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Experimental results with a variable geometry ejector using R600a as working fluid
Paulo R. Pereira, Szabolcs Varga, João Soares, Armando C. Oliveira, António M.Lopes, Fernando G. de Almeida, João F. Carneiro
PII: S0140-7007(14)00166-2
DOI: 10.1016/j.ijrefrig.2014.06.016
Reference: JIJR 2822
To appear in: International Journal of Refrigeration
Received Date: 25 March 2014
Revised Date: 11 June 2014
Accepted Date: 29 June 2014
Please cite this article as: Pereira, P.R., Varga, S., Soares, J., Oliveira, A.C., Lopes, A.M., de Almeida,F.G., Carneiro, J.F., Experimental results with a variable geometry ejector using R600a as working fluid,International Journal of Refrigeration (2014), doi: 10.1016/j.ijrefrig.2014.06.016.
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Experimental results with a variable geometry ejector
using R600a as working fluid
Résultats expérimentaux avec un éjecteur à géométrie variable utilisant R600a comme réfrigèrent
Paulo R. Pereira, Szabolcs Varga*, João Soares, Armando C. Oliveira, António M.
Lopes, Fernando G. de Almeida, João F. Carneiro
Department of Mechanical Engineering
Faculty of Engineering, University of Porto
Rua Dr Roberto Frias, 4200-465 Porto – Portugal
* corresponding author
e-mail: [email protected]
Phone: +351 220414366
Fax: +351 225082153
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Abstract
Experimental results with the first laboratory scale variable geometry ejector (VGE)
using isobutane (R600a) are presented. Two geometrical factors, the area ratio and the
nozzle exit position, can be actively controlled. The control of the area ratio is achieved
by a movable spindle installed in the primary nozzle. The influence of the spindle
position (SP) and condenser pressure on ejector performance are studied. The results
indicate very good ejector performance for a generator and evaporator temperature of 83
ºC and 9 ºC, respectively. COP varied between 0.4 and 0.8, depending on operating
conditions. The existence of an optimal SP, depending on the back pressure, is
identified. A comparison of the benefit of applying the variable geometry design over a
fixed geometry configuration is assessed. For example, for a condenser pressure of 3
bar, an 80% increase in the COP was obtained when compared to the performance of a
fixed geometry ejector.
Keywords: Ejector cooling system; Variable geometry design; Experimental work; Performance enhancement.
Mots-clés: Système de refroidissement par éjecteur; Conception à géométrie variable;
Travaux expérimentaux; Amélioration de la performance.
Symbols
COP coefficient of performance
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diameter (m)
mass flow rate (kg s-1)
NXP nozzle exit position (mm)
pressure (bar)
heat (kW)
area ratio,
temperature (°C)
Greek letters
λ entrainment ratio
Subscripts
condenser
critical
diffuser
evaporator
heat generator
constant area section
primary nozzle throat
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primary nozzle exit
1. Introduction
The increased demand on human thermal comfort has led to a massive growth of air
conditioning applications over the last two decades, in both developed and developing
countries. In order to counteract its negative effect on electricity consumption, the
development of efficient and reliable thermally driven cooling systems seems to be a
logical solution, especially in the context of supplying the necessary heat from already
available solar collectors. Ejector cooling fits well into these requirements, since
ejectors are simple in construction, they have long durability and require little
maintenance. This has been realised by many researchers, and a significant effort has
been dedicated to improve ejector cooling efficiencies since the mid-nineties. A recent
review on the advances of ejector technology can be found in Chen et al. (2013).
Such as most thermally driven cooling cycles, ejector refrigeration is strongly
influenced by operating conditions and the properties of the working fluid. The
influence of working temperatures and pressures on ejector cooling cycle performance
is relatively well established. Several experimental (e.g. Chunnanond and Aphornratana,
2004b; Selvaraju and Mani, 2006; Yapıcı et al., 2008) and theoretical works have been
carried out (e.g. Hemidi et al., 2009; Huang et al., 1999; Varga et al., 2009a) to assess
the effect of generator, evaporator and condenser conditions. Comparative studies of
cooling performance using different refrigerants are limited to theoretical analyses, such
as in (Sun, 1999) and (Cizungu et al., 2001). More recently, studies (Roman and
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Hernandez, 2011; Varga et al., 2013b) that involved the analysis of hydrocarbon
refrigerants, for ejector cooling, such as R600a, concluded that they provide high
performance; however, precautions should be taken into consideration due to their
flammability.
Ejector design, for a given working fluid and cooling capacity, has a strong dependence
on the operating conditions (Varga et al., 2009b; Yapıcı et al., 2008). In other words, an
ejector with fixed geometry only works with high COP in a narrow range of operating
temperatures/pressures, called design conditions. Operating conditions influence ejector
geometry mostly through the area ratio, rA (ratio between the constant area section and
primary nozzle throat area). A potential solution to this problem has led to the concept
of a variable geometry ejector (VGE). Sun (1996) was one of the first authors analysing
geometrical requirements for an ejector using water as working fluid. Recently, Dennis
and Garzoli (2011) presented research results of a VGE using R141b under variable
operating conditions. In both cases the advantages of the concept were clearly
demonstrated; however, no technical solutions were given. Potential technical
implementations for controlling the area ratio can be found in Kim et al. (2006) for air;
in Elbel and Hrnjak (2008) for carbon dioxide (R744); and in Ma et al. (2010) and
Varga et al. (2011) for water as the working fluid. The use of air as a working fluid has
a very limited application in refrigeration. R744 has a very low critical temperature and
high critical pressure; therefore, its use requires adequate technical features and robust
construction that would contribute to an increase in the initial cost of the cooling
system. Water has been used widely as a refrigerant in ejector cooling; however, it has
the disadvantage of resulting in relatively low COPs for moderate generator
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temperatures, which is the case when using low-cost solar collectors as primary heat
source (Varga et al., 2013b).
Numerical studies with hydrocarbon refrigerants (e.g. Roman and Hernandez, 2011),
such as R600a (e.g. Pridasawas and Lundqvist, 2007), have shown that a good cooling
cycle performance can be achieved, even for moderate generator temperatures (80 – 90
ºC). Although R600a is already widely used in domestic refrigerators, because of the
flammable nature of R600a there is only a very limited number of experimental
investigations regarding ejector cooling. According to the authors’ knowledge, until this
date the only experimental investigations with R600a in an ejector cooling system were
presented by Butrymowicz et al. (2012) and Butrymowicz et al. (2013). In these works
relatively low COPs were obtained (~0.15); however, the generator temperatures
applied were also very low (63.5 ºC). In a previous study, Varga et al. (2013a)
developed a CFD model to assess the benefit of a VGE design for an R600a ejector.
Predicted entrainment ratios varied in a wide range (0.15 – 0.65) depending on the
operating conditions. It was concluded that installing a spindle in the primary nozzle
could lead to increased ejector performance, with an increase as high as 177% for low
condenser pressures, when compared to a fixed geometry design. Based on these
findings, an R600a ejector prototype with variable geometry was developed and
installed in a test rig. The ejector geometry can be actively controlled by changing the
area ratio through a movable spindle and by changing the nozzle exit position. The
present work summarises the first results obtained with the experimental setup. The
benefit of applying a variable geometry over a fixed geometry is experimentally
demonstrated for the first time.
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2. Experimental Setup
The experimental ejector test rig is composed by: i) the ejector refrigeration cycle with
the variable geometry ejector; plus three sub-cycles, which are: ii) one on the high
temperature side connected to the generator; iii) one on the low temperature side
connected to the evaporator; and iv) one on the heat dissipation side connected to the
ejector cycle condenser, as shown in Fig. 1. The objective of the three sub-cycles (ii-iv)
was to simulate operational conditions, so that the heat transfer in the generator,
evaporator and condenser can be independently controlled.
In the ejector cycle, the motive (primary) fluid is pressurized according to the desired
saturation temperature by using a magnetic drive sliding vane pump (Gemmecotti,
Italy). The desired pressure was set by adjusting the pump velocity with a general
purpose frequency inverter. Upstream to the fluid circulating pump (after the condenser)
a vapor separator was installed in order to assure that the working fluid enters on the
suction side in liquid state. In addition, the pump was placed at the lowest location of
the rig in order to prevent cavitation problems. From the pump, the high pressure
motive fluid enters the generator that is connected to the external heat source (electric
heater). The generator is composed by two plate heat exchangers (Alfa Laval, Sweden).
In the first one, the motive fluid receives energy mostly in the form of sensible heat,
while it leaves the second heat exchanger with a small amount of superheat (~ 5 - 15
ºC). The primary fluid is led to the variable geometry ejector where it mixes with the
secondary stream coming from the evaporator. A more detailed discussion of the VGE
is presented in Section 3. The mixed R600a vapour leaving the ejector is condensed in
another plate heat exchanger (Alfa Laval, Sweden) connected to a water chiller. On the
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low pressure side of the ejector, there is an expansion valve (Swagelok, USA), that is
manually adjusted to the desired evaporator pressure. The cooling effect is obtained in
the evaporator also with a plate type configuration (Alfa Laval, Sweden). All the
installation, including connecting pipes was supplied with thermal insulation (Armacell,
Germany), in order to minimize heat losses.
The high temperature heat source, supplying the thermal energy for the generator of the
cooling cycle, was a 6 kW electric temperature controller (SHINI, Taiwan). It can
produce high temperature pressurized water up to 120 °C. A thermal bath (JULABO,
Germany), with a heating capacity of 2 kW was used to simulate the thermal load in the
evaporator. The condenser was connected to a water chiller of about 1.4 kW cooling
capacity. The water chiller is equipped with a cold water tank of approximately 150 l to
provide steady state operation for a given period of time.
In order to monitor the system variables along the ejector cycle, a number of
instruments were installed including: pressure transducers (Kobold, Germany) with 6
and 25 bar measuring range and an accuracy of 0.5% of the full scale; calibrated T-type
thermocouples (Tecnisis, Portugal) with a maximum error of 0.35 °C; RTDs (KIMO,
France) with 0 to 100 °C range and an accuracy of ±0.08 °C; and two variable area flow
meters (Kobold, Germany) with an accuracy of 2.2% of the full scale. Additionally, the
water temperature at the inlet and outlet of both the generator and the evaporator sub-
cycles were measured in order to evaluate the global performance of the system. The
water flow rate through the generator was monitored with a high precision
electromagnetic flow meter (ABB, USA) with an accuracy of 0.5% of the read value.
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On the evaporator side, the water flow rate was constant (13.3 l/min) and it was
manually determined by measuring the displaced water volume with a chronometer
(estimated error of 3%). For the locations of the installed sensors the reader is referred
to Fig. 1. The data acquisition system integrated a data logger module (HP Agilent
34970A, USA) connected to a personal computer. A control and monitoring application
was developed in LabVIEW 2011 (National Instruments, USA). The data were sampled
every 10 s and saved in a text file for data analysis, carried out in MS Excel and
Engineering Equation Solver (EES) (F-Chart, USA). A general overview of the rig with
its key components is presented in Fig. 2.
3 Ejector Operation and VGE design
The key component of the experimental setup, and the heart of the refrigeration cycle, is
the ejector. The schematic cross section of a typical ejector is shown in Fig. 3. The high
pressure stream (motive or primary fluid) coming from the generator (g) enters the
primary nozzle at low velocity (Varga et al., 2013a). Because of the converging section,
the primary flow gets accelerated and choked in the nozzle throat (dt). In the divergent
section, the stream is expanded such that it fans out from the primary nozzle to the
suction chamber typically with supersonic speed and low static pressure (Chunnanond
and Aphornratana, 2004a; Varga et al., 2013b). At this condition there is a difference
between the pressure at the mixing chamber and the evaporator (Chen et al., 2014), the
former being lower. This pressure difference draws the secondary fluid from the
evaporator (e) into the mixing chamber. Due to the velocity difference between the two
fluids and the resulting shear, the secondary stream gets accelerated. Under normal
conditions, the secondary fluid starts mixing with the primary fluid after it gets choked.
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Due to the interaction of both fluid streams and the ejector wall, the mixing process
downstream the nozzle exit plane is rather complex. The flow of the primary jet can be
characterized by a series of oblique/normal shock waves, known as the shock train
(Bartosiewicz et al., 2006; Bouhanguel et al., 2011). During this process, the static
pressure of the motive stream gradually increases and levels with the pressure of the
secondary fluid. Somewhere in the constant area section, or in the beginning of the
diffuser (depending on operation conditions), the mixing process completes and a final
shock occurs. From there, the flow essentially becomes subsonic and thus the increasing
cross section of the diffuser leads to deceleration of the mixed streams and to a
simultaneous increase of the static pressure. The exit pressure is governed by the
condenser (c) conditions (ejector back pressure/temperature).
Entrainment ratio (λ) and cooling cycle COP are the most important global performance
indicators for characterising an ejector. The entrainment ratio is defined as:
(1)
In the experimental test rig, the flow meters measured the volumetric flow rates of the
secondary and the mixed streams at evaporator outlet and at condenser inlet,
respectively (see Fig. 1). The conversion to mass flow rate was carried out by
determining the fluid density, based on the pressure and temperature readings at the
same locations, and using the property functions of EES. Additionally, the primary mass
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flow rate was obtained through the difference between mass flow rates of the mixed
stream and the secondary fluid. Once λ is determined, COP was obtained by:
(2)
In Eq. 2 the enthalpy change in the evaporator (∆he) and in the generator (∆hg) were
determined using EES physical property functions with the measured pressure and
temperature data. Note that in Eq. 2, it is assumed that the pump energy consumption is
negligible.
As all thermally driven cooling technologies, the ejector cycle has a strong dependence
between its performance and the operational conditions, i.e. the generator, evaporator
and condenser temperatures (Varga et al., 2013b). This is particularly important for the
generator temperature in the context of solar driven ejector cooling systems, where
temperature oscillations are expected during operation. For instance, when a fixed
geometry ejector is designed to operate at Tg=90 °C, for any higher value, the primary
flow rate increases, leading to a decrease in the entrainment ratio. This decrease is a
combined effect of the increased primary mass flow (larger heat input), and also the fact
that the primary stream leaves the nozzle slightly under-expanded, which results in
decreasing the secondary mass flow (reduced cooling). In contrast, a decrease in the
generator temperature results in insufficient momentum transfer, and thus mixing
between the two streams, leading to poorer ejector performance (Yapıcı et al., 2008).
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Regarding the condenser temperature (pressure), the performance of the ejector cooling
system is limited by the critical back pressure (pc,cr) (Chunnanond and Aphornratana,
2004a). Fig. 4 presents a typical operating curve of an ejector where the entrainment
ratio is depicted with the condenser pressure, at specified generator and evaporator
temperatures. It is possible to distinguish two operating regimes depending on the
condenser pressure. In the double choking region, the entrainment ratio is constant and
independent from condenser pressure. This can be explained by the choking of the
secondary flow in the ejector tail (Varga et al., 2013a), namely that when a gas flow is
chocked, the mass flow rate is only dependent on the upstream condition. In contrast, in
the single choking region the secondary flow remains subsonic, its mass flow rate
depends on both upstream and downstream conditions and thus λ quickly falls with pc.
The boundary between the two regions is defined as the critical condenser pressure
(back pressure) and it can be considered as another performance parameter. The higher
pc,cr is, the wider the range of condenser pressures at which the ejector operates with
constant performance. Since beyond pc,cr, the secondary flow is not choked, ejector
operation depends also on downstream conditions, and thus the pressure in the mixing
chamber increases with condenser pressure. Eventually, it can be higher than evaporator
pressure, leading to a reverse flow on the secondary side of the ejector, and thus ejector
failure (see Fig. 4) (Allouche et al., 2014). Regarding the evaporator temperature
(pressure), an increase in pe results in higher secondary mass flow rate and consequently
in an increase of both entrainment ratio and COP. It should be noted that the critical
condenser pressure also increases with evaporator pressure (Chunnanond and
Aphornratana, 2004a).
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Considering the benefits presented in the introduction, a variable geometry ejector
prototype was developed. The variable geometry is achieved by active control of: i) the
nozzle exit position (NXP) in the converging part of the ejector tail; and ii) the area
ratio, with a movable spindle that changes the primary nozzle cross section on the high
pressure side of the nozzle throat. These two degrees of freedom can be independently
set, allowing a full adaptation to the operating conditions. The desired position of the
spindle (SP) and NXP can be adjusted by two actuators driven by small stepper motors.
Fig. 5 presents a schematic view of the flow channel of the ejector prototype with its
characteristic dimensions, including NXP and SP. As can be seen from Fig. 5, the
spindle can travel forward to a position where it completely blocks the free passage for
the primary fluid, by touching the primary nozzle wall. This position was considered the
zero position. Thus, each SP (in mm) was referred to this zero position when moving
the spindle upstream. Likewise, the nozzle exit can be moved upstream into the mixing
chamber, up to a point when there is no free cross section available for the secondary
flow. This was considered again as zero for the NXP. A given position is measured
upstream from this zero position (in mm).
4. Experimental Procedure and Data Analysis
The main objective of the present experimental approach is to demonstrate the
performance benefits of the variable geometry ejector prototype, based on operating
conditions and spindle position. In this preliminary work, a series of experiments were
carried out by using constant generator and evaporator conditions, with controlled
temperatures of 85 °C and 15 °C in the electric heater and the thermal bath respectively,
for several spindle positions. Due to some limitations inherent to the capacity of the
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pump and the high temperature heat source, it was only possible to collect experimental
data for a range of spindle positions between 3.75 to 6.25 mm, representing an area ratio
variation ranging from 9.6 to 5.4. During these tests, the nozzle exit position has been
fixed at 3 mm. This position was previously found to provide the lowest pressure on the
suction side of the ejector, by changing NXP while maintaining constant generator
conditions and the expansion valve closed. In order to make sure that both SP and NXP
are in the correct position during each run, both variables were set to zero (fully closed)
and then retracted to the desired set points using the stepper motors. Each experiment
started when the temperature of the working fluid at the primary inlet of the ejector
reached steady state, at approximately 83 °C (approx. 15 ºC superheat) and pressure of
about 10 bar. At the outlet, the condenser pressure always started below the critical back
pressure by adjusting the condenser temperature with the water chiller in the beginning
of the test run. During the test, it was allowed to increase constantly beyond pc,cr. This
way, the entire operating curve of the ejector could be determined, identifying double
choking, critical operation and single choking regimes in a single experimental run. The
expansion valve was adjusted to set an evaporator pressure of about 2 bar, resulting in
an evaporator temperature of approximately 9 °C (with 2 °C of superheat).
Each experimental run was stopped when one of the recorded flow rates became smaller
than the minimum range indicated for the flow meters. The data were then analyzed
according to the ejector flow regimes. Mass flow rates and enthalpies were calculated
from the measured data using EES. Below the critical back pressure, both and
remained constant, and so did the performance parameters (COP and λ). Here, the
ejector performance was assessed by simple arithmetic average. Beyond the critical
back pressure, the ejector operates in single choking regime, with decreasing secondary
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mass flow rate, COP and λ. In this regime, a linear regression line for the COP and λ as
a function of back pressure (temperature) was fitted. Then, the critical back pressure
was determined by intersecting this regression line with average values of the double
chocking regime (see Fig. 4).
5. Results and Discussion
Fig. 6 shows the experimental cooling cycle COP as a function of condenser pressure,
for a spindle position of 5.00 mm (Tg=83 °C, Te=9 °C). It can be seen from Fig. 6 that
the critical back pressure was approximately 3.48 bar, corresponding to a condensation
temperature of 24.8ºC. Below pc,cr, COP was approximately constant with a value of
about 0.58. This corresponds to an entrainment ratio of 0.72 and a cooling capacity of
1.52 kW. Beyond the critical back pressure, ejector performance fell quickly, with a
COP as low as 0.2 for a condenser pressure of 3.9 bar (Tc=28.7ºC).
It is important to note that the entrainment ratio follows the same evolution as COP. The
spindle tip position influences the primary nozzle throat area and thus affects . Fig. 7
presents the steady state primary and secondary mass flow rates as function of spindle
position, for constant generator and evaporator pressures/temperatures. Additionally, a
line representing the equivalent fixed geometry ejector primary mass flow rate is
indicated. When the spindle traveled between 3.75 and 5.25 mm, the mass flow rate
variation was almost constant (linear) leading to an overall increase of about 74 %.
When SP was varied between 5.25 to 5.75 mm, the change in was only about 5%.
This latter variation was within the range of the accuracy of the flow meters, and
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therefore it can be concluded that a further increase in SP had almost no influence on
. One may also observe that was kept almost constant around 5.5 g/s, for all
spindle positions.
The determination of the operating curve (e.g. in Fig. 6) was carried out for all the
spindle positions within the considered range. Fig. 8 depicts the results obtained for
spindle positions from 4 to 6 mm, with constant Tg and Te of 83 °C and 9 °C,
respectively. The critical operational line resulting from the experimental data is also
indicated in Fig. 8. Since the generator and evaporator conditions were kept constant
during these experiments, this line can be considered as the optimal operation line as a
function of the condenser pressure. For instance, for a condenser pressure of 3.2 bar
(Tc= 21.9 °C) (vertical solid line), operating the ejector with a SP of 5 mm, resulted in a
COP of 0.58 (horizontal dotted line). Moving SP to 4.5 mm, the ejector operated with a
COP of 0.68 (horizontal solid line), which represents a 16% performance improvement.
By moving SP further to 4 mm, the measured COP was 0.65 (horizontal dashed line),
which was actually 7% smaller than the COP obtained with SP=4.50 mm. Therefore,
this latter value of SP can be considered as optimal, allowing the ejector to operate near
critical condition. With an increase in the condenser pressure, e.g. to 3.4 bar (Tc= 23.9
°C) (vertical dashed line) with 4.50 mm SP, the ejector works under single choking
regime and COP drops to approximately 0.5 (horizontal double dotted line). However, if
the spindle is opened to 5.00 mm, the ejector continues working in double choking
regime and the COP is higher (approximately 0.58).
The evolution of the evaporator cooling capacity, COP and λ with spindle position, for
the cases with double choked flow inside the ejector is presented in Fig. 9. One may
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note that both performance indicators decreased with SP between 3.75 and 5.25 mm.
This tendency can be explained by the fact that during the experiments, secondary flow
rate remained essentially constant, while the primary flow rate increased, leading to a
constant decrease of the performance indicators (see Fig. 7). This phenomenon is also
demonstrated by the approximately constant cooling capacity, as shown in Fig. 9. For
spindle positions between 5.75-6.25 mm, the performance parameters remained almost
constant, since the primary mass flow rate became independent of the SP, as also shown
in Fig. 7. The small variation on the cooling capacity (relative standard deviation bellow
3%) can be associated to the repeatability of the experiment runs, since not all
experimental conditions could be fully controlled (e.g. room temperature).
Based on Fig. 9, one may conclude that for optimal operation, the optimal SP is 3.75
mm. However, closing the spindle, the critical back pressure drops as indicated by Fig.
8. These two characteristics allow the optimization of the VGE control (SP), and the
results are shown in Fig. 10, where COP, pc and SP are depicted in the same diagram.
The dotted curves in the figure indicate a critical operating line for COP and pc,cr. This
analysis allows the identification of optimal SP control, as well as the assessment of the
benefit of applying a VGE instead of a fixed geometry ejector. Fig. 10 shows that pc,cr
varied from 2.8 to 3.8 bar (Tc from 17.4 to 27.8 °C) depending on the spindle position.
The critical back pressure of 3.8 bar (Tc= 27.8 °C) was obtained with an SP of 5.75 mm
or higher. The corresponding cooling cycle COP was approximately 0.48. This value of
COP also corresponds to the performance of a fixed geometry ejector working in double
choking regime, since in this case the spindle has no influence on the primary mass flow
rate and thus on ejector operation. Condenser pressure typically depends on the climatic
conditions and it varies with time. In a situation when the ambient temperature
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decreases with time (end of the day), condenser pressure also reduces. For instance,
considering a condenser pressure of about 3.65 bar (about 26 °C condenser temperature)
(solid arrows), the spindle position that resulted in optimal operation conditions was
5.25 mm, ensuring critical operation with a COP of 0.5. This represents an improvement
of about 6.4%, comparing to the fixed geometry ejector under the same generator and
evaporator conditions. Likewise, if the condenser pressure decreased about 0.65 bar
(condenser pressure of 3 bar, about 20 °C condenser temperature) (dashed arrows), the
optimal operation condition corresponding SP was 4.00 mm, and this corresponds to a
COP improvement of about 70%. The benefits of a VGE with R600a working fluid
have already been demonstrated by Varga et al. (2013a), using CFD simulations.
However, these are the first experimental results that empirically prove the benefit of a
VGE over a fixed geometry ejector.
The choice of the optimal spindle position can be perhaps better understood through the
experimental data presented in Fig. 11. The figure shows COP as a function of the
spindle position, for 3 sets of condenser pressures. For each set, it is possible to select
the optimal SP resulting in the highest COP. For instance, for a condenser pressure of
3.48 bar (Tc= 24.7 °C), if the SP is 5.00 mm, COP is higher than with the ejector
operating with a SP position above or below that value, and the ejector runs under
critical condition. The figure also shows the same tendency for 3.22 bar (Tc= 22.1 °C)
and 3.76 bar (Tc= 27.5 °C) condenser pressures. In general, the higher the condenser
pressure, the more open the spindle should be for optimal operation.
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An attempt was made to compare the present performance results to previously
published data using fixed geometry ejectors and R600a as working fluid. A direct
comparison is however difficult, since most published data with R600a rely on
simulation studies. For example, Pridasawas and Lundqvist (2007) presented simulation
results. In their work, the entrainment ratio, as function of the condenser and generator
(70 to 120°C) temperatures, for an evaporator temperature of 15 °C, was studied. For a
Tg= 80°C, λ was about 0.32. Since the present experimental results were obtained for a
considerably lower evaporator temperature of 9 °C and somewhat higher generator
temperature (83ºC), it could be expected that the entrainment ratio would be lower than
0.32. In the present work λ was found to be always above 0.5 (see Fig. 9), for all studied
spindle positions. As mentioned before, Butrymowicz et al. (2013) carried out
experimental analysis with an R600a ejector which, according to the authors’
knowledge, are the only experimental data with isobutane available in the literature until
this date. The ejector performance indicators were presented for a single generator and
evaporator temperature, 63.5ºC and 7ºC, respectively. Considerably lower values were
reported (λ= 0.19 and COP= 0.15) which suggests that these preliminary results can be
considered rather promising.
6 Conclusions
In the present paper, the experimental performance results with a small (1.6 kW cooling
capacity) variable geometry ejector using isobutene (R600a) were presented. The work
aimed to characterize the performance of the VGE prototype, with active adjustment of
the primary nozzle geometry, for fixed generator and evaporator conditions, while
changing condenser pressure.
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Under constant upstream and downstream conditions, changing the spindle position
(SP) resulted in a variation of about 80% in the primary flow rate, without significantly
affecting the stream pressure. Within the range of condenser pressures considered, COP
values ranged between 0.45 and 0.88, which can be considered as excellent, especially
for the generator and evaporator temperatures applied.
The benefits of using a variable geometry ejector, compared to a fixed geometry design,
were also demonstrated. It was found that the larger the difference between the actual
and design condenser pressure, the larger the benefits of using the VGE. The
improvement in COP was as high as 85%. The existence of an optimal SP, dependent
on operating conditions, was also experimentally demonstrated for the first time.
The present work should be considered as preliminary, since it only covers a single
generator and evaporator temperature condition. In order to develop an adequate
algorithm that allows for the optimal control of the SP and also NXP, the presented
experimental work will be extended to cover a range of operational conditions
(generator, evaporator and condenser temperatures) that can be expected in a solar
driven air-conditioning system. This way, a complete optimal operation map can be
built. Future work will allow the full characterisation of the VGE cooling cycle, with
the objective to develop and implement an adequate control strategy depending on
operating conditions.
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Acknowledgments
The present work was developed within the framework of the “Investigation into an
Improved Ejector for Variable Operating Conditions” research project. The authors
wish to acknowledge the financial support of “Fundação para a Ciência e a Tecnologia
(FCT)”, through contract PTDC/EME-MFE/113007/2009.
The authors wish to acknowledge Armacell for supplying thermal insulations used in
the test rig. The authors would also like to express their gratitude to BaxiRoca for
providing the equipment used to charge the ejector cycle.
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Fig. 1 – Schematic drawing of the experimental test rig.
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Fig. 2 – A photograph of the experimental apparatus (1 - ejector; 2 - generator, 3 - evaporator, 4 - condenser, 5 - circulating pump, 6 - expansion valve, 7 - vapor separator, 8 - SP and NXP stepper motors).
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Fig. 3 – Schematic view of a typical ejector.
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Fig. 4 – Ejector operation regimes for constant Tg and Te.
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Fig. 5 – Detailed dimensions of the ejector internal geometry.
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Fig. 6 – COP as function of condenser pressure for SP=5.00 mm, with Tg= 83 °C and Te= 9 °C. Condenser temperatures range from 17 ºC (2.8 bar) to 30 ºC (4 bar).
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Fig. 7 – Influence of spindle position on the primary and secondary flow rates for Tg= 83 °C.
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Fig. 8 – COP as function of condenser pressure for different spindle positions, with Tg= 83 °C and Te= 9 °C. Condenser temperatures range from 15 ºC (2.6 bar) to 31 ºC (4.2 bar).
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Fig. 9 – Performance parameters and evaporator cooling capacity at different spindle positions, with Tg= 83 °C and Te= 9 °C.
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Fig. 10 – Influence of the spindle position on COP and critical back pressure, with Tg= 83 °C and Te= 9 °C. Condenser temperatures range from 12 ºC (2.4 bar) to 31 ºC (4.2 bar).
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Fig. 11 – COP as a function of the spindle position for three different condenser pressures (temperatures), with Tg= 83 °C and Te= 9 °C.
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• An experimental study with a variable geometry ejector with isobutane is
presented
• The influence of the spindle position on the ejector performance is assessed
• The critical condenser pressures are determined
• The existence of an optimal spindle position was experimentally verified
• The performance improvements compared to a fixed geometry ejector are
demonstrated
