Low charge transport refrigerator (II). Theoretical and experimental investigation/Mazos dozes transportinis saldytuvas (II). teorinis ir eksperimentinis tyrimas.
Vaitkus, L.
1. Introduction
The first part of this work [1] presented a review of refrigerant charge studies and analysed general strategies of charge reduction. Also
from the analysis of available experimental data it was found that for
low charge system it is advantageous to reduce subcooling in condenser the charge should just ensure complete condensation. Such system could
be the system with high pressure liquid receiver, in which liquid
subcooling is not allowed by design.
Further the applicability of general charge reduction strategies
for refrigerating plant with eutectic system is analyzed together with
some aspects specific to low temperature systems. Experimental results
obtained for low charge eutectic systems are also presented.
2. System architecture consideration
General strategies for charge reduction are (i) use of indirect
system, (ii) use of low-pressure receiver in the suction line rather
than the common high pressure receiver in the liquid line and proper
receivers sizing, (iii) capillary tube used as expansion device, the
diameters and lengths of pipes, (iv) use compressor with small internal
volume and small oil charge, (v) use nonmiscible oils, (vi) use of
compact, low-volume heat exchangers, preferably minichannel heat
exchangers, (vii) use evaporator's direct expansion supplying.
The recommendation to use indirect system can be explained by
significant pipe lengths in split systems and high refrigerant amounts
in pipes. Macchi et al. [2] reported that in the studied cases the ratio
between the charge in the liquid pipes and the total charge was 60% for
the direct expansion system and 40% for the flooded system. Small plants
with split system have separate condensation groups and also often
contain a significant part of their charge in the liquid pipes.
Therefore minimization of liquid pipe diameter and length is one of
necessary steps for charge reduction. With indirect systems the distance
between system components may be minimal, which reduces refrigerant
charge. Palm [3] gives another example, when the charge of indirect
supermarket system is expected to be 11% of the charge of conventional
centralized system. Here analyzed transport refrigerators are already
made compact and implementation of indirect system will not offer
further decrease of liquid line length.
Jensen and Skogestad [4] investigated optimal operation of simple
refrigerating systems and compared different designs. In their analysis
the "active" charge is defined as the total mass accumulated
in the process equipment in the cycle (condenser, evaporator, compressor
...) but excluding any adjustable mass in liquid receivers. It is
assumed that refrigerant holdup change by filling or leaking in receiver
with a constant active charge does not affect the operation of the
cycle. Thus receiver makes operation independent of the total charge in
the system. The two main cases where the receiver is placed are (i) on
the high pressure side after the condenser and (ii) on the low pressure
side after the evaporator. The superheat in evaporator is not optimal,
but some subcooling in condenser is optimal contrary to popular believe.
This claim is supported by theoretical and experimental investigation by
Corberan et al. [5] as well as experimental data from Tassou and Grace
[6], Choi and Kim [7, 8], Cho et al. [9] and Primal et al. [10] (see
discussion in [1]).
Jensen and Skogestad [11] also investigated selection of controlled
variables for simple refrigeration cycles. It was found that for ammonia
refrigeration cycle a good policy is to have no subcooling. Further
savings at about 2% are obtained with some optimal sub-cooling (larger
savings are expected for cases with smaller heat exchanger areas). It
was found that the best control strategy is to fix the temperature
approach at the condenser exit (the difference between temperature of
sub-cooled liquid before subcooling control valve and ambient
temperature).
Palm [3] also recommends using a low pressure receiver in the
suction line rather than the common high pressure receiver and capillary
tube as expansion device in order to decrease refrigerant charge. Low
pressure receiver with the correct charge should ensure operation of the
system without superheat at the evaporator outlet and highest
Coefficient of performance (COP). Capillary tube ensures that no charge
is trapped in a liquid line. Such expansion device is cheap and requires
no initial setting once correctly sized. Evans et al. [12] used the
system with low pressure receiver and four capillary tubes as expansion
devices in their multitemperature commercial refrigerator cabinet.
The low pressure receivers are often called accumulators.
Parameters of available low pressure accumulators and their influence on
system performance are investigated by Wang et al. [13, 14]. They
experimentally and theoretically studied low pressure accumulator with
an oil bleeding hole at the bottom of j-tube.
Such system with low pressure receiver may be very promising for
eutectic systems. Still, its successful implementation requires
development of the receiver with proper parameters. We tested such
system with currently commercially available low pressure receivers and
found the results not good enough--the measured system capacity and COP
was lower comparing to results of baseline system with thermostatic expansion valve (TEV). The available low pressure receivers
(accumulators) do not provide the outlet vapour quality close to
saturated outlet, which would be required to ensure high efficiency.
Since the hardware components for implementation of the system with low
pressure receiver are not currently available, the traditional system
with a high pressure receiver and TEV or electronic expansion valve
(EEV) was chosen for low charge eutectic system development.
3. Features of eutectic systems and the baseline system
Most of the works, discussed previously investigate refrigerating
system working at static conditions. As distinct from these systems, the
eutectic systems always operate at nonstatic conditions. In this sense
the eutectic system is similar to household refrigerator, but the
evaporation temperature in case of eutectic system is varying in much
wider range. Another difference in operation mode is that eutectic
systems are built for maximal cooling capacity and sometimes operate
nonstop through the night, while in household refrigerators the duration
of compressors on-off cycle can be used as additional control variable.
The major tests for eutectic systems performance estimation is
temperature pull-down test. During this test the warm refrigerator with
the eutectic plate temperature equal to ambient temperature is turned
on, and runs until the thermostat cut-out temperature is reached. In the
analyzed refrigerators the eutectic mixture with -33[degrees]C
crystallization temperature was used and the cut-out temperature (air)
for summer conditions was -36[degrees]C. The pull-down time and energy
consumption may be measured and performance estimated for different
inside air temperature intervals. Still, when refrigerator is back to
depot and turned ON during everyday operation, the eutectic solution in
the plates is almost completely thawed and temperature of air in
refrigerated space is close to -20[degrees]C. The period when average
air temperature in refrigerated space goes from -20 to -33[degrees]C is
considered as best indicating the everyday performance of the system
Usually it is required that the eutectic system was able to perform
pull-down at 35[degrees]C ambient temperature for 'north' and
at 38[degrees]C for 'south' applications. During the initial
stage of pull-down test the temperature of eutectic plates is equal to
ambient temperature and the evaporation temperature is lower than
eutectic plate temperature by the certain temperature difference. At
higher evaporation temperatures the heat rejection, condensing pressure
and compressors current would increase a few times comparing to nominal
operating conditions. In order to keep these parameters inside of
operation envelope, the suction pressure in such system must be limited.
Most of eutectic systems are equipped with direct expansion (DX)
evaporator with mechanical TXV and high pressure (HP) liquid receiver.
Usually these systems also have suction--liquid heat exchangers (SLHX).
Simplified diagram of such system is given on the Fig. 1.
The eutectic refrigerators chosen for the investigation and
described in this paper are developed and produced as refrigerated
bodies with an integrated refrigerating plant for delivery vans
(permissible total weight 3.500 kg), to be used for the distribution of
refrigerated goods (vegetables, meat, ice cream etc.) to the end
consumers. The baseline systems (A) and (B) are traditional design, use
R507A refrigerant and are equipped with scroll compressors with liquid
injection through the discharge temperature control (DTC) valve. The
mechanical outlet pressure downstream regulator (OPR) is used for
maintaining a predetermined maximum suction pressure. This regulator is
mounted on a suction line just before the compressor. The maximum
operation pressure (MOP) is chosen experimentally during the pull-down
test at maximal rated ambient temperature (35 or 38[degrees]C) and must
keep the condensing pressure and compressors current within allowable
limits. The specifications of the major components of both baseline
systems are given in Table.
[FIGURE 1 OMITTED]
The experimental systems (C) and (D) were designed and realized in
collaboration between Kaunas University of Technology and JSC Carlsen
Baltic. The project was supported by Lithuanian State Studies
Foundation. The objective of the project was to investigate the
possibility of charge reduction for System (A) below 2.5 kg and for
System (B) below 3 kg.
The initial specific charge of system (A) was about 2650 g/kW, and
for system (B) was about 2817 g/kW. The specific charge of baseline
systems is on the same level as specific charge of competing eutectic
systems, but is significantly higher comparing to other refrigerating
systems. Hrnjak and Litch [15] give compilation of specific charges, and
according to their data for commercially available air-cooled ammonia
chillers the specific charge is in a range of 125-160 g/kW. However,
these systems can not be directly compared--the used refrigerants have
different properties, the systems operate at different temperatures and
have different designs of condensers and especially evaporators.
Since the control system was not defined in the specifications of
the project, various control systems were considered. The most
cost-effective solution would be a system with capillary tube as
expansion device. As mentioned previously, prototype of such system was
tested, but did not offer acceptable performance comparing to baseline
system. According to our estimation the main obstacle here was the lack
of suitable low pressure accumulator (with saturated outlet). Also, the
capillary tube capacity ensuring effective operation of eutectic system
at all the range of operation conditions could not be selected. This
means that the additional valve is needed.
If we consider a system with a small liquid line diameter and
additional control valve, we get a system similar to the one proposed by
Barsanti [16]. The patent describes refrigerating system, which is
claimed to offer the refrigerant charge reduction of 80%. The system has
no thermostatic valve and the expansion is performed through the piping.
The diameter of these pipes is much smaller comparing to liquid lines
usually used. In addition to that, evaporator liquid feeding is limited
by pulsing solenoid. This would be similar to pulse-width-operated EEV,
except that the valve has no metering orifice. Actually it is claimed
that the system should work with a plain on-off solenoid valve. Since
the valve has no metering orifice, the system is claimed to be not
sensitive to gaseous phase refrigerant before the valve. The system also
has no receiver on high or low pressure sides.
Let's analyze of the proposed system. The system without
receiver is possible--most of the works discussed in [1] analyze systems
without receiver. The charge minimization is achieved by charging the
refrigerating system the minimal refrigerant quantity necessary to
system operation. Often the refrigerating systems include significantly
refrigerant receivers that permit to store a refrigerant quantity
definitely higher than strictly necessary working quantity. Such system
may face minor leaks without adding refrigerant. However, from the
ecological point of view such practice is unacceptable since it will
increase total leakage significantly. In the system designed for minimal
refrigerant charge the destination of receiver should be only to
compensate fluctuations of active charge, which may occur at different
operating conditions. If system is working at close to constant
operating conditions the receiver is not needed, which is common case
for heat pump systems. The eutectic system analyzed here must comprise
of properly seized receiver since its operating conditions vary in a
wide range.
The clause considering small diameter liquid line is also
clear--the reduction of liquid pipe diameter is one of the success keys
in a charge minimization strategy. The diameter must be optimized taking
into account the associated pressure losses--the decreased diameter
increases a risk of partial liquid vaporization prior to expansion
device and a bad evaporator's supply. Possible strategies to design
a refrigerating plant with reduced diameter of liquid pipes depend on
system architecture. For refrigerating systems without liquid
recirculation the smaller liquid line diameter may be compensated by (i)
better liquid subcooling in suction--SLHX or economizer or by (ii) an
additional pump at the condenser outlet.
We performed simulation of refrigerating system with SLHX using the
mathematical model presented in [16, 17] with refrigerant properties
from [18] and compressors parameters from [19]. According to the
simulation, smaller liquid line diameter also has additional advantage
in the systems comprising of counter-flow welded tube SLHX. While the
diameter of liquid line decreases, the increase of heat transfer
coefficient from the liquid side is faster than the decrease of heat
transfer area, and kA value of SLHX increases. The pressure losses in
the liquid line will also increase rapidly, but that is not a problem as
far as required refrigerant capacity is ensured. Therefore the
combination of smaller diameter liquid line and EEV without metering
orifice makes sense for charge reduction as well as higher thermal
efficiency of SLHX.
One may doubt the suggestion in [20] to use refrigerating system
without SLHX. This could be considered when the system with absolutely
minimal charge is on target, since generally without the SLHX the vapour
quality at the inlet of evaporator will be higher, bringing in some
charge decrease. However, this is the case when measure reducing
refrigerant emissions also decreases system energy efficiency. In-deep
investigation of SLHX influence on system performance is given for
example in Klein et al. [21]. We also performed experiments in order to
estimate the influence of SLHX to low temperature refrigerating system.
Two refrigerating systems were tested in parallel--the reference system
(A) and otherwise identical system with SLHX from the system (B). The
system with higher efficiency SLHX at 'OFF' instance reached
both lower average eutectic plate temperature and lower average air
temperature. From ON to OFF the pull-down energy consumption decreased
by 1.74 kWh (15% decrease). The measured performance improvement is
higher than obtained through simulation of SLHX. In order to explain the
observed improvement, the deeper analysis including influence of SLHX on
heat transfer and hydraulic losses in evaporator is required. Still, the
experiment demonstrates the importance of the SLHX for efficiency of
eutectic system.
When analyzing the baseline eutectic system in its nominal
operation mode we noticed that the system may operate with significantly
lower refrigerant charge without negative impact on performance.
However, such system with 'filled on demand' charge
demonstrated inadequate performance during pull-down test at higher
ambient temperatures. We come to conclusion that this performance
degradation is related to one more design feature, which as far as we
know is entirely overlooked in literature, but still strongly influences
the total refrigerant charge in eutectic systems. This feature is a mean
to ensure the MOP.
4. MOP and cycle parameters
The eutectic system is low evaporation temperature system with the
maximum operation pressure limitations imposed by compressors maximal
electrical current as well as maximal pressure in condenser. The
eutectic systems mostly use the mechanical outlet OPR for maintaining a
predetermined maximum suction pressure, which is chosen experimentally
during the pull-down test at maximal rated ambient temperature.
Let's analyse performance of such system during the initial
stage of pull-down at high ambient temperatures. The ambient temperature
[t.sub.amb] = 38[degrees]C and the same is initial eutectic plate
temperature. The TXV is limiting the refrigerant supply to evaporator
according to superheat at outlet, i.e. expansion device capacity (EDC)
or MOP are not a limiting factors. For the analyzed cycle conditions the
difference between condensing and evaporation pressures is about 11
bars, which is enough for operation of TXV. In such case the heat
transfer surface dedicated to evaporation is almost constant and the
difference between the plate and evaporation temperatures correlates
well with cold capacity of the compressor--as cold capacity decreases,
the temperature difference also decreases. For the baseline system (A)
the OPR pressure is set to 1.7 bar and the condensing temperature should
not exceed 55[degrees]C. The effectiveness of SLHX in system (A) is
approximately equal to [[eta].sub.t] = 0.5. The compressor's cold
capacity according to manufacturer's data is approximately 1.5 kW
(a bit lower than nominal) and the temperature difference in evaporator
can also be assumed a bit lower than in nominal mode of operation (7 K);
the evaporation temperature then is equal to 31[degrees]C. Let's
assume the vapour superheat at evaporator outlet equal to 6 K and the
temperature before OPR will be about 46[degrees]C. With such assumptions
the simplified diagram of the cycle is given on the Fig. 2.
[FIGURE 2 OMITTED]
If the refrigerating system is controlled by EEV, the OPR valve is
not used, since the controller itself may limit the refrigerant supply
to evaporator according to suction pressure--MOP function is overriding
the superheat control. Simplified diagram for such system is given on
Fig. 3. Let's analyse such system at the same conditions as in
previous example. Since OPR valve is not present in the system, the EEV
is limiting refrigerant supply to the evaporator so that the pressure in
the evaporator does not exceed the MOP settings, and evaporation
temperature is approximately equal to -35[degrees]C. Difference between
the plate and evaporation temperatures is then greater than 70 K--about
10 times bigger, than assumed in the previously analyzed case. Since the
compressors cold capacity in both cases is approximately the same, the
heat transfer area dedicated for evaporation proportionally decreases,
increasing the heat transfer area dedicated for refrigerant
superheating. At the initial stage the temperature of superheated vapour
at evaporator outlet is close to plate temperature. During further
operation the plate temperature decreases, the evaporator is gradually
filled with liquid and the superheat in evaporator decreases to
~6[degrees]C. The corresponding p-h diagram is the Fig. 4 (cycle a).
During the baseline system operation its p-h diagram also undergoes
transformations--the evaporation temperature decreases, vapour superheat
in SLHX increases and pressure drop in OPR decreases. Eventually the
initial cycle from Fig. 2 also transforms into the cycle a from Fig. 4.
From this moment neither OPR nor MOP function of electronic controller
does influence the system performance.
[FIGURE 3 OMITTED]
[FIGURE 4 OMITTED]
In real refrigerating system the discharge temperature is
controlled by DTC valve, but on the given diagrams (Figs. 2 and 4) the
compression is assumed isentropic. This assumption makes no difference
for analysis of the evaporator. However, for analysis of the condenser
the vapour temperature after compression directly influences the
refrigerants mass flow in condenser (when liquid injection is used, the
refrigerant mass flow in condenser is bigger than refrigerant mass flow
in evaporator).
Another simplification is constant condensing temperature. For the
conditions on the Figs. 3 and 5 (cycle a) the condenser's heat
rejection is not strictly the same, and the condensing temperatures
should be different. However, the heat rejection differs by less than
6%, and the average temperature difference would change about
0.5[degrees]C. For the simplicity reason this change was neglected.
5. Simulation results
In order to estimate the refrigerant charge distribution in two
phase regions, the Premoli et al. [22] correlation was chosen. According
to Rice [23], Premoli correlation is one of the most effective for
refrigerating systems. According to Kuijpers et al. [24], only the
Premoli correlation gives satisfactory results for the charge
calculation in a condenser. According to Farzad and O'Neal [25],
models depending on the mass flux, such as Premoli correlation, predict
a refrigerant charge that is more significant and is more effective.
The refrigerant properties required for the simulation were taken
from REFPROP database [18]. The influence of oil in the condenser and
evaporator for fluid properties was not taken into account. The
refrigerant mass flow rate in the evaporator was estimated according to
the data, provided by compressors manufacturer [19].
The evaporators of both baseline systems comprised of two parallel
branches, made of [empty set]15 x 1 tube. Each branch was controlled by
separate TXV according to superheat in the branch. However, such
evaporator setup is not suitable for system with EEV. The controllers
currently available on the market can only control one branch per
controller. In the best case the setup for two branches system can
comprise of cold room controller + additional slave controller, both
sharing the signal from the single pressure probe. While such setup can
be used, the cost increase is too high for competitive transport
refrigerator.
For the systems with EEV therefore are two options--to use parallel
design with distributor or serial design. The parallel design makes it
possible to decrease the evaporator's tubes diameter which would
result in lower evaporator's volume, lower cost, weight,
refrigerant charge and even potentially better efficiency. We made the
preliminary tests of such system, and obtained controversial results.
The tests confirmed the higher efficiency of the parallel system, but
also revealed some problems caused by uneven refrigerant distribution
(mainly caused by technological restrictions). Therefore the serial
evaporator design was chosen for eutectic system with EEV. In order to
keep hydraulic losses within acceptable limits, the diameter of
evaporator's tube was increased to [empty set]18 x 1.
Calculating the refrigerant charge we considered the [empty set]15
x 1 two branch evaporator design for baseline system and [empty set]18 x
1 serial design for system with EEV.
For the baseline system (A) at the conditions corresponding to Fig.
2 the calculated refrigerant charge in evaporator is approximately 1.4
kg, while for conditions corresponding to Fig 4 (cycle a) the charge
decreases to only 0.44 kg. During this cycle transformation at initial
stage of pull-down the refrigerant charge in the evaporator decreases by
almost 1 kg, which makes approximately 25% of total charge in baseline
system (A). For the system (C) with EEV at the conditions corresponding
to Fig 4 (cycle a) the refrigerant charge in the evaporator is equal to
0.54 kg, which is more than for baseline system (A); the charge
increases due to bigger volume of the evaporator. However, at the
initial stage of pull-down for the system with EEV the charge in the
evaporator is smaller due to higher vapour quality and smaller tube
length dedicated to evaporation. During further stage of
pull-down--crystallization of eutectic plates (Fig. 4, cycle b)--the
heat rejection decreases, condensing temperature slightly decreases,
evaporation temperature decreases to -43 [degrees]C. As a result the
vapour quality after expansion changes insignificantly and the charge in
evaporator remains almost the same.
The required refrigerant charge at high pressure side remains
almost constant for all analyzed stages of the baseline system (A)
operation at 38[degrees]C ambient temperature. According to the
simulation, about 0.84 kg of refrigerant can be found in condenser, 1.2
kg in receiver and about 0.27 kg in liquid line (tube [empty set]10 x 1,
L = 5.5 m). Calculating the charge in the receiver we made the same
assumption as Rajapaksha and Suen [26] that 1/6 of the receiver is
filled with liquid refrigerant. Then in the system (A) at initial stage
of pull-down the charge is about 3.71 kg, at -35.5[degrees]C evaporation
temperature it decreases to 2.75 kg and during crystallization remains
about 2.73 kg. Some additional charge remains dissolved in the
compressor oil and in filter--dryer. The calculation of charge dissolved
in oil is quite complicated, since oil is circulating in the system and
may be at different temperatures and pressures. Thorough evaluation of
all the piping was also not performed. Still, the maximal calculated
value of refrigerant charge will be close to the charge, which is
actually filled into the baseline system (A).
In Corberan et al. [4], Primal et al. [27] the charge distribution
between the condenser and evaporator is different, but these sources
analyze the systems with liquid sub-cooling in the condenser. As noted
Hrnjak and Litch [15], the liquid sub-cooling is a large contributor to
total charge since the sub-cooling region holds most of the total charge
inventory. In systems with the high pressure receivers the sub-cooling
in the condenser is not possible, and refrigerant charge in the
evaporator is relatively bigger. For the eutectic system the charge in
the evaporator is also increased due to relatively big volume of the
evaporator.
Neglecting the charge increase in the evaporator during the initial
stage of pull-down, the active charge could be considered almost
constant. Even for different ambient temperatures the variation of
active charge is very small. For example during the crystallization at
nominal conditions (Fig. 5, cycle c) the active charge in baseline
system is 2.53 kg. Some charge increase in the evaporator is compensated
by the charge decrease on the high pressure side. The charge increase
during the initial stage of pull-down is caused by interaction of TXV
and OPR, while in the system with EEV during this stage of operation the
charge in the evaporator even decreases.
Unfortunately, the system with EEV may not always be competitive
due to higher cost. For the low-cost low-charge systems the alternative
control system was proposed. Such system could comprise of the
mechanical TXV with MOP function. The lowest MOP pressure for the
available expansion valves corresponds to -25[degrees]C evaporation (2.6
bar for R507A), which could be considered acceptable. The experimental
system with such TXV was built out of baseline system (B)--for this
system the OPR settings correspond to -28[degrees]C, which is very close
to nominal MOP of TXV. Unfortunately, during the tests the MOP function
in mechanical valves was not accurate enough, and suction pressure
regulator OPR was still required. When OPR settings were set to nominal,
the evaporation pressure increased, and resulted evaporation pressure at
initial stage of pull-down was about 6 bar. The systems (B) and (D) are
also equipped with higher efficiency SLHX with [[eta].sub.t] = 0.5.
Operation of such system has similarities with both previously
analyzed systems. Initially the refrigerant supply to the evaporator is
limited by MOP function of TXV and only fraction of evaporator surface
is used for heat transfer. While the plate temperature decreases, the
evaporator is filled with liquid until all the evaporator is working and
TXV is controlled according to superheat in evaporator. This point
corresponds to the biggest refrigerant charge in the evaporator due to
low vapour quality and high pressure. Note that evaporation pressure at
this point was found to be much higher, than rated MOP pressure of TXV.
Resulting p-h diagram for such system at previously described conditions
is given on Fig. 5 (cycle a). Estimated refrigerant charge is equal to
1.1 kg, which is significantly less than maximal charge estimated for
baseline system (A). Comparing the charges one should also have in mind
that evaporator's volume in baseline system (B) is ~33% bigger.
From this point the evaporation pressure decreases until it equalizes
with OPR pressure (Fig. 5, cycle b). The charge in evaporator at this
point decreases to 0.68 kg.
[FIGURE 5 OMITTED]
The successful charge minimization strategy also include careful
selection of the receiver. Previously, when the charge minimization was
not considered a priority, the fundamental principle in selecting the
size of liquid vessels was to choose them large enough that during
operation they never become completely full of liquid or completely
empty. However, the complete filling of the receiver with liquid should
not cause a problem if no valve is present between the receiver and the
condenser. During the normal operation the receiver should only
compensate the variation of active charge in evaporators. The service
technicians commonly wish having a possibility to move all the charge
from the low pressure side to the high pressure side (e.g. for replacing
compressor). In such case the liquid refrigerant can be stored not only
in receiver, but also in the condenser, i.e. the common volume of the
receiver and condenser should be able to hold the entire refrigerant
from the system (with some reserve volume).
Also, it is considered that liquid should not be permitted to
completely drain from the HP receiver, because this would adversely
affect the performance of control valves. If the SLHX is implemented,
significant amount of vapour may be condensed in its liquid line without
affecting the control valves. From our experience, the refrigerating
system performance is not affected significantly as far as clear liquid
with some bubbles of the vapour can be seen in a sight glass (after
receiver). Still we support the recommendation to have at least 1/6 of
the receiver volume filled with liquid in order to avoid incomplete
condensation. The baseline system (A) comprised of 4.3 litres receiver.
Following the discussed principles the volume of the receiver for
corresponding low charge system (C) was decreased to 1 litre. Then at
55[degrees]C condensing temperature the refrigerant charge in the
receiver (with 1/6 volume full of liquid) decreased from 1.23 kg to 0.29
kg.
Another parameter--the diameter of liquid line was already
discussed previously. The tube for liquid line was changed from [empty
set]10 x 1 to [empty set]6 x 1, which decreased refrigerant charge in
liquid line from 0.265 kg to 0.065 kg. The liquid line diameter may be
decreased even further (system with [empty set]5 x 1 liquid line was
successfully tested), but further charge reduction is insignificant.
For the low charge system (D) with mechanical control the 1.5 litre
receiver was chosen. Also the condenser from the baseline system (B) is
not suitable for low charge system--at high ambient temperatures such
condenser alone holds almost 1.5 kg of refrigerant, which makes the
system with 2.9 kg charge beyond the reach. This condenser was replaced
with special low-volume (2.25 instead of 3.98 liters) condenser with the
same capacity. According to the simulation, such system should be able
to work with 2.9 kg of refrigerant. The decreased tube diameter causes
some increase of pressure drop, but estimated effect on condensing
temperature is insignificant.
6. Test results of low charge systems
The low charge experimental systems (C) (with 2.4 kg charge) and
(D) (with 2.9 kg charge) were developed and built using previously
discussed principles and recommendations. Their test results were
compared to baseline systems (A) and (B). The tests were successful--the
low-charge systems demonstrated similar performance as baseline systems
with regular charge in a whole range of ambient temperature. No
evidences of insufficient charge were observed.
Also it was found that 2.4 kg system with EEV (system (C)) has some
reserves for charge decreasing--no performance degradation was observed
when the charge was further decreased by 400 g. The thorough experiments
were not done since the objectives of the project were already reached.
The system with mechanical TXV and OPR (system D) does not have reserves
for further charge decrease with current components.
Another question is considering significance of MOP control for
eutectic systems operating at high ambient temperatures. One could
think, that previously discussed 'initial period' lasts for a
few minutes, and therefore could be neglected when selecting the charge.
The Fig. 6 gives the temperatures, measured on the outside surface of
the tubes, entering the eutectic plates. As can be seen, it takes more
than 5 hours from the beginning of the test to the moment, at which all
the evaporator is working and the MOP function is no longer interfering
with control. During this period the system filled 'on demand'
may operate with incomplete condensation, causing significant
degradation of performance.
On the Fig. 6 also two MOP corrections can be seen, when the system
was turned off by the high pressure protection, after which the MOP
settings were reduced.
All the previously discussed systems were equipped with copper
tube/aluminium fin coil condensers. Parameters of (B) condenser would be
typical to the condensers, used in refrigerating industry, while the (A)
condenser is special low-volume design. The further charge decrease is
possible with the implementation of new low-volume condenser. The
aluminium microchannel condenser, similar to the one described in [15],
was also tested and demonstrated superior performance. The inner volume
of the condenser was only 1.345 litres, which is only 60% of current
low-volume condenser, used in systems (A), (C) and (D). Comparing to
condenser of system (B), the volume of new condenser is only 34%.
Implementation of such condenser alone allows decreasing the refrigerant
charge by 0.3 kg for systems (A), (C) and (D), while for system (B) the
charge decrease is 1 kg. For the system based on system (B) and equipped
with lower diameter liquid line and such condenser we may decrease the
volume of the high pressure receiver from 6.9 to 4.6 litres, while the
refrigerant charge for such system decreases from 6.0 to 4.5 kg even
without implementing the new control system. Such the system was also
tested and no performance degradation was observed.
[FIGURE 6 OMITTED]
When microchannel condenser was implemented together with new
control system, even better results were achieved. The new system was
based on the baseline system (A), but equipped with the new condenser,
higher capacity SLHX from system (D) (liquid line [empty set]6 x 1 mm),
2.3 litre high pressure receiver and new mechanical control system (TXV
with MOP + OPR). The system was tested thoroughly and proved to be
perfectly functional with 2.5 kg refrigerant charge. The efficiency of
the system even increased due to higher capacity of SLHX--the measured
COP was higher by ~9% when compared to baseline system (A). Such system
is also very cost-efficient, since the refrigerant charge and energy
consumption is decreased without increasing system cost.
7. Conclusions
The objective of this project was to develop two low-charge
eutectic refrigerating systems. During the research it was found, that
big variation of active refrigerant charge in baseline systems is
related to significant increase of refrigerant charge in evaporator at
initial stages of pull-down. This charge increase depends on the
controls used in the system in order to maintain maximum operation
pressure. The biggest charge increase is for traditional combination of
TXV and OPR--in this case the charge increase was ~1 kg, i.e. 25% of
total charge in baseline system (A). For low-charge systems it was
proposed to use EEV or combination of TXV with MOP function and OPR. As
far as we know the influence of MOP control to refrigerant charge was
not previously discussed in literature.
Next, it is important to reduce the internal volume of condenser
and evaporator. With implementation of michrochanel condenser the charge
in the condenser may be decreased by 40-70%, comparing to traditional
copper tube/aluminium fin coil condensers. This makes 7 to 17% of the
total refrigerant charge in the baseline systems. With the parallel
plate evaporator setup, the further charge reduction is possible.
The minimization of liquid line diameter may also offer some charge
reduction without any negative effect on performance. The potential of
this measure depends on the length of the liquid line. In analysed case
the charge decrease was ~5% of total charge in baseline system.
The last important measure for charge reduction is to decrease the
volume of receiver. If low-charge system is on target, the function of
receiver should be limited to compensation of active charge variation.
In the analysed case the refrigerant charge in the receiver was
decreased by ~1 kg--from 1.23 kg in system (A) to 0.29 kg in system (C).
This charge decrease makes ~25% of total charge in baseline system (A).
http://dx.doi.org/ 10.5755/j01.mech.18.1.1290
Acknowledgements
Lithuanian State Studies Foundation and JSC Carlsen Baltic provided
support for this work.
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L. Vaitkus
Kaunas University of Technology, K. Donelaicio str., Kaunas, 44239,
Lithuania, E-mail: liutauras.vaitkus@ktu.lt
Received February 02, 2011
Accepted January 25, 2012
Table
Specifications of the main components of systems
Components Specifications System
(A)
Refrigerant Charge, kg. 4.1
Condenser Inner volume, (l) 2.25
Inner area, ([m.sup.2]) 1.19
Outer area, ([m.sup.2]) 26.4
Evaporator Tube, (mm) [empty set]15 x 1
Total length, (m) 45
Inner volume, (l) 5.97
SLHX Liquid line, (mm) [empty set]10 x 1
Length, mm 5.5
Receiver Inner volume (l) 4.6
Components System
(B) (C) (D)
Refrigerant 6.0 2.4 2.9
Condenser 3.98 2.25 2.25
1.58 1.19 1.19
25.3 26.4 26.4
Evaporator [empty set]15 x 1 [empty set]18 x 1 [empty set]15 x 1
60 45 60
7.96 9.05 7.96
SLHX [empty set]10 x 1 [empty set]6 x 1 [empty set]6 x 1
7.5 5.5 7.5
Receiver 6.9 1.0 1.5