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The Performance of Refrigeration Expansion Devices

Report for:ELECTRICITY ASSOCIATION TECHNOLOGY LTD

Contract Number: U 1420

Completed at: Department of Mechanical Engineering King's College London

Investigators: GA Vinnicombe and GA Ibrahim

Period: January 1991 to August 1991
Summary

Experiments have been conducted to determine the range of pressure differentials i.e., condensing temperatures for a constant evaporating temperatures, over which a refrigerating system can be successfully operated. It has been confirmed that the characteristics of the expansion valve limits this useful range because of instability at part load and high condensing temperature conditions.

The effect of pressure drop in the liquid line leading to the occurrence of flash gas has been investigated both theoretically and experimentally. The theory agrees well with the experimental results. It is shown that flash gas reduces refrigeration capacity when condensing temperature is low and the expansion valve is fully open. In addition it is shown that flash gas can be eliminated by using either a subcooler or HY-SAVE LIQUID LINE BOOSTER PUMP but, when the operating conditions require a fully open expansion valve, the refrigerating capacity can only be restored to the value when no flash gas was present if a LIQUID LINE BOOSTER PUMP is used.

1. Introduction
For a given set of conditions the compressor power, and to a lesser extent the refrigeration capacity, of a refrigeration system is very much effected by the condensing pressure. The higher the condensing pressure then the smaller will be the refrigerating capacity and the larger will be the compressor power. This effect is demonstrated in Fig.1 which shows the calculated refrigerating capacity and coefficient of performance for a typical system producing chilled brine, at -13`C, over a range of condensing temperatures for both evaporative and dry surface air cooled condensers. It can be seen that, for the most efficient operation of a system, it should run at as small a condensing temperature as possible and incidentally, use an evaporative condenser. Consequently, for best performance, full advantage should be taken of the cooling capacity of the ambient conditions. However, because of the requirements of the expansion device, whose duty is to adjust the refrigerant flow into the evaporator to maintain the required evaporator exit conditions, this is often not done. In fact, in winter, ambient cooling is usually restricted in order to maintain a high condensing pressure.

Almost invariably expansion devices in refrigeration systems take the form of some type of automatically adjustable orifice. In the flooded evaporator the adjustment will be by means of a float switch to maintain the refrigerant level and, in the more common thermostatic expansion valve (TEV), the adjustment is made in response to a temperature sensor at the evaporator exit to maintain a constant refrigerant superheat. This report is concerned only with thermostatic expansion valves. In the TEV there will be a maximum size of orifice when the valve is fully open and, if the conditions require to be larger than this to maintain superheat, the valve will be unable to satisfy the demand and the result will be a superheat greater than that set and, because of the reduced area for boiling, a corresponding reduction in refrigerating capacity. The overall state of affairs is illustrated in Fig.2 which shows how the equilibrium refrigerant mass flow rate varies as a function of pressure difference (Pc-Pe)across the system. Pc is the condensing pressure and Po is the evaporating pressure. At low pressure differences i.e., up to a pressure difference of ^P1 , the refrigerant flow is determined by the, fully open, orifice size of the TEV. The mass flow in this region is, to a first approximation, proportional to the root of the pressure difference and, in fact, manufactures published data suggests that this first approximation is the true condition. When the pressure difference is grater than ^P1 then the flow is determined by the capacity of the compressor and the size of the orifice in the TEV will be reduced automatically to maintain the set superheat. From Fig.2 it is seen that the refrigerant flow and therefor the refrigerating capacity is small at low pressure differences i.e., low condensing temperatures and this is the reason put forward why the condensing temperature must be maintained high. However, as a counter argument to this , it can be seen in Fig.2 that if the point ^P1 was set at the lowest attainable pressure difference for the system for the given conditions i.e., by using a relatively large orifice then the problem with the valve in restricting the capacity would be overcome. Unfortunately there is a limit to this because another factor which limits the range of TEV operation is the tendency for them to become unstable when the orifice is required to be set small to satisfy the conditions i.e., at high condensing temperatures and part load conditions when there is a reduced refrigerant flow requirement. This instability (hunting) results in a cyclic overfeeding and underfeeding of the evaporator and the most serious consequence can be refrigerant liquid being admitted to the compressor leading to possible mechanical failure. manufactures recommend that in order to avoid this instability the capacity should not be reduced to less than 30% of the declared capacity. Part load conditions can be obtained by reducing the swept volume rate of the compressor and this is a common operating condition. A typical part load condition is shown in Fig.2 and the region where hunting is likely to be obtain is indicated.It is clear therefore that the use of a large orifice in the TEV could allow the system to run at low condensing temperatures but the operating range would not be extended because the maximum condensing temperature and the amount of part load would then be limited. One of the purposes of this report is to investigate experimentally the arguments addressed above and to test if they are reasonable.

The other area of TEV operation which this report will examine is that of the effect, on performance, of the presence of refrigerant vapour (flash gas)in the liquid at entry to the TEV. Now the data from which system performance is predicted is obtained usually from manufactures catalogue and this data assumes that refrigerant enters the TEV in the liquid state. Invariably the refrigerant will leave the condenser in the liquid state but, if there is significant pressure drop in the liquid line from the condenser to the TEV, it is possible that, under certain circumstances, vapour will form in the refrigerant and so will enter the expansion valve with the liquid. As a result of this the mean specific volume of the refrigerant will be increased and so it is likely that the mass flow rate through the expansion valve and consequently the refrigerating capacity will be less than if there was no vapour present. This report address this area also in order to provide information on the likely effects of flash gas and ways in which it can be avoided in particular by examining the usefulness of subcooling and of a proprietary booster pump which can be inserted into the liquid line to re-establish the pressure difference.

2.Equipment
The equipment comprised a laboratory refrigeration plant which was representative of one to provide chilled water in practical air conditioning application. It had a design refrigerating capacity of 10kW and could reduce the temperature of the chilled water flowing at a rate of 0.4kg/s from 12`C to 6`C when the evaporating temperature was 0`C. The refrigerant was R-22. The evaporator (chiller) was a plate heat exchanger with a basic rating of 1.7kW/K and the cooling load was provided by electrical heaters in the recirculating chilled water line. The heaters could be thermostatically controlled to maintain a constant return temperature to the evaporator at full and part load conditions. The condenser was water cooled with a shell and tube configuration. The compressor was a two cylinder reciprocating open type with a water cooled head. Part load conditions were obtained predominantly by using different pulley combinations but, because the available pulleys did not conform exactly to the required part load conditions, they were supplemented with an evaporator pressure regulator valve. The full load capacity curve for the compressor is shown in Fig.3. It will be noted that the rated evaporating temperature is less than the actually used but this presented no problems when a driving motor of sufficient power was provided. Two sizes of TEV`s and a microprocessor controlled valve - activated by a stepper motor - were to be tested and these were all incorporated into the system for selection by suitable stop valves when required. The manufacture's data on refrigerant flow rate as a function of pressure difference for the two TEV`s is shown in Fig.4. No detailed data on the electronic valve was available but its orifice size indicated that its capacity was equivalent to the smaller TEV. As part of the test programme the possibility of using TEV`s in combination to overcome their individual deficiencies and so operate over a wider pressure range was to be investigated so means were provided for this to be undertaken in the form of a pressure sensor and solenoid valves. The liquid line from the condenser to the expansion valves was made as short as possible to minimise the built in pressure drop and so avoid possible vapour formation. However, in order to investigate the effects of the occurrence of vapour on the performance of the expansion valve, a needle valve was incorporated in the liquid line so that pressure drop of various amounts could be effected. A sight glass was placed immediately in front of the expansion valve to give a visual indication of the presence of vapour. An investigation into the possibility of preventing the formation of vapour in the liquid line by using both a liquid subcooler and a refrigerant booster pump was to be undertaken and so both of these devices were installed. The subcooler was a water cooled plate heat exchanger and the pump was of a self-unloading centrifugal type which is designed specifically to overcome the problem of vapour in the liquid line.

The test programme required the condenser to be operated, sometimes, at low condensing temperatures which were not attainable in the laboratory using the normal service cooling water. Thus in order to obtain a satisfactory cold supply an ice bank was constructed from which sufficiently cold water could be extracted. The compressor for this ice-bank was, in fact, the one used for the test programme and so, before any testing was undertaken at low condensing temperatures unattainable using the normal water supply, the system was first used to charge the ice bank.

Specific details of the equipment is as follows:

Compressor             Bitzer type IV(W)
Evaporator            SWEP type B8/26

Expansion Valves:
Thermostatic -        Danfoss type TEX 2-1.5
                              Danfoss type TEX 2-2.3
Electronic -          Egelhof type RTC1 / MPS20

Booster Pump -      HY-SAVE LPA

3.Test Program

3.1 No Flash Gas
The first set of tests to be undertaken was to monitor system performance when the condensing temperature was progressively reduced and, at the same time, ensuring that no flash gas was present at the entry to the expansion valve.

At the onset of the programme two types of tests were envisaged, both to be at constant evaporating temperature. In the first case the system performance was to be monitored at a fixed refrigerating capacity as the condensing temperature was progressively reduced. In the second case, performance was to be monitored when the return chilled water temperature to the evaporator was maintained constant, probably by reducing the load, as the condenser temperature was reduced. It was soon realised, however, that the results from the first type of test would have very   limited use because they would be specific to the system under test. They would not, therefore, have any general application nor would it be possible to make comparisons with manufacturers data. As a consequence this method of testing was not pursued. Nevertheless a few tests of this type were performed in the early stages and a result has been included for completeness.

The main body of testing in this section (no flash gas) was to be undertaken with the return chilled water temperature to the evaporator maintained constant. The performance of two TEV`s and the electronic valve were to be compared with manufactures predictions at full (10kW), three-quarter, half and quarter capacity for the evaporator nominally at 0`C and the condensing temperature within the range from 35`C down to as small a value as possible. The rated capacity of the smaller TEV (7.5kW) was less than the design capacity of the rig so it would be tested at full, two-thirds and one-third capacity. In addition an opportunity would be taken with this smaller valve to test the manufactures contention that the valve was capable of a larger duty than the declared duty.

3.2 With Flash Gas

The effect of flash gas on performance was to be investigated by repeating some of the tests described above but this time introducing pressure drops of various amounts into the liquid line with the aid of a needle valve. It was expected that the pressure drop would have two separate effects. In the first case, when there was considerable sub cooling after the condenser, no flash would be generated even after the pressure drop and so there would be a reduction in flow rate only as a result of reduced pressure drop across the expansion valve. In the second case, when flash gas was generated, the reduction in flow rate as a result of the drop in pressure would be compounded when this reduced pressure also gave rise to vapour formation.

4. Results and Discussion

4.1 Testing With No Flash Gas
Results typical of those obtained during early testing when the refrigerating capacity was maintained constant is shown in Fig.5 where the superheat is plotted against condensing temperature. In the initial stage, as the condensing temperature was reduced, the system remained stable and maintained the load by operating at a higher chilled water and superheat temperature. However this condition could not be maintained and at lower condensing temperatures the system became, in effect, unstable and the water and superheat temperatures increased continuously to an unacceptable level at which time the system was shut down. Fig.5 is plotted on a base of equivalent condensing temperature because, in these tests the low pressure differentials were obtained by making use of a pressure reducer in the liquid line to the TEV.

The main set of tests were conducted with the return chilled water temperature and the evaporator maintained constant at 12`C and 0`C respectively. The result of the tests at duty greater than design duty for the expansion valve is shown in Fig.6. It can be seen that they confirm the manufactures contention that the valves can operate at duties larger than the design duty. For this particular valve a duty 50% greater than the declared duty is apparent over the whole range of condensing temperatures.

The result of the tests at full and part load duty for each of the valves is shown in Fig.7 to Fig.9 inclusive. With the icebank installed the minimum condensing temperature attainable was approximately 10`C at full capacity down to approximately 5`C at minimum capacity. It can be stated that over all the conditions imposed, apart from at high condensing temperature and low load where valve instability was observed, the system performed most satisfactory and no practical difficulties were experienced. In fact at the low condensing temperatures the compressor ran far more smoothly and quietly. The measured refrigerating capacity together with the manufactures rated capacity is shown in Fig.7 and in Fig.8 for the TEV`s and the measured refrigerating capacity, only, for the electronic valve is in Fig.9. Two regions of the results are important for this investigation. The first one is that at low condensing temperatures where the desired refrigerating capacity cannot be maintained even when the expansion valve is fully open. The second is at part load conditions and high condensing temperatures where expansion valve instability is possible because it needs to be almost closed. With respect to the first region it can be seen that both the TEV`s behaved broadly as expected and their capacities were significantly larger than the declared capacities. It is noticeable that the extra capacity of the valve in excess of the declared capacity is very significant and allows the system to maintain the set load condensing temperatures much lower than that predicted from the declared data. In practice advantage should be taken of this. Unfortunately, exact improvements cannot be included at the design stage because, clearly, the enhancement will be different for different values as a result of manufacturing tolerances. Nevertheless in seeking to operate at lower condensing temperatures some accounts should be taken of this extra capacity as it seems unreasonable not to utilise the valve's actual capacity. The concept of safety factors in this consideration seems to have no relevance. The results for the electronic valve confirmed the view that it's capacity was similar to that of the smaller TEV. The refrigerating capacity is below the desired value at the low condensing temperatures results from the fact that, for the test conditions, the refrigerant flow could not maintain a full wetted area in the evaporator. A consequence of this is an increase in the superheat beyond the set point and this confirmed in Fig.10 where a typical result of the increase of superheat with reduced condensing temperature is displayed. It will be noted that, as expected, the superheat increase is asymptotic towards the chilled water entering temperature. It is evident that this increase is superheat at reduced pressure difference will not always occur. For these tests it appears because an attempt is made to maintain constant chiller conditions i.e., constant evaporator and chilled water return temperature from the load together with a constant water flow rate. Report incomplete.. by request you can obtain a hard copy.. please use the contact us link above