# HEFAT2012 9 International Conference on Heat Transfer, Fluid Mechanics and Thermodynamics ```HEFAT2012
9th International Conference on Heat Transfer, Fluid Mechanics and Thermodynamics
16 – 18 July 2012
Malta
SIMULATION MODEL OF A SMALL POWER REFRIGERATION CYCLE
DEMONSTRATION UNIT
Martínez-Suárez J.A.a, Sieres J.a*, Martín E.b
Área de Máquinas y Motores Térmicos, Escuela de Ingeniería Industrial, Campus Universitario LagoasMarcosende, University of Vigo, 36310 Vigo, Spain
b
Área de Mecánica de Fluidos, Escuela de Ingeniería Industrial, Campus Universitario Lagoas-Marcosende,
University of Vigo, 36310 Vigo, Spain
* Author for correspondence: E-mail: [email protected], Tel: +34 986 811997, Fax: +34 986 811995
a
ABSTRACT
This paper is focused on a small power refrigeration cycle
demonstration unit to be used at different operating conditions
and in various fields of study. The refrigeration unit is planned
to be used principally for two different kinds of studies:
thermodynamic analyses of the refrigeration system
performance at different evaporation and condensation
temperatures; and determination of the heat transfer coefficients
during the condensation and evaporation processes.
In order to design and select its different components, a
detailed simulation model was developed. The model takes into
account specific data, characteristics and dimensions of the
main components.
In the paper, the main components of the refrigeration unit
are described, the mathematical simulation model is detailed
and the main results of the simulation model are presented and
analyzed.
NOMENCLATURE
Cp
d
DS
f
h
k
L
m
N
Ns
p
pm
q
q”
R
RC
R2
sm
T
V
Vt
W
[J/(kg K)]
[m]
[K]
[Hz]
[J/kg]
[W/(m K)]
[m]
[kg/s]
[kPa]
[-]
[W]
[W/m2]
[K/W]
[-]
[K], [ºC]
[m3/s]
[m3]
[W]
Specific heat capacity
Diameter
Degree of superheat
Frequency
Specific enthalpy
Thermal conductivity
Length
Mass flow rate
Speed regime
Synchronous speed
Pressure
Number of magnetic pole pairs per phase
Heat flow
Heat flux
Thermal resistance
Pressure ratio
Coefficient of determination
Slip
Temperature
Volumetric flow rate
Compressor displacement
Power
Greek symbols
α
[W/(m2 K)]
∆Tlm
[K]
ρ
[kg/m3]
η
[-]
Heat transfer coefficient
Logarithmic mean temperature difference
Density
Efficiency
Subscripts
ac
air
cl
C
e
E
elec
i
in
is
n
out
ov
sat
t
ref
Vol
w
Air-cooled heat exchanger
Air
Convective liquid flow
Condenser
External
Evaporator
Electric
Internal
Inlet
Isentropic
Nucleate pool boiling
Outlet
Overall
Saturation conditions
Tube wall
Refrigerant
Volumetric
Water
INTRODUCTION
Air-conditioning and refrigeration installations account for
about 30 percent of worldwide energy consumption . Thus
deep knowledge of this type of installations is of great
importance in the actual energy and industrial scenario. Even
though, there are different refrigeration technologies in the
market, the most extended one is the vapor compression
technology.
As a result, most engineering courses all over the world
cover, in more or less depth, fundamental and advanced topics
related with vapor compression systems. Usually the preferred
method of teaching technological disciplines is an alternation
between traditional classrooms, in which the teacher introduces
the contents of the subject matter or solves exercises, and other
activities such as the use of multimedia files, web-based
applications, simulation software and experimental equipment.
In the field of refrigeration a broad number of examples exist,
such as those found in Refs. [2-5].
1658
leaves it as a superheated vapor flow. The degree of superheat
at the evaporator outlet is controlled manually by a commercial
thermostatic expansion valve.
The condenser consists of two concentric copper tubes of
1 m length which are oriented vertically. The inner and outer
tubes are of diameters 15.87/14.13 mm and 9.52/7.92 mm,
respectively. The refrigerant condensation takes place as it
flows downwards inside the inner tube. Water is used as the
cooling medium which flows through the annular space formed
by the concentric tubes, in counter current with the refrigerant
flow.
The cooling water flows in a closed loop consisting of a
circulating pump and a forced air-cooled heat exchanger,
constructed with copper tube of 3/8” in staggered arrangement
and continuous aluminum fins. The cooling water temperature
can be adjusted by controlling the airflow of the fan of the aircooled heat exchanger by means of a variable frequency drive.
The compressor is a commercial hermetic compressor with
a displacement of 3.13 cm3. The refrigerant mass flow rate is
controlled by varying the compressor supply frequency by
means of a variable frequency drive.
Other components of the experimental demonstration unit
are a thermostatic expansion valve, solenoid valve, filter and
proper instrumentation, as depicted in Figure 1.
A number of experimental equipment for lecture in vapor
compression refrigeration systems exist in the market.
However, this type of equipment is usually either expensive or
limited to a set of experiments focused on the demonstration
principles of the technology rather than on the possibility of
carrying out detailed and controlled thermodynamic or heat
transfer analyses.
This paper is focused on a small power refrigeration
demonstration unit based in the vapor compression cycle. The
refrigeration unit is planned to be used principally for two
different kinds of studies: thermodynamic analyses of the
refrigeration system performance at different evaporation and
condensation temperatures; and determination of the heat
transfer coefficients during the condensation and evaporation
processes.
The design and selection of its components are focused to
meet the following challenges:
1. The user should be able to perform basic thermodynamic
analyses; as a result, the experimental unit has to be equipped
with proper instrumentation.
2. The user should be able to perform (and repeat)
experimental tests at the desired operating conditions (basically
condensation and evaporation temperatures); thus, the
experimental unit has to be equipped with adequate operational
or electronic control systems.
3. The user should be able to calculate the condensation and
evaporation heat transfer coefficients. This requires that the
instrumentation used and the operating conditions found in the
condenser and evaporator should be studied in detailed in order
to be favorable for measuring the heat transfer coefficients with
a reasonable accuracy.
In order to design and select its different components, a
detailed simulation model was developed. The model takes into
account specific data, characteristics and dimensions of the
main components. In the paper, the main components of the
refrigeration unit are described, the mathematical simulation
model is detailed and the main results of the simulation model
are presented and analyzed.
Figure 1 Schematic representation of the experimental
demonstration unit.
DESCRIPTION OF THE DEMONSTRATION UNIT
The mathematical simulation is based on an experimental
demonstration unit which operating scheme is depicted in
Figure 1. The demonstration unit consists of the main
components found in a vapor compression refrigeration unit:
compressor, condenser, expansion device and evaporator. The
compressor and expansion device are commercially available
but the heat exchangers (condenser and evaporator) have been
specially design to perform different heat transfer analyses. The
refrigerant used is R134a.
The evaporator consists of a vertical copper tube of
diameter 9.52/7.92 mm and 1 m length. The evaporator is
covered by a flexible power heating cable that supplies a
uniform heating power along the evaporator length. The total
heating power supplied to the evaporator (equal to the
evaporator cooling capacity) is adjusted modifying the heating
cable supply voltage by means of a rheostat. The refrigerant
flows vertically from the bottom to the top of the evaporator.
The refrigerant enters the evaporator as a two-phase flow and
SYSTEM MODELLING
A steady state mathematical model has been developed in
order to simulate the vapor compression demonstration unit.
The model takes into account specific data and dimensions of
the components in order to predict the real operating conditions
of the unit. The following general assumptions have been
considered:
1. The system operates at steady-state conditions.
2. The pressure drops in the refrigerant pipes circuit are
negligible.
3. Heat transfer between the heat exchangers and the
surroundings is negligible
4. The heat losses from the refrigerant and water lines are
negligible.
5. Saturated liquid conditions are considered at the outlet of
the condenser.
6. In order to obtain the thermodynamic state at the
compressor outlet, the compression process is considered to be
1659
isentropic. Later, the overall compressor efficiency is used to
calculate the electric power of the compressor.
In the following sections the basic characteristics and
equations of the main components of the experimental unit are
detailed.
0.7
ηvol (60 Hz)
Efficiency
0.6
Compressor
The Danfoss TL3G hermetic compressor was selected with
a displacement of 3.13 cm3. The refrigerant used is R134a. The
manufacturer reports data for the capacity and power
consumption of the compressor for a condensing temperature of
55 ºC and an evaporating temperature range of -30 to 15 ºC,
measured according to the test conditions of EN
12900/CECOMAF and ASHRAE standards.
The pressure ratio of the compressor is defined as the ratio
of the discharge pressure to suction pressure. According to
assumption 2, these pressures are the condensation and
evaporation pressures, respectively. Then, the pressure ratio is
given by equation (1):
RC =
pC
pE
2⋅π ⋅ f
pm
ηov (60 Hz)
0.4
0
5
10
15
20
Compression ratio (RC)
Figure 2 Volumetric efficiency and overall efficiency of the
compressor as a function of the pressure ratio.
The isentropic power of the compressor (Wis) can be
calculated as stated in equation (6):
Wis = mref ⋅ (h2is − h1 )
(1)
(6)
where h2is is the specific enthalpy of the refrigerant vapor at the
discharge pressure and specific entropy of the refrigerant vapor
entering the compressor.
The overall efficiency of the compressor is obtained as the
ratio of the isentropic power to the real power consumption
obtained from the manufacturer data:
(2)
η ov =
Wis
Welec
(7)
Only the compressor power consumption is given by the
manufacturer, so only the overall efficiency could be
determined and other compressor efficiencies such as the
indicated or electrical efficiency could not be calculated.
Figure 2 also shows the values of the compressor overall
efficiency as a function of the pressure ratio obtained from the
manufacturer data. With respect to the results of the volumetric
efficiency, two main differences are observed. The first one is
that a linear dependence between the overall efficiency and the
pressure ratio cannot be assumed, so a higher order polynomial
fit should be considered. The second difference is that the
overall efficiency is affected by the compressor supply
frequency (50 or 60 Hz). Anyway, due to the information is
only limited to two frequency values (50 and 60 Hz), the effect
of the frequency was neglected and the overall efficiency was
assumed to be only a function of the pressure ratio. Eq. (8) is
the result of the regression analysis with a coefficient of
determination R2 of 80%.
(3)
Ns − N
(4)
Ns
where Ns is the synchronous speed, f is the supply frequency, pm
is the number of magnetic pole pairs per phase and s is the slip.
In our case pm = 1 and a typical value of sm = 5% for small
motors has been assumed.
sm =
The volumetric efficiency was obtained by a simple
regression analysis from the refrigeration data capacity
provided in the manufacturer technical data. Figure 2 shows the
values of the volumetric efficiency as a function of the pressure
ratio. It can be seen that a linear dependence exists between the
experimental values of the volumetric efficiency and the
pressure ratio, and also that this dependency is not affected by
the compressor supply frequency (50 or 60 Hz). Equation (5) is
the result of the regression analysis with a coefficient of
determination R2 higher than 99%.
ηvol = −0.02372⋅ RC + 0.7396
0.5
0.2
The manufacturer reports data for different operating
conditions and supply frequencies of 50 and 60 Hz. The
compressor speed (N) is related to the supply frequency by
equations (3) and (4):
Ns =
ηov (50 Hz)
Best
interpretation
0.3
The refrigerant mass flow rate can be obtained from the
compressor displacement (Vt), speed regime (N), volumetric
efficiency ( ηVol ) and the refrigerant density at the suction
conditions ( ρ1 ), as stated in equation (2):
mref = ρ1 ⋅Vt ⋅ηvol ⋅ N (2⋅π )
ηvol (50 Hz)
ηov =1.24⋅10−4 ⋅ RC 3 − 5.37⋅10−3 ⋅ RC 2 + 0.05786⋅ RC + 0.2139 (8)
Thermostatic valve
The thermostatic valve is assumed to be perfectly isolated,
so the refrigerant suffers an isenthalpic process (h3 = h4). The
(5)
1660
In this expression, q′n′ is the nucleate pool boiling heat flux
calculated from Rohsenow  based on the assumption that
the bulk of the liquid is stationary. The term qc′′ is the singlephase convective heat flux obtained from equation (16):
thermostatic valve guarantees a constant degree of superheat at
the evaporator outlet, then:
(9)
T1 = Tsat ( p E ) + DS
where Tsat represents the saturation temperature at the
evaporator pressure and DS is the constant degree of superheat
controlled by the thermostatic valve.
(16)
qc′′ =α cl ⋅(TtE −TE )
where αcl is calculated from the Dittus-Boelter correlation 
based on the assumption that a single-phase liquid flow at
temperature TE flows inside the tube. Rohsenow and Griffith
 recommend calculating αcl by replacing the coefficient
0.023 with 0.019 in the Dittus-Boelter correlation.
Condenser
The condenser consists of two concentric copper tubes
oriented vertically of 1 m length. The heat transfer rate from the
refrigerant to the cooling medium (water) can be obtained by
applying energy balances for each stream – equations (10) and
(11) – and the heat transfer equation (12).
qC = mref ⋅ (h2 − h3 )
(10)
qC = mw ⋅Cpw ⋅ (T5 − T6 )
(11)
qC =
∆Tlm,C
Refrigeration system
The system COP is obtained from equation (17):
COP =
In this equation, the refrigeration capacity is calculated as
stated in equation (14) and the electric power consumption of
the compressor is calculated using equations (6) to (8).
(13)
Cooling water pump
The Wilo Star-RS 25/7 circulation pump is used for the
cooling water circuit. The total head (kPa) for different water
flow rate values (m3/s) was obtained from the manufacturer
technical data. These values were fitted by a simple regression
analysis which result yields equation (18) with a coefficient of
determination R2 higher than 99%.
and RC is the overall thermal resistance in the condenser:
RC = RiC + RtC + ReC
1
ln(deC / diC )
1
+
+
α iC ⋅π ⋅diC ⋅ LC 2⋅π ⋅kt ⋅ LC α eC ⋅π ⋅ deC ⋅ LC
The condensation process takes place inside the inner tube
of the condenser, so it is affected by the vapor flow which also
flows downwards. The liquid film is then expected to be thinner
than in the absence of vapor flow, so a higher condensation
heat transfer coefficient should be expected. In this work, the
average heat transfer coefficient (αiC) has been obtained from
McNaught and Butterworth .
The heat transfer coefficient (αeC) for turbulent flow in the
concentric annular duct was calculated according to Petukhov
and Roizen , using a modified form of the Gnielinski 
correlation for turbulent flow in tubes as a function of the inner
and outer tubes diameters ratio.
∆p = −1.468 ⋅10 7 ⋅Vw2 −1.4272 ⋅10 4 ⋅Vw + 65.57
qac = qC = q15 ⋅
where ∆Tlm,ac
∆Tlm,ac
(19)
15
is the logarithmic mean temperature difference
between the water and air flows, and q15 is the heat transfer rate
for a temperature difference of 15 K. Equation (19) is based on
performance data for an air cooled refrigerant condenser and
not an air cooled water flow heat exchanger. However, equation
(19) was considered to be valid (or at least conservative) based
on the following considerations: a quick evaluation of the
overall heat transfer coefficient indicated that the air side
thermal resistance was the higher one; the water flow rate is
very high, so the water side temperature is nearly constant; and,
the water side flow regime is turbulent, so high water side heat
transfer coefficients are obtained.
(14)
The boiling heat transfer process inside the vertical tube of
the evaporator is modeled by the additive formula
recommended by Rohsenow and Griffith :
q′E′ = qn′′ + qc′′
(18)
Air-cooled heat exchanger
The forced air-cooled heat exchanger has a heat transfer
area of 1 m2 and it is constructed with eight copper tubes of
9.5/7.9 mm in staggered arrangement and continuous aluminum
fins. The configuration is of a single-pass, cross-flow heat
exchanger with both fluids unmixed.
The heat transfer rate is estimated based on manufacturer
data (EN-327:2000) by equation (19).
Evaporator
The evaporator consists of a vertical copper tube of 1 m
length. A uniform heating power (qE) along the evaporator
length is supplied by means of a flexible power heating cable.
The heat transfer rate can be related with the refrigerant
conditions by applying the energy balance for the refrigerant
flow, as stated in equation (14).
qE = mref ⋅(h1 − h4 )
(17)
(12)
RC
where ∆Tlm,C is the logarithmic mean temperature difference
=
qE
Welec
(15)
An energy balance on the air side gives equation (20):
1661
qac =Vair ⋅ ρ air ⋅Cpair ⋅ (Tair ,out − Tair ,in )
where Vair is the nominal air flow rate.
Thermodynamic analysis
Table 2 shows the numerical results of the simulation
model, considering the data in Table 1. It can be seen, that for
the given data, the system COP is 1.72. The maximum ambient
temperature of 29.4 ºC indicates that if the ambient temperature
is higher than this value then it would not be possible to carry
out this experiment (for the given values of condensation and
evaporation temperatures). In contrast, if the ambient
temperature is lower than the calculated value, then the
experiment can be easily attained if the speed of the fan of the
air-cooled heat exchanger is adjusted to the required value by
means of the variable frequency drive. Similarly, in order to
operate at the input evaporation temperature, the heating cable
supply voltage should be adjusted by means of the rheostat in
order to obtain the heating power value (equal to the cooling
power) shown in Table 2.
(20)
Cooling water circuit
A total tube length of 2 m of copper tube of diameter
15.9/14.1 mm is considered for the cooling water circuit,
excluding the condenser length and the tube-side length of the
air-cooled heat exchanger. The pressure drops in the connecting
tubes as well as the pressure losses in the piping components,
air-cooled heat exchanger and flow-meter were considered in
conjunction with the pump curve to determine the water flow
rate .
RESULTS AND DISCUSSION
The mathematical model described in the previous section
has been programmed using Engineering Equation Solver
(EES) . The model has been used to simulate and analyze
the performance of the experimental vapor compression
demonstration unit under different operating conditions. The
input parameters and operating conditions considered for the
analysis are indicated in Table 1.
Table 2. Calculated parameters of the experimental demonstration unit
Parameters
Vapor compression system
Evaporation pressure
Table 1. Operating conditions and components data.
Parameters
Values
Compressor
Displacement
3.13·10-6 m3
Values
350 kPa
Condensation pressure
Refrigerant mass flow rate
Cooling power
1320 kPa
1.61 10-3 kg/s
217 W
Condensation heat power
Compressor isentropic power
Compressor electric power
263 W
46 W
126 W
Number of magnetic pole pairs per phase
Number of phases
Slip
1
1
5%
COP
Cooling water circuit
Water flow rate
1.72
Supply frequency
Volumetric efficiency
Overall efficiency
50 Hz
Eq. (5)
Eq. (8)
Inlet water temperature to condenser (T6)
Outlet water temperature from condenser (T5)
Air-cooled heat exchanger
39.0 ºC
39.2 ºC
Condenser
Length
Tube material
Inner tube diameters
Outer tube diameters
Condensation temperature
Evaporator
Length
Tube material
Maximum required ambient temperature (Tair,in)
Logarithmic mean temperature difference
1m
Copper
15.9/14.1 mm
9.5/7.9 mm
50 ºC
9.5/7.9 mm
5 ºC
Degree of superheat
Water pump
Pump curve
5K
Eq. (18)
Air-cooled heat exchanger
Nominal temperature difference
Nominal heat transfer rate (q15)
15 K
450 W
Nominal air flow rate
Number of tubes
Tubes length
0.115 m3/s
8
0.262 m
Tube material
Tubes diameters
Copper
9.5/7.9 mm
29.4 ºC
8.8 ºC
It is known that for a vapor compression system, the COP
increases with increasing values of the evaporation temperature
and decreasing values of the condensation temperature. Figure
3 shows the numerical results predicted by the simulation
model for condensation and evaporation temperature values in
the range from 35 to 60 ºC and from 0 to 15 ºC, respectively.
Figure 4 shows the results for the pressure ratio for the same
cases analyzed in Figure 3. It can be seen that the pressure ratio
increases with increasing values of the condensation
temperature and with decreasing values of the evaporation
temperature. As a result, an opposed trend should be expected
for the compressor volumetric efficiency, which is confirmed in
the numerical results shown in Figure 5. Figure 5 also shows
the results for the overall compressor efficiency. For the range
of condensation and evaporation temperature values analyzed,
the pressure ratio is always lower than 6; then, from equation
(8) and figure 2, it should be expected that the overall
compressor efficiency increases with increasing values of the
condensation temperature and with decreasing values of the
evaporation temperature, as confirmed in Figure 5.
1m
Copper
Tube diameters
Evaporation temperature
Thermostatic valve
3.34·10-4 m3/s
1662
The range of evaporation and condensation temperature
values that would be possible to test with the refrigeration cycle
demonstration unit will be constrained by the ambient air
temperature. Figure 6 shows the calculated ambient
temperature values (Tair,in(max)) required to perform these
experimental tests. If the actual ambient temperature is lower
than the values shown in figure 6, then the experiment can be
attained reducing the speed of the fan of the air-cooled heat
exchanger by means of the variable frequency drive. In
contrast, if the ambient temperature is higher than the values
calculated in figure 6, then it would not be possible to perform
the experiment with the corresponding evaporation and
condensation temperature values. Results shown in figure 6
indicate that Tair,in(max) decreases with the decreasing values of
the condensation temperature and with increasing values of the
evaporation temperature. This last behavior is expected, since
when the evaporation temperature is increased for a fixed value
of the condensation temperature, then the compression ratio
decreases and, according to figure 4, the volumetric efficiency
increases. As a result, higher values of the refrigerant mass
flow rate and higher values of the cooling and condensation
powers are obtained.
4.5
4.0
COP
3.5
3.0
2.5
2.0
1.5
1.0
0
5
10
15
TE (ºC)
35
40
45
50
55
60
Figure 3 System COP as a function of the evaporation
temperature for different values of the condensation
temperature.
6.0
5.0
RC
50
4.0
40
Tair,in(max) (ºC)
3.0
2.0
1.0
0
5
10
15
30
20
10
TE (ºC)
0
35
40
45
50
55
60
0
5
10
15
TE (ºC)
Figure 4 Pressure ratio (RC) as a function of the
evaporation temperature for different values of the
condensation temperature.
35
40
45
50
55
60
Figure 6 Maximum required ambient temperature as a
function of the evaporation and condensation temperatures.
0.7
Heat transfer analysis
The refrigeration cycle demonstration unit is also planned to
be used for the determination of the heat transfer coefficients
during the condensation and evaporation processes.
In a first stage, the condensation heat transfer coefficient
will be studied for different pure refrigerants and refrigerant
mixtures. Since only fluid temperatures are measured (wall
temperatures are not measured) a tentative technique for
determining the condensation heat transfer coefficient will be
the use of the Wilson plot method  or some of its
modifications . In order to apply the Wilson plot method
successfully, at least the following conditions are desirable: a)
the temperature difference between the condensing refrigerant
and the water stream is high enough to be only gently affected
by the accuracy of the temperature sensors; b) the main thermal
resistance in the condenser is the condensation side thermal
resistance; and, c) during different experimental tests the wall
0.6
Efficiency
ηVol
0.5
ηov
0.4
0.3
0
5
10
15
TE (ºC)
35
40
45
50
55
60
Figure 5 Volumetric and overall compressor efficiencies as
a function of the evaporation temperature for different values of
the condensation temperature.
1663
conductive and water convective thermal resistance remain
nearly constant when compared with the overall thermal
resistance variation.
Figure 7 shows the results for the logarithmic mean
temperature difference in the condenser ( ∆Tlm,C ). It can be
correlations can be as high as 30%. Therefore a sensitivity
analysis of the effect of the heat transfer coefficients on the
overall thermal resistance has been performed. The results of
this analysis are shown in Figure 9. The water heat transfer
coefficient, the condensation heat transfer coefficient and the
tube wall thermal conductivity have been multiplied by a factor
varied from 0.1 to 10. It can be seen that the condensation heat
transfer coefficient (αiC) has the most significant effect on the
overall thermal resistance (RC). The water side heat transfer
coefficient (αeC) is the next more important factor, though its
effect is much lower. Finally, the tube wall thermal resistance is
negligible. These results show that the experimental unit design
and operating conditions are adequate to experimentally
measure the condensation heat transfer coefficient even if the
water side heat transfer coefficient calculation cannot be
estimated with accuracies better than 30%.
seen that for all the operating conditions considered, ∆Tlm,C is
always higher than 8 ºC, which will be easily measured (with
sufficient accuracy) in the future experiments.
∆Tlm,C (K)
20
15
10
0.20
αiC
αiC
αeC
αeC
0.15
0
5
35
40
10
TE (ºC)
45
RC
5
15
0.10
50
55
60
Figure 7 Logarithmic mean temperature difference in the
condenser as a function of the evaporation and condensation
temperatures.
0.05
0.00
The results of the thermal resistances in the condenser are
collected in Figure 8. It can be seen that the overall thermal
resistance in the condenser (RC) increases with the evaporation
and condensation temperatures. For all the cases analyzed,
results in Figure 8 show that the thermal resistance of the
condensation process accounts for more than 90% of the overall
thermal resistance in the condenser.
0.046
0.1
0.044
CONCLUSIONS
This paper covered the design and simulation of a small
power refrigeration cycle demonstration unit to be used at
different operating conditions and in various fields of study. In
order to predict the performance of the refrigeration unit, the
simulation model accounted for specific data, characteristics
and dimensions of the main components.
The results of the simulation model indicated that the
refrigeration unit will be able to be used for two different kinds
of studies: thermodynamic analyses of the refrigeration system
performance at different evaporation and condensation
temperatures; and determination of the heat transfer coefficients
during the condensation and evaporation processes.
The evaporation temperature will be experimentally
controlled by adjusting the total heating power supplied to the
evaporator by means of a rheostat that varies the heating cable
supply voltage. On the other hand, the condensation
temperature will be controlled by varying the airflow of the fan
of the air-cooled heat exchanger by means of a variable
frequency drive. However, it was shown that for some
combinations of the evaporation and condensation
temperatures, it may not be possible to carry out the experiment
because it would require too low ambient air temperatures.
98
0.038
94
RiC/RC (%)
RC (K/W)
96
0.036
92
0.034
0.032
90
5
10
15
TE (ºC)
35
40
45
50
55
10
MF
0.042
0.04
1
Figure 9 Effect of the heat transfer coefficients on the
condenser thermal resistance.
100
0
ktkt
60
Figure 8 RC and ratio RiC/RC as a function of the
evaporation and condensation temperatures.
All previous results rely on the assumption that the heat
transfer coefficients can be determined or estimated from
published correlations. However, typical accuracies of these
1664
ACKNOWLEDGMENTS
The authors would like to acknowledge the financial
support
from
the
“Xunta
de
Galicia”,
Project
09REM004303PR.
 Petukhov B.S. Roizen L.I., Generalized relationship for heat
transfer in a turbulent flow of a gas in tubes of annular section, High
Temp. (URSS), Vol. 2, 1964, pp. 65-68.
 Gnielinski V., New equations for heat and mass transfer in
turbulent pip and channel flow, Int. Chem. Eng., Vol. 16, 1976, pp.
359-368.
 Rohsenow W.M. Griffith P., Correlation of maximum heat flux
data for boiling of saturated liquids, AIChE-ASME Heat Transfer
Symp., Louisville, Ky., 1955.
 Rohsenow W.M., A method of correlating heat transfer data for
surface boiling liquids, Trans. ASME, Vol. 74, 1952.
 Dittus PW. And Boelter L.M.K., Heat transfer in automobile
radiators of the tubular type, Int. Commun. Heat Mass Transfer, Vol.
12, 1985, pp. 3-22.
 Gersten K., Ducts, in: Hewitt G.F. (Ed.), Heat Exchanger Design
Handbook, Begell House, New York, 1998, 2.2.2.
 Klein S.A., Engineering Equation Solver (EES), F-Chart
Software, Middleton, WI, 1992-2006.
 Wilson E.E., A basis for rational design of heat transfer apparatus,
Trans. ASME, Vol. 37, 1915, pp. 47–82.
 Fernández-Seara J., Uhía F.J., Sieres J., Campo A., A general
review of the Wilson plot method and its modifications to determine
convection coefficients in heat exchange devices, Appl. Therm. Eng.,
Vol. 27, 2007, pp. 2745-2757.
REFERENCES
 Buzelin L.O.S, Amico S.C., Vargas J.V.C., Parise J.A.R.,
Experimental development of an intelligent refrigeration system, Int.
J. Refrigeration, Vol. 28, 2005, pp. 165–175.
 Sieres J., Fernández-Seara, J., Simulation of compression
refrigeration systems, Comput. Appl. Eng. Educ., Vol. 14, 2006, pp.
188–197.
 CoolPack 1.49:
 Tan F.-L., Ameen A., Fok S.-C., Engineering courseware on
refrigeration cycle simulation, Comput. Appl. Eng. Educ., Vol. 5,
1997, pp. 115–120
 Abu-Mulaweh H. I., Computer-controlled air-conditioning
experimental apparatus, Comput. Appl. Eng. Educ., Vol. 18, 2010,
pp. 458–468
 McNaught J.M., Butterworth D., Film condensation of pure
vapour, in: Hewitt G.F. (Ed.), Heat Exchanger Design Handbook,
Begell House, New York, 1998, 2.6.2.
1665
```