 0.0125 0.013
refrigerant flow rate (kg/=)

Figure 5 Relative error for evaporator thermal capacity against refrigerant flow rate
Figure 5 Erreur relative pour la capacit~ thermique par rapport au d~bit du frigorigkne

temperature has been measured all along the heat has been conducted, comparing the predictions of the
exchangers with T thermocouples. Pressure has also model with the experimental results.
been measured at the inlet and outlet of the heat For this comparison only the correlations on the
exchangers with pressure transducers of 0.6% precision refrigerant side have been modified, leaving the air-side
full scale, allowing 1% accuracy on the evaporator and correlation (Gray and Webb) unchanged. According to
condenser pressure. The flowmeter used is a turbine-type the results of Wang eta/. 32, the uncertainty in the air side
specially calibrated for the refrigerant R134a and for the type of exchanger studied lies within 10%, which
adequate for the range of measurements, which works would cause a maximum uncertainty of 6% of the heat
with an accuracy of 1%. Special care was taken to insure transferred on the refrigerant side. Therefore, it is not
a correct degree of subcooling in the condenser in order possible to distinguish between correlations which
to avoid the flashing of vapour in the flowmeter. All the produce a difference in the results within this margin.
electrical signals were registered by computer. Thermo- This small uncertainty is due to the similar order of
couples were individually calibrated, together with the magnitude of the relative resistances in both sides of the
acquisition board. heat exchanger, due to the great surface of the air side.
In order to determine the distribution of the air flowing For instance, in the middle of the condenser, the
through the heat exchangers, sufficiently long tunnels representative values of the thermal resistances of both
were built around them, and the velocity profiles at the sides are: 18.2 mK/w for the refrigerant side and
inlet cross-sections were measured by means of a Pitot 25.6 mK/w for the air side.
tube and a digital micro-manometer. The results showed For the heat transfer during boiling, the calculated
that the air flow had an almost uniform profile, with a results obtained with the correlations of Gungor and
mean value of 2.54 m/s, therefore, flow uniformity has Winterton ~3, Collier and Pulling ~4, Chaddock and
been assumed, using the mean value as the uniform Noerager 17, VD115 and Shah 16 agree quite well with
velocity. the experimental results, giving always slightly lower
Unfortunately, it was not possible to control the values. Therefore, the two correlations that show the best
ambient conditions for the condenser, neither for the predictions would be the Gungor one 13 and that
evaporator. Therefore, in order to spread at maximum recommended by the VDI Atlas ~5. This last one has
the range of the study, measurements were taken at days been chosen as the one to be finally implemented in the
with different ambient conditions, and the evaporator model because it gradually covers the whole range of
load was varied through the modification of the incident qualities.
air flow rate and the use of electric heaters. The range of For the frictional pressure drop, the correlations by
conditions achieved was: refrigerant flow rate between Soliman and Azer 24, Chisholm and Sutherland 23, VDI ~s
0.0097 and 0.0130, evaporator pressure between 2.46 and Martinelli and Nelson 26 show similar results and are
and 3.59 bar and condenser pressure between 8.20 and closest to the experimental ones. They all slightly
10.96 bar. underpredict the pressure drop for the evaporator
studied. Therefore, the best solution seems to be the
Martinelli and Nelson 26 correlation, but modified for
Experimental and simulation results qualities over 80% with another correlation that would
A study of the above mentioned correlations found in the tend gradually to the one-phase vapour limit value. In
literature for heat transfer and frictional pressure drop any case, the correlation by VDI ~5 has been chosen
 Modelling of plate finned tube evaporators and condensers 279

30

x
I!120
X
X
15
I
x
10 I ~ I i -- I I

0.0095 0.01 0.0105 0.011 0.0115 0.012 0.01 25 0.013

relrl~,,rant Ilow rate (kg/s)

Figure 6 Degree of superheat at evaporator exit against refrigerant flow rate

Figure 6 Niveau de surchauffe h la sortie de l'~vaporateur en fonction du d~bit du frigorigkne

20

15

10

0
0
-5

-10

-15

-20
0.0095 0.01 0.0105 0.011 0.0115 0.012 0.0125 0.013
refrigerant flow rate (kg/s)

Figure 7 Relative error for degree of superheat at evaporator exit against refrigerant flow rate
Figure 7 Erreur relative pour le niveau de surchauffe ~ la sortie de l'~vaporateur en fonction du d~bit du frigorigbne

because it is consistent throughout the complete range of illustrate the quality of the simulation results. The
qualities and gives very similar results. following parameters have been measured and analysed
For the heat transfer during condensation, the results for the validation of the model using the finally selected
obtained with the correlations by Cavallini 27, VDI tS, correlations.
Shah 29 and by Traviss 28 are very similar and close to the
experimental ones. The correlation by VD115 has been
Evaporation
finally selected because it provides very good results and
offers again a consistent definition throughout the range Figure 4 shows a comparison between measured and
of qualities. It should be noted that although this selected calculated values of the thermal capacity for the entire
correlation gives accurate results, for qualities greater data set. Figure 5 shows the corresponding relative error.
than x = 0.995 a discontinuity in the VDI curve appears, All of the calculated data lie within +3.5% and - 3 %
therefore, an interpolation between this value and the of the measured evaporator capacity. The simulation
vapour-only coefficient has been implemented in the either overpredicts or underpredicts the data, with a
model. mean ratio of calculated to measured capacity of 1.075.
A sample of 20 data points has been chosen to Figure 6 shows a comparison between measured and
280 J. M. Corberan and M. Garcia Mel6n

18
16

14
12
[ lO

i- 8

6
4

2
0 7

0 2 4 6 8 10 12 14 16 18 20
tube nr.

• T wall measured • T int. measured - - e - T int. calculated

Figure 8 Evolution of the temperature along the evaporator

Figure 8 Evolution de la tempdrature le long de l'evaporateur

12CXJ0

10000
A t x
Z x x
• xXx
E x •

i 4000

I I I I ~ I
O.C.x3K~ 0.01 0.0105 0.011 0.0115 0.012 0.0125 0.013
refflgenmt flow rate (Icg/s)

Figure 9 Total evaporator pressure drop against refrigerant flow rate

Figure 9 Chute de pression totale dans I'dvaporateur en fonction du ddbit du frigorig~ne

calculated values of the degree of superheat at the 15 and 16. Furthermore, it can be seen that the predicted
evaporator outlet for the entire data set. Figure 7 shows evolution agrees well with the experimental results.
corresponding relative error, with reference to the total During the change of phase the temperature slightly
temperature variation along the evaporator.All of the decreases because a slight pressure drop occurs.
simulated data lie within 12% and - 8 . 5 % of the This kind of representation has also been used to
measured degree of superheat. determine which correlation, for the evaporation coeffi-
Figure 8 shows an example of the fluid temperature cient, agrees best with the experimental results, because
distribution obtained by the model for the evaporator. it gives more precise information than the global results
The measured fluid temperature at the inlet and outlet between inlet and outlet of the evaporator.
points, and the outer temperature distribution at the tube
wall are included in order to make evaluation of the
Pressure drop
results possible.
As can be observed in this figure the simulation Figure 9 shows the trends in both measured and
predicts the end of the evaporation in tube 15 which predicted refrigerant pressure drop between evaporator
approximately agrees with the experimental results, inlet and outlet for the entire data set. Figure 10 shows
which show the end of the phase change between tubes the corresponding relative error.
 Modelling of plate finned tube evaporators and condensers 281

25

20

15

10

5
g
0 I I I I I I

-5

-10

-15

-20

-25
0.0095 0.01 0.0105 0.011 0.0115 0.012 0.01 25 0.013
refrigerant f l o w rate (kg/s)

Figure 10 Relative error for the evaporator pressure drop against refrigerant flow rate
Figure 10 Erreur relative pour la chute de pression totale dans l'~vaporateur en fonction du d~bit du frigorigkne

•experimental X model [

~ 2.5
X

L~
1ul
U 2 X

1.5 I I I I I I I
0.009 0.0095 0.01 0.01 05 0.011 0.0115 0.012 0.0125 0.01 3
refrigerant flow rate (kg/s)
Figure 11 Condenser thermal capacity against refrigerant flow rate
Figure 11 Capacitd thermique du condenseur en fonction du d~bit du frigorigkne

10

0 t t • I • ~ t

-2 • !
• &• • • A •
-4

-6

-8

-10
O.OOg5 0.01 0.0105 0.011 0.0115 0.012 0.0125 0.013
refrigerant f l o w rate (kg/s)

Figure 12 Relative error for condenser thermal capacity against refrigerant flow rate
Figure 12 Erreur r~lative pour la capacit~ thermique du condenseur en fonction du d~bit du frigorigkne
282 J. M. Corberan and M. Garcia Mel6n

• experimental X model
X

g3
I=

J2 x ~ x
• t
x
x

0 I I I I I I
0.009 0.0005 0.01 0.0105 0.01 1 0.011 5 0.01 2 0.0125 0.013
reldgerant flow rate
Figure 13 Degree of subcooling at condenser exit against refrigerant flow rate
Figure 13 Degr~ de sous-refroidissernent fi la sortie du condenseur en fonction du ddbit du frigorigkne

I I I L

q)

-2

-3 •
0.009 0.C£,~6 0.01 0.0105 0.011 0.0115 0.012 0.0125 0013
~ m ~ II~ rate (kg/$)
Figure 14 Relative error for condenser thermal capacity against refrigerant flow rate
Figure 14 Erreur relative pour le degr~ de sous-refoidissement ~ la sortie du condenseur en fonction du d~bit du frigorigkne

The simulation model underpredicts the pressure drop. Figure 13 shows a comparison between measured and
All except one data point are underpredicted with a calculated values of the degree of subcooling at the
maximum error of 15%. The ratio of simulated to condenser outlet for the entire data set. Figure 14 shows
measured pressure drop has a mean value of 0.91. It the corresponding relative error, with reference to the
should be noted that the correlation chosen gave one of total temperature variation along the condenser.
the highest frictional drop values. Therefore, the under- As in the capacity case, the simulation normally
prediction could be caused by the calculation of the underpredicts the data with a maximum error of - 3 % .
pressure drop in the 180 ° bends. Figure 15 presents an example of calculated fluid
temperatures inside the condenser, compared with the
measured outer wall temperatures and the inlet and outlet
Condensation
refrigerant temperatures.
Figure 11 shows a comparison of measurements and It shows that the simulation predicts the end of the
predictions of the thermal capacity for the entire data set. condensation in tube 19, which approximately agrees
Figure 12 shows the corresponding relative error. with the experimental results that show the end of the
As can be observed, the simulation normally under- phase - change between tubes 18 and 19. Again, as it
predicts the data, with a maximum error of - 5 . 5 % and a was done for the evaporator, this figure has been used to
mean ratio of simulated to measured capacity of 0.98. determine which correlation, for the condensation
 Modelling of plate finned tube evaporators and condensers 283

8O

70

60

50
A
~4o
I--

20

20

10

0 5 10 15 20 25
tube nr.

• T wall measured • T int. measured - - 0 - - T int. calculated

Figure 15 Evolution of the temperature along the condenser

Figure 15 Evolution de la tempdrature le long du condenseur

coefficient, agrees best with the experimental results, range of operating conditions covered by the tests.
because it gives more precise information than the global Unfortunately, the installation did not allow an extensive
results between inlet and outlet of the condenser. range study. A wider range and more detailed refrigerant
temperature measurements would provide more accurate
experimental results, which are planned tbr the near
Conclusions
future.
A computer simulation model for plate finned tube heat
exchangers has been developed, capable of predicting
the heat transfer of an evaporator or condenser with an
accuracy between _+ 5% in the studied range. References
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