Experimental study on R-134a refrigeration system using
a two-phase ejector as an expansion device
Praitoon Chaiwongsa, Somchai Wongwises
*
Fluid Mechanics, Thermal Engineering and Multiphase Flow Research Laboratory (FUTURE), Department of Mechanical Engineering,
King Mongkut’s University of Technology Thonburi, Bangmod, Bangkok 10140, Thailand
Received 8 May 2006; accepted 14 May 2007
Available online 2 June 2007
Abstract
This paper is a continuation of the authors previous work. In the present paper, the performance of the refrigeration cycle using a
two-phase ejector as an expansion device is experimentally investigated. Refrigerant R-134a is used as working fluid. Motive nozzles having three different outlet diameters are tested. New experimental data that have never been seen before are presented on the effects of the
external parameters i.e. heat sink and heat source temperatures on the coefficient of performance and various relevant parameters i.e.
primary mass flow rate of the refrigerant, secondary mass flow rate of the refrigerant, recirculation ratio, average evaporator pressure,
compressor ratio, discharge temperature and cooling capacity. The effects of size of the motive nozzle outlet on the system performance
are also discussed.
Ó 2007 Published by Elsevier Ltd.
Keywords: Two-phase ejector; Refrigeration; Liquid-recirculation; Coefficient of Performance; Motive nozzle; Expansion device
1. Introduction
In vapor refrigeration cycle, in order to accommodate
the low temperature evaporation process, a saturated or
sub-cooled liquid produced from condensation at high
temperature must be expanded through an expansion
device to low-pressure. Throttling process or isenthalpic
process is the process occurring through an expansion
device. This process produces the thermodynamic losses
and causes a larger amount of the refrigerant to flash into
a vapor than in the isentropic process.
In order to reduce the thermodynamic loss in the expansion process, various researchers have attempted to look
for other appropriate devices. Due to the low cost, no moving parts and ability to handle two-phase flow without
damage, an ejector is considered as an attractive expansion
device in the refrigeration system.
*
Corresponding author. Tel.: +662 470 9115; fax: +662 470 9111.
E-mail address: somchai.won@kmutt.ac.th (S. Wongwises).
1359-4311/$ - see front matter Ó 2007 Published by Elsevier Ltd.
doi:10.1016/j.applthermaleng.2007.05.005
To the best of the authors’ knowledge, information on
the application of an ejector as an expansion device in a
refrigeration cycle is still limited.
Kornhauser [1] investigated the thermodynamic performance of the ejector expansion refrigeration cycle by using
R-12 as a refrigerant under standard conditions, 15 °C
and 30 °C for evaporator and condensor temperatures,
respectively. A theoretical COP improvement of up to
21% over the standard cycle was found. This result was
based on ideal cycle and constant mixing pressure in the
ejector. Harrell et al. [2] used a R-134a two-phase ejector
and the test rig to estimate the COP of the refrigeration
cycle. It was found that the COP improvement ranged from
3.9% to 7.6%. Menegay et al. [3] developed a bubbly flow
tube to reduce the thermodynamic non-equilibrium in the
motive nozzle with R-12 as the refrigerant. This device
was installed upstream of the motive nozzle. The COP of
the system using the bubbly flow tube can be improved
up to 3.8% over the conventional cycle under standard conditions. However, they reported that the result was not as
good as was expected and study of the ejector expansion
Nomenclature
COP
cp,hw
Dni
Dnt
Dne
Lni
Lnc
Lnd
m_ p
m_ s
m_ e
m_ hw
coefficient of performance
specific heat at constant pressure of the hot
water (kJ/kg k)
inlet diameter of motive nozzle (mm)
throat diameter of motive nozzle (mm)
outlet diameter of motive nozzle (mm)
inlet length of motive nozzle (mm)
convergent length of motive nozzle (mm)
divergent length of motive nozzle (mm)
primary mass flow rate of refrigerant (kg/s)
secondary mass flow rate of refrigerant (kg/s)
flow rate of the vaporized refrigerant (kg/s)
mass flow rate of hot water (kg/s)
refrigeration cycle should be extended. Domanski [4] concluded that the ejector efficiency was very sensitive to the
theoretical COP of the ejector expansion refrigeration
cycle. Nakagawa et al. [5] showed that the longer divergent
part provided a longer period of time for the two-phase
flow to achieve equilibrium. He concluded that the longer
the length of the divergent part of the motive nozzle, the
higher the motive nozzle efficiency could be obtained.
The papers mentioned above focused on operating with
a dry-expansion evaporator by installing an expansion
valve downstream of the liquid-vapor separator. However,
the purpose of the ejector is to replace the throttling valve.
With this consequence, any throttling device in the system
should be avoided.
Up to now, there have been only two papers, published
by Disawas and Wongwises [6] and Wongwises and
Disawas [7], dealing with this issue. In their experimental
apparatus, the evaporator was flooded with refrigerant
and became a liquid-recirculation system, in which, in addition to serving as an expansion device, the ejector also
acted as a refrigerant pump for the low-pressure side of
the system.
Although, some information is currently available on
the refrigeration cycle using a two-phase ejector as an
expansion device, the detailed investigation is still lacking.
Especially, it can be noted that no attention has been paid
to the effects of geometric parameters on the performance
of the refrigeration system. This paper is the third in a series and is a continuation of the authors’ previous work.
The main concern of the present study is to experimentally
investigate the effect of the external parameter i.e. heat sink
temperature and heat source temperature, the effect of relevant geometric parameters of the ejector on the performance of the refrigeration cycle using a two-phase ejector
as an expansion device. Moreover, the effect of the outlet
diameter of the motive nozzle on the performance of the
refrigeration system, which has never before appeared in
the open literature, is presented.
me
ne
Qevap
se
Tsink
Tsource
Thw,in
Thw,out
Wcomp
mixing chamber exit
motive nozzle exit
heat transfer rate at evaporator (kW)
suction nozzle exit
heat sink temperature (°C)
heat source temperature (°C)
hot water temperature at the inlet of the evaporator (°C)
hot water temperature at the outlet of the evaporator (°C)
electrical power supplied to the compressor
(kW)
2. Experimental apparatus and procedure
Fig. 1 shows the schematic diagram of the experimental
apparatus. A commercial R-134a is used as the working
fluid. The refrigerant loop consists of the vapor compression cycle components: compressor, condenser, expansion
valve and evaporator, and other accessory parts – the oil
separator, liquid receiver, filter/drier, sight glass and the
accumulator. The operating conditions of the apparatus
are similar to those of a typical air-conditioning application. The principal modifications from the standard refrigeration system are the addition of a two-phase ejector and
a liquid–vapor separator.
Refrigerant is discharged by a two-cylinder single stage
reciprocating unit, driven by an electric motor. The speed
of the motor is varied by an inverter to regulate the refrigerant flowing through the motive nozzle. Compact plate
heat exchangers are used for the condenser and evaporators. The evaporator referred to in this paper is the main
evaporator as shown in Fig. 1. A filter/drier, placed downstream of the receiver, is provided to keep the circulating
refrigerant free from harmful substances: moisture and foreign particles that might remain in the system. An oil separator is used to keep the oil content in the refrigerant to a
minimum.
The motive and the suction mass flow rates are measured by volumetric flow meters located downstream of
the sight glass and of the liquid-vapor separator, respectively. All flow meters are specially calibrated for R-134a
from the manufacturer. The total capacity of all refrigerant
flow meters is 0.3 to 3.3 LPM. The manufacturer’s listed
accuracy is 0.1% of the full scale. The temperatures are
measured by T-type thermocouples having accuracy of
0.1 °C. All the temperature-measuring devices are well calibrated in a controlled temperature bath using standard
precision mercury glass thermometers. Bourdon gauges,
calibrated against the dead weight test, are used to measure
the pressures. All static pressure taps are mounted flush in
Oil
Separator
Condenser
2′
Stirrer
RTD By-Pass
Valve
CDU
Electrical
Heater
Condensing
Unit
.
mp
1 Inverter
Sight
Glass
Cold-Water
Pump
Cold-Water
Tank
2
3
Flow
Meter
Compressor Motor
Filter-Drier
Receiver
Sight
Glass
Primary
Flow
Meter
.
mp
OC/CE
OE/CC
. .
m p+ ms
Ejector
OE/CC 3′
4
LiquidVapor
Separator
OE/CC
.
ms
Accumulator
OC/CE
6
.
ms
Expansion
Valve
Main
Evaporator
Secondary
Flow
Meter
OE/CC
5
Sight
Glass
Sub-Evaporator
Flow
Meter
Stirrer
RT
By-Pass
Valve
Hot-Water
Pump
Electrical
Heater
P
T
Condenser
Liquid-Vapor
Separator
4
.
mp
.
.
m p + ms
1
ne se me
Ejector
.
ms
6
: Pressure Gauge (High/Low)
: Thermocouple (T-Type)
2′
3
3′
Hot-Water
Tank
OE/CC : Open for Ejector Mode, Closed for
Conventional Mode
OC/CE : Open for Conventional Mode, Closed
for Ejector Mode
Evaporator
2
Compressor
5
Fig. 1. Schematic diagram of experimental apparatus.
the tube wall. Please note that the sub-evaporator shown in
Fig. 1 was not used in the present study. It was prepared
for the experiment with high cooling load. The main evaporator has enough capacity for the cooling load used in the
present study.
In order to collect data at various conditions, the heat
load to the evaporator is supplied by using the hot water
loop. The water in the tank is heated with a 4.5 kW electric
heater and supplied through the evaporator by the circulating pump. The condenser rejects heat to the water coming
from a cold water tank. The water is cooled by a separated
refrigeration system using R-22 as refrigerant. The separated refrigeration system consists of a condensing unit
with a capacity of approximately 2.6 tons of refrigeration,
Mixing Chamber
Motive Nozzle
Diffuser
Mixed
Refrigerant
Primary
Refrigerant
Suction Chamber
Secondary
Refrigerant
Nozzle Rod
O-Rings
φ22.0
Mixing Chamber and Diffuser
φ10.0
Motive Nozzle
φ30.0
Locking Nut
Suction Chamber
30.0
110.0
90.0
Fig. 2. Ejector assembly.
a capillary tube, filter/drier, sight glass and a helical copper
tube coil immersed in water being cooled in an insulated
tank. The test runs are done at the cooling load ranging
between 1.8 and 3.0 kW.
The schematic diagram of the two-phase ejector as
shown in Fig. 2, is designed in three main parts: the motive
nozzle, the suction chamber, and the mixing chamber with
diffuser. The motive nozzle throat area is designed according to the Henry and Fauske model [8]. This model is used
because it considers the metastable effect of the expansion
of saturated liquid into the liquid–vapor mixture region.
The remaining cross-sectional areas of the ejector are
designed according to the homogeneous equilibrium model
(HEM) (Kornhauser [1], Sherif et al. [9]). HEM is based on
the assumption that vapor and liquid are in thermal and
mechanical equilibrium. Furthermore, the mixing process
is assumed to occur at constant pressure. The other dimensions, including the lengths of each section and the convergent and divergent angles, are based on recommendations
from the ASHRAE Handbook [10] and from Nakagawa
and Takeuchi [5]. Brass is used as material for the ejector.
The three main parts of the ejector are connected by fine
screws. Three o-rings, as shown in Fig. 2, are used to prevent refrigerant leakage. The detail drawing of the tested
motive nozzle is shown in Fig. 3.
In the present study, three motive nozzles having different outlet diameter (Dne) of 2.0, 2.5 and 3.0 mm are investigated. The inlet diameter (Dni), inlet length (Lni),
convergent length (Lnc), throat diameter (Dnt) and diver-
gent length (Lnd) of all three motive nozzles are 6, 32, 6,
0.9, 20 mm, respectively. Usually, a comparison of the
cycle performance can be made by two approaches. The
first one is based on internal parameters i.e. evaporating
and condensing temperatures. This method requires the
different modes to be compared at the same evaporating
and condensing temperatures. The second one is based
on external parameters such as the inlet temperature and
the flow rate of the heat transfer fluid (HTF) (Hoegberg
et al. [11], Giuliani et al. [12]). This method allows each
mode of operation to operate under the same external conditions. In this paper, the comparison is based on the second method. This method is selected because it is more
likely in real life. Water is used as the heat transfer fluid.
Hot water acts as a heat source of the evaporator and heat
source temperature means temperature of hot water. While
cold water acts as a heat sink of the condenser and heat
sink temperature means temperature of cold water.
According to the experimental conditions, the test runs
are done at heat source temperatures ranging between
8 °C and 16 °C while the volume flow rate of hot water is
kept constant at 12 LPM. Each value of heat source temperature is tested at varying heat sink temperatures of
26.5, 29.5, 32.5, 35.5 and 38.5 °C. The volume flow rate
of cold water is fixed at 14 LPM. These volumetric flow
rates of cold water and hot water are employed to prevent
water from freezing on the surface of the equipment. The
compressor speed is maintained at 450 rpm by controlling
the inverter frequency. It should be noted that this speed
Dnt
Dne
Dni
Lni
Lnc
Lnd
Fig. 3. Motive nozzle.
is appropriate for this experimental set up. This speed is
employed after problems were encountered at several different speeds used in previous experiments. It has been
found that if compressor speeds of over 450 rpm are used,
the pressure in the liquid–vapor separator decreases. This
enables the liquid in the separator to become increasingly
vaporized and finally result in the increase of the vapor
temperature. This high temperature vapor probably causes
the compressor to become damaged. Also if compressor
speeds of lower than 450 rpm are used, the amount of
liquid in the separator gradually increases. Finally, the
liquid refrigerant floods the outlet of the separator and
flows to the compressor. This results in compressor failure.
The water temperatures are kept constant at the required
values in both the hot and the cold water tanks. They are
then circulated through the condenser and evaporator at
constant temperatures and constant volume flow rates by
the circulating pumps.
3. Results and discussion
Fig. 4 shows a pressure-enthalpy diagram of the
two-phase ejector refrigeration cycle for Dne = 2.5 mm, at
Tsource = 8 °C and Tsink = 26.50 °C. Fig. 4 is only an example, however the experimental results from other experimental conditions give the same tendency. The cycle is
separated into two loops; primary refrigerant loop and secondary refrigerant loop. In the primary refrigerant loop,
vapor from the liquid–vapor separator is drawn in the compressor cylinder during its suction stroke and is compressed
to pressure p2 and temperature T2 during the compression
stroke and delivered out from the compressor at condition
2 passes on to condenser at 2 0 in which cooling water is
supplied to remove heat from the vapor. Vapor is therefore
first cooled to the saturated temperature at pressure p2 and
further removal of heat. Condensation at the high temperature produces a sub-cooled liquid refrigerant at point 3.
The high pressure liquid is now further expanded through
motive nozzle at point 3 0 . The primary refrigerant will be
mixed with the secondary refrigerant at mixing chamber.
The mixture will be compressed through diffuser and flow
to point 4 and to the liquid–vapor separator. In the secondary refrigerant loop, since high pressure primary refrigerant
is supplied to the nozzle inlet and is expanded in the mixing
chamber, refrigerant vapor originating from the evaporator at point 6, is entrained with the high velocity refrigerant
jet and compressed through mixing chamber into the diffuser at point 4. Vapor refrigerant from the liquid–vapor
separator is sucked by compressor while sub-cooled liquid
passes on to the inlet of the evaporator at point 5 and
absorbs heat from hot water supplied to the evaporator.
The outlet of the evaporator appears a point 6. The cycle
continues on and on.
Motive nozzles used in the present study have different
sizes of outlet diameter. This different outlet diameter
directly affects the shape of the divergent angle of the
motive nozzle. The following results are gained from experimental conditions in which the refrigerant at the entries of
the motive nozzle and the evaporator are sub-cooled with
temperatures of 0.1–3.5 °C and 0.1–1.1 °C, respectively,
while the refrigerant at the inlet of the compressor is superheated with the temperature of 2.0–4.5 °C. The experiment
yields similar variations of most variables affecting the performance of the cycle, with significant differences in certain
conditions. However, the variations of variables tested with
this range of outlet diameter are not favorably obvious.
Fig. 5 shows the variation of the primary mass flow rate
with heat sink temperature from using the various sizes of
motive nozzle for the different heat source temperatures of
8, 12 and 16 °C. The graph indicates that an increase or
decrease of the heat source temperature does not affect
the primary mass flow rate, even though it has direct effect
on the increase or decrease of temperature and pressure in
the evaporator. This varying pressure has very little effect
on the pressure difference between the inlet and the discharge of the motive nozzle. Therefore, the variation of
the heat source temperature has no significant effect on
the primary mass flow rate.
Concerning the variation of the primary mass flow rate
with heat sink temperature at heat source temperatures of
8, 12 and 16 °C, as shown in Fig. 5, it is found that the primary mass flow rate tends to increase when the heat sink
temperature increases. This is because the temperature
and pressure of the condenser increase with increasing heat
sink temperature. This higher condenser pressure results in
the increase of the pressure at the inlet of the motive nozzle.
As a result, the pressure difference between the inlet and the
outlet of the motive nozzle also increases which, in turn,
10.0
10.0
Motive Nozzle Dne = 2.5mm.
TSource = 8oC
Pressure (MPa)
TSink = 26.5oC
1.0
2
3
1.0
2'
3'
4
5
1
6
Ideal
Actual
ne
me
se
0.1
0.1
150
200
250
300
350
400
450
500
Enthalpy (kJ/kg)
Fig. 4. Pressure–enthalpy diagram of the two-phase ejector refrigeration cycle in the present study.
Primary Mass Flow Rate (kg/s)
.04
.03
Motive Nozzle :
Dne = 3.0 mm.
Dne = 2.5 mm.
Dne = 2.0 mm.
TSource = 8 oC
TSource =12 oC
TSource =16 oC
Subcooling at the Inlet :
- Motive Nozzle : 0.1-3.5 oC
- Evaporator
: 0.1-1.1 oC
.02
.01
0.00
24
26
28
30
32
34
36
38
40
Heat Sink Temperature (οC)
Fig. 5. Comparison of primary mass flow rates of refrigerant at heat source temperatures of 8, 12 and 16 °C.
increases the primary mass flow rate. When comparing the
primary mass flow rate gained from the three motive nozzles, it is found that all motive nozzles give similar primary
mass flow rate.
Fig. 6 shows the comparison of the secondary mass flow
rate gained from using various sizes of motive nozzle at the
heat source temperatures of 8 and 16 °C. The graph shows
that the change of the heat source temperature has a small
Secondary Mass Flow Rate (kg/s)
.04
Subcooling at the Inlet :
- Motive Nozzle : 0.1-3.5 oC
- Evaporator
: 0.1-1.1 oC
.03
.02
.01
24
Motive Nozzle :
Dne = 3.0 mm.
Dne = 2.5 mm.
Dne = 2.0 mm.
26
28
30
32
TSource = 8 oC
TSource =16 oC
34
36
38
40
Heat Sink Temperature (οC)
Fig. 6. Comparison of secondary mass flow rates of refrigerant at heat source temperatures of 8 and 16 °C.
effect on the secondary mass flow rate. In addition, the
results suggested that the secondary mass flow rate depends
mainly on the primary mass flow rate and the pressure at
the outlet of the motive nozzle.
The comparison of the secondary mass flow rates at
Tsource = 8 °C, reveals that motive nozzle having
Dne = 3.0 mm produced a slightly higher secondary mass
flow rate than the other nozzles, as more clearly evident
at high heat sink temperatures. Concerning the changes
in the secondary mass flow rate and heat sink temperature
when using 3 different sizes of motive nozzle, it is found
that the secondary mass flow rate increases when the heat
sink temperature increases. This is because the higher heat
sink temperature makes the pressure difference between the
inlet and the outlet of motive nozzle higher, resulting in a
higher velocity of the primary mass flow rate that causes
the pressure to drop at the outlet of the motive nozzle. This
pressure drop at the outlet of the motive nozzle increases
the pressure difference between the outlet of the motive
nozzle and the outlet of the evaporator. Hence, the secondary mass flow rate increases accordingly.
The recirculation ratio is defined as the ratio between
the refrigerant mass flow rate delivered to the evaporator
ðm_ s Þ and flow rate of the vaporized refrigerant ðm_ e Þ.
Fig. 7 shows the variation of the recirculation ratio with
heat sink temperature from using the various sizes of
motive nozzle at heat source temperatures of 8 and
16 °C. The result shows that the recirculation ratio
decreases with an increase in heat source temperature. This
is because the heat source temperature increases with
increasing rate of heat transfer at the evaporator, which
results in the increasing vaporized mass flow rate, while
the secondary mass flow rate changes slightly. Ultimately,
the recirculation ratio also increases.
The comparison of the recirculation ratio when using
motive nozzles at heat source temperatures of 8 and
16 °C as shown in Fig. 7, indicates that motive nozzle having Dne = 3.0 mm produces the higher recirculation ratio
than the others. This is because the recirculation ratio is
the ratio of the secondary mass flow rate and vaporized
mass flow rate in the evaporator obtained from the calculation. The calculation result indicates that the vaporized
mass flow rate is lowest when using motive having
Dne = 3.0 mm. Considering the change in the recirculation
ratio and heat sink temperature gained from the three different sizes of motive nozzle, it is found that the recirculation ratio tends to increase when the heat sink temperature
increases. This is due to the ejector behavior. That is
because the secondary mass flow rate increases with
increasing heat sink temperature, while the vaporized mass
flow rate drops. The higher heat sink temperature causes
the heat transfer at the evaporator to drop.
Fig. 8 shows the variation in the average evaporator
pressure as the heat sink temperature changes, from using
the various sizes of motive nozzle at heat source temperatures of 8, 12 and 16 °C. The graph indicates that the
higher heat source temperature gives higher average evaporator pressure. This is because the change in the heat
source temperature directly affects the changes of temperature and pressure in the evaporator.
The changes in the average evaporator pressure and heat
sink temperature as shown in Fig. 8, indicate that when the
heat sink temperature increases, the average evaporator
pressure will also increase slightly – even though the
increase or decrease of the heat sink temperature does
not have a direct effect on the change of pressure in the
evaporator. The increase of the average evaporator pressure may result from other causes, such as, the heat transfer
5.0
Recirculation Ratio
4.0
Subcooling at the Inlet :
- Motive Nozzle : 0.1-3.5 oC
- Evaporator
: 0.1-1.1 oC
3.0
2.0
Motive Nozzle :
Dne = 3.0 mm.
Dne = 2.5 mm.
Dne = 2.0 mm.
1.0
0.0
24
26
28
30
32
TSource = 8 oC
TSource =16 oC
34
36
38
40
Heat Sink Temperature (οC)
Fig. 7. Comparison of recirculation ratio at heat source temperatures of 8 and 16 °C.
0.6
Average Evaporator Pressure (MPa)
Superheating at the Compressor Inlet : 2.0-4.5 oC
0.5
0.4
0.3
Motive Nozzle :
Dne = 3.0 mm.
Dne = 2.5 mm.
Dne = 2.0 mm.
0.2
0.1
24
26
28
30
32
TSource = 8 oC
TSource =12 oC
TSource =16 oC
34
36
38
40
Heat Sink Temperature (οC)
Fig. 8. Comparison of average evaporator pressure at heat source temperatures of 8, 12 and 16 °C.
from surrounding into the system. When comparing the
average evaporator pressures gained from the three
different sizes of motive nozzle, it is found that all motive
nozzles produce almost the same average evaporator
pressure.
Fig. 9 shows the relation between the compressor pressure ratio and the heat source temperature from using the
various sizes of motive nozzle at the heat sink temperatures
of 26.5, 32.5 and 38.5 °C. It is clear that the compressor
pressure ratio increases with increasing heat sink tempera-
ture. This is because the heat sink temperature has a direct
effect on the temperature and pressure in the condenser.
When the condenser pressure increases, the pressure at
the condenser outlet also increases. This results in the
increase of the compressor pressure ratio. The compressor
pressure ratio for specific heat source and heat sink temperatures obtained from all motive nozzles is similar.
Fig. 10 shows the discharge temperature plotted against
heat source temperature for the various sizes of motive
nozzle at heat sink temperatures of 26.5, 32.5 and
5.0
Superheating at the Compressor Inlet : 2.0-4.5 oC
Compressor Pressure Ratio
4.0
3.0
2.0
Motive Nozzle :
Dne = 3.0 mm.
Dne = 2.5 mm.
Dne = 2.0 mm.
1.0
TSink = 26.5 oC
TSink = 32.5 oC
TSink = 38.5 oC
0.0
8.0
10.0
12.0
14.0
16.0
ο
Heat Source Temperature ( C)
Fig. 9. Comparison of compressor pressure ratio at heat sink temperatures of 26.5, 32.5 and 38.5 °C.
80
Superheating at the Compressor Inlet : 2.0-4.5 oC
Discharge Temperature (oC)
70
60
50
40
Motive Nozzle :
Dne = 3.0 mm.
Dne = 2.5 mm.
Dne = 2.0 mm.
30
TSink = 26.5 oC
TSink = 32.5 oC
TSink = 38.5 oC
20
8.0
10.0
12.0
14.0
16.0
ο
Heat Source Temperature ( C)
Fig. 10. Comparison of discharge temperature at heat sink temperatures of 26.5, 32.5 and 38.5 °C.
38.5 °C. The plot shows that the increase of heat sink temperature makes the discharge temperature rise accordingly.
This is because when the heat sink temperature increases,
the compressor pressure ratio will also increase and, consequently, the compressor temperature ratio defined as the
ratio between the temperature at the compressor outlet
and the temperature at the compressor inlet, will increase
accordingly.
The result from this figure indicates that the discharge
temperature tends to decrease when the heat source temperature increases. The discharge temperatures gained
from all motive nozzles for specific heat source and heat
sink temperatures are similar.
The cooling capacity can be calculated according to the
following equation:
Qevap ¼ m_ hw cp;hw ðT hw;in T hw;out Þ
where m_ hw is the mass flow rate of hot water (kg/s), cp, hw is
the specific heat at constant pressure of the hot water (kJ/
kg k), Thw, in is the hot water temperature at the inlet of the
evaporator (°C), Thw, out is the hot water temperature at the
outlet of the evaporator (°C).
4.0
Superheating at the Compressor Inlet : 2.0-4.5 oC
Cooling Capacity (kW)
3.0
2.0
Motive Nozzle :
Dne = 3.0 mm.
Dne = 2.5 mm.
Dne = 2.0 mm.
1.0
TSink = 26.5 oC
TSink = 38.5 oC
0.0
8.0
10.0
12.0
14.0
16.0
Heat Source Temperature (οC)
Fig. 11. Comparison of cooling capacity at heat sink temperatures of 26.5 and 38.5 °C.
Fig. 11 shows comparison of the cooling capacity
obtained from all motive nozzles at heat sink temperatures
of 26.5 and 38.5 °C. It can be clearly seen that the cooling
capacity decreases with increasing heat sink temperature.
As a consequence of ejector operation, the higher heat sink
temperature causes the temperature difference between the
cold water from the heat sink and the refrigerant from the
condenser to decrease. As a result, the heat transfer rate of
the condenser decreases accordingly. Considering the
energy conservation, it is found that when the heat transfer
rate of the condenser decreases, the heat transfer rate of the
evaporator will also decrease accordingly.
The changes in the cooling capacity and heat source
temperature as shown in Fig. 11, indicate that the cooling
capacity increases as the heat source temperature increases.
The change of the heat source temperature is similar to the
change in the cooling load of the system. When the
temperature of the hot water from the heat source
increases, the heat absorption of the refrigerant in the
evaporator also increases. The amount of heat absorbed
by the refrigerant is actually the average heat transfer rate
of the evaporator, which equals the cooling capacity of the
system. The comparison of the cooling capacity gained
from using the 3 different outlet diameter of motive nozzle
7.0
Superheating at the Compressor Inlet : 2.0-4.5 oC
Coefficient of Performance (COP)
6.0
5.0
4.0
3.0
Motive Nozzle :
Dne = 3.0 mm.
Dne = 2.5 mm.
Dne = 2.0 mm.
2.0
1.0
TSink = 26.5 oC
TSink = 32.5 oC
TSink = 38.5 oC
0.0
8.0
10.0
12.0
14.0
16.0
Heat Source Temperature (οC)
Fig. 12. Comparison of coefficient of performance at heat sink temperatures of 26.5, 32.5 and 38.5 °C.
reveals that there is no significant effect on the cooling
capacity.
The system COP is defined as the ratio between the cooling capacity and the electrical power supplied to the compressor and can be written as
COP ¼
Qevap
W comp
where Wcomp is the electrical power supplied to the compressor (kW). This electrical power is directly obtained
from the buit-in function of the inverter.
Fig. 12 shows the variation of the system’s COP versus
heat source temperature from using the various sizes of
motive nozzle at heat sink temperatures of 26.5, 32.5 and
38.5 °C. It can be clearly seen that the COP decreases as
the heat sink temperature increases. This is because when
the heat sink temperature increases, the cooling capacity
of the system decreases, while the compressor pressure
ratio increases, resulting in an increase of electrical power
input of the compressor. Ultimately, this results in the
decrease of the COP of the system. Comparison of the system’s COP, at heat sink temperature of 26.5, 32.5 and
38.5 °C, all motive nozzles tend to give similar COP.
4. Conclusions
Variables affecting performance and varying directly
with the heat sink temperature include the primary mass
flow rate, secondary mass flow rate, recirculation ratio,
compressor pressure ratio, and discharge temperature. On
the other hand, the cooling capacity varies inversely with
the heat sink temperature while the average evaporator
pressure varies only a little and tends to vary directly. Concerning the effect of the heat source temperature on the variation of variables affecting performance, the results are
found to be contrary with the effect of the heat sink temperature, with the exception of the average evaporator pressure which varies directly. On the other hand, the
primary mass flow rate and the secondary mass flow rate
tend to be slightly increased as the heat source temperature
increases. The use of motive nozzles having different outlet
diameters in the range of 2.0–3.0 mm yields insignificant
effects on the system performance. However, although the
ejector cycle has higher performance over the standard
cycle, some disadvantage should be considered e.g. high
refrigerant charge, high refrigerant flow, piping insulation
and installation cost.
Acknowledgements
The authors would like to express their appreciation to
the Thailand Research Fund (TRF) for providing financial
support in this study.
References
[1] A.A. Kornhauser, The use of an ejector as a refrigerant expander, in:
Proceedings of the 1990 USNC/IIR-Purdue Refrigeration Conference, 1990; 10–19.
[2] G.S. Harrell, A.A. Kornhauser, Performance tests of a two-phase
ejector, in: Proceedings of the 30th Intersociety Energy Conversion
Engineering Conference, Orlando, FL, 1995; 49–53.
[3] P. Menegay A.A. Kornhauser, Improvements to the ejector expansion refrigeration cycle, in: Proceedings of the 31th Intersociety
Energy Conversion Engineering Conference, Washington DC, 1996;
702-6.
[4] P.A. Domanski, Theoretical evaluation of the vapor compression
cycle with a liquid-line/suction-line heat exchanger, economizer, and
ejector. Nistir-5606, National Institute of Standards and Technology,
March, 1995.
[5] M. Nakagawa, H. Takeuchi, Performance of two-phase ejector in
refrigeration cycle, in: Proceedings of the Third International Conference on Multiphase Flow. Lyon. France, Jun. 8–12, 1998; 1-8.
[6] S. Disawas, S. Wongwises, Experimental investigation on the
performance of the refrigeration cycle using a two-phase ejector as
an expansion device, Int. J. Refrigeration 27 (6) (2004) 587–594.
[7] S. Wongwises, S. Disawas, Performance of the two-phase ejector
expansion refrigeration cycle, Int. J. Heat Mass Transf. 48 (2005)
4282–4286.
[8] R.E. Henry, H.K. Fauske, The two-phase critical flow of onecomponent mixtures in nozzles, orifices, and short tubes. ASME
Trans. J. Heat Transf. May, 1971; 179–87.
[9] S.A. Sherif, W.E. Lear, J.M. Steadham, P.L. Hunt, J.B. Holladay,
Analysis and modeling of a two-phase jet pump of a thermal
management system for aerospace applications, Int. J. Mech. Sci. 42
(2000) 185–198.
[10] ASHRAE, ASHRAE Handbook – Guide and Data Book. American
Society of Heating, Refrigerating and Air Conditioning Engineering.
1969; Chapter 13, 151–58.
[11] M. Hoegberg, L. Vamling, T. Berntsson, Calculation methods for
comparing the performance of pure and mixed working fluids in heat
pump applications, Int. J. Refrigeration 16 (6) (1993) 403–413.
[12] G. Giuliani, N.J. Hewitt, F. Marchesi Donati, F. Polonara, Composition shift in liquid-recirculation refrigeration systems: an experimental investigation for the pure fluid R134a and the mixture R32/
134a, Int. J. Refrigeration 22 (6) (1999) 486–498.