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The novel use of phase change materials in refrigeration plant. Part 3: PCM for control and energy savings
Applied Thermal Engineering 27 (2007) 2911–2918 www.elsevier.com/locate/apthermeng
The novel use of phase change materials in refrigeration plant. Part 3: PCM for control and energy savings Fuqiao Wang a
a,*
, Graeme Maidment a, John Missenden a, Robert Tozer
b
London South Bank University, Faculty of Engineering, Science and the Built Environment, 103 Borough Road, SE1 0AA, London, UK b Waterman Gore-Mechanical and Electrical Consulting Engineers, Versailles Court, 3 Paris Garden, SE1 8ND, London, UK Received 23 August 2004; accepted 17 June 2005 Available online 10 May 2007
Abstract Phase change materials (PCMs) have been recognized as energy storage tanks since the 1980s. The ÔtankÕ has been introduced into the refrigeration system to enable its capacitance to take account of fluctuations in the daily cooling load. However, this part of the paper will present a novel control purpose of using PCM in refrigeration systems. The novel application of PCMs in refrigeration systems at different positions in the refrigeration cycle circuit with a shell and tube structure has been investigated extensively by the novel mathematical model presented in part 2 of the paper. The results show that for energy savings, PCMs at different positions give coefficient of performances (COPs) up to 8% through lowering the sub-cooling. PCMs also improve the system COPs up to 4% and 7% for the thermostatic expansion valve (TEV) and orifice systems, respectively, by minimizing the superheat. Further benefits such as system stabilization were also observed in this investigation. Ó 2005 Elsevier Ltd. All rights reserved. Keywords: Phase change materials; Refrigeration; Air conditioning; Control; Thermal energy storage
1. Introduction 1.1. The different purpose of using PCM from traditional usage The traditional use of PCM in refrigeration system is of a cold storage tank. During Ôoff-peakÕ periods, the refrigeration system produces cooling for the PCM storage system. The stored cooling energy will be released during peak periods to assist the refrigeration system to meet the peak load. PCM storage system used as a storage tank achieves smaller size chiller, lower running cost and higher efficiency. A recent development in the application of PCMs in refrigeration cycles is shown in Fig. 1 [1]. This improves *
Corresponding author. Tel.: +44 20 78157634; fax: +44 20 78157699. E-mail address: [email protected] (F. Wang). 1359-4311/$ - see front matter Ó 2005 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2005.06.010
cycle control stability and also results in large energy savings. Compared with a conventional system higher COP can be achieved by increasing sub-cooling by PCMB, minimizing superheat by PCMC and reducing condenser pressure by PCMA. These novel methods of using PCMs for controlling purpose can be introduced together in traditional vapour compressor refrigeration systems. 1.2. The benefits to the system The primary benefit of increased sub-cooling is higher COP. For example, a 15 K increase in sub-cooling levels has been reported to lead to approximately a 10% improvement in peak plant efficiency [4]. A refrigeration system incorporating a phase change material in contact with the liquid lines (as shown in Fig. 1) enables the refrigeration system to benefit from lower night time ambient conditions. At night lower temperatures
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F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918
Nomenclature COP Pc Pe PCM PCMA
coefficient of performance pressure in the condenser (Pa) pressure in the evaporator (Pa) phase change material PCM heat exchanger located between the condenser and the compressor
PCMB PCM heat exchanger between the condenser and the TEV PCMC PCM heat exchanger between the evaporator and the compressor t temperature (°C) TEV thermal expansion valve
Fig. 1. PCM heat exchanger used individually at different positions in the refrigeration cycle.
generate lower condenser pressures (for the case of PCMA) as well as increased liquid sub-cooling (for the case of PCMB). The use of a PCM in the liquid line then provides an artificial cooling load on the condenser during daytime, which will avoid the necessity for inefficient condenser control, as well as allowing the night time cold energy to be stored for release during daytime operation. At the exit from the evaporator the refrigerant vapour must be dry to avoid droplet damage in the compressor and most systems are designed to produce superheated refrigerant to ensure this. The degree of superheat is usually controlled by a thermostatic expansion valve (TEV), which is a proportional controller responding to the control signal, the degree of superheated vapour at the evaporator outlet. In practice an undesirable oscillating behaviour of the TEV control known as ÔhuntingÕ occurs when a low superheat is applied. In extreme cases, the TEV operates as an on/off controller. The use of higher superheat can reduce the unstable operating range. However, this reduces the efficiency of the evaporator. A superheat setting of 5 K is often cited as an efficient compromise, but this fluctuates with load, and many refrigeration systems require higher superheat settings for stable operation. As a consequence, up to 20% of the heat exchanger surface area is required to transfer heat to the dry vapour thus using it ineffectively [3]. The higher the superheat is, the lower the efficiency of the system is. This is in part due to a lower evaporation temperature which increases the compression pressure ratio and work of compression, however it also causes the compression process to begin
further in the superheat zone and this causes the isentropic enthalpy change across the compressor to be greater. The use of a suction line PCM heat exchanger overcomes this by minimizing the degree of superheat in the suction line. The purpose of the suction line PCM exchanger is to fix the superheat in the suction line by providing thermal capacitance and enabling the evaporator to operate approaching wet, which will further improve system efficiency. In order to investigate quantitatively the benefits achieved by the novel use of PCM in the system, a validated mathematical model was used. The details of the mathematical model can be found in part 2 of the paper. Here the results of these novel ideas which include the use of PCM heat exchangers in different position as illustrated in Fig. 1 are introduced and discussed. The mathematical model was based on the following specific plant. 1.3. System to be simulated The test plant is based on an IMI Marstair impact cooling split system model C60/E. The indoor unit comprises an evaporator (IMI impact CU6E) and a fan. The outdoor unit (IMI CUE60) is comprised of a condenser, a compressor and a propeller fan. The details of the plant geometric parameters can be referred from part 1 of the paper. This paper will exam the benefits of the novel uses of PCM heat exchangers as shown in Fig. 1 individually by using the model that is developed in part 2 of this paper, and also discuss some parameters which affect the per-
F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918 3.2
COP
3
2. PCMB increasing the sub-cooling
2.9 2.8
During the simulation, PCM heat exchanger B was assumed to be charged with E21 solution that is made by EPS Ltd Company. This is Grauber salt eutectic solutions complete with nucleating agent, gelling agent and stabilizers. However, the bulk of the solution is Grauber salt. Its physical properties are listed in Table 1. The heat exchanger effective length is 4.32 m and its diameter is 110 mm. In the simulation, only the melt temperature was considered to change from 16 °C to 28 °C, the other parameters were fixed. Mathematical investigation shows that different phase change temperatures have different influences on the system. The results are shown in Fig. 2 which assumes the air inlet temperature to the condenser was fixed at 30 °C, and the air inlet temperature to the evaporator was fixed at 24 °C. The system was charged of
Table 1 PCM materials physical properties Fusion heat (kJ/kg)
Transition temperature (°C)
Thermal conductivity (W/m K)
Density (kg/m3)
150 140
21 4.2/3.5
0.43 0.44
1480 1469
35
tpcmb=16 tpcmb=19 tpcmb=22 tpcmb=25 tpcmb=28
30 25 20 15 10
2.7
2.1. Different phase transition temperatures with different sub-cooling
5 0
2.6 0
1000
2000 3000 Time (sec)
4000
0
0.04
0.03
0.02
0.01
16_mid_discharge 28_mid_discharge 16_mid_charge 28_mid_charge
16_end_discharge 28_end_discharge 16_end_charge 28_end_charge
0 0
500
1000
1500
2000 3000 Time (sec)
4000
3200 g of Freon22(R22) in the model, and the geometric sizes of the heat exchangers including the PCMB heat exchanger and TEV system were fixed. Fig. 2 shows the COP result from this investigation and identifies how the system COP varies with the PCMB temperature and time. It can be clearly seen that a higher COP can be achieved by lowering the phase change temperature. This is because greater sub-cooling is achieved with lower phase change temperature as also shown in Fig. 2. However, the phase change temperature will also affect the ability to recharge the PCM. Fig. 3 compares two PCMB types with different phase transition temperatures, namely 16 °C and 28 °C. Their phase interfaces moving against time are shown for the discharge (melt) process. The lower temperature PCMB can be seen to discharge faster than the higher temperature version. Fig. 3 also compares charging (solidification) of the same two PCMB systems in an environment of 12 °C, it can be seen that the lower temperature PCMB recharges more slowly than the higher PCMB option. It can be seen that the lower temperature PCMB takes approximately 4 times as long to recharge
0.05
16_entrance_discharge 28_entrance_discharge 16_entrance_charge 28_entrance_charge
1000
Fig. 2. System COP and the sub-cooling various with PCMB temperature and time.
0.06
Phase-interface position diameter (m)
PCMA&B PCMC
tpcmb=28 tpcmb=25 tpcmb=22 tpcmb=19 tpcmb=16
3.1
Sub-cooling (˚C)
formance of these three systems. The numerical investigation initially starts with the PCM heat exchanger at position B, then at C, and finally at A as shown in Fig. 1.
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2000
2500
3000
3500
4000
Time (sec)
Fig. 3. PCMB phase interface moving against time during charging/discharging period.
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3.1. The investigation on PCMC system—TEV system
than to discharge. The recharge usually takes place at night when low ambient temperatures occur. The ability of the PCM to recharge will depend upon the night time temperature, recharge duration and phase change temperature. Lower phase change temperature may improve COP but may limit recharge and therefore careful selection is important for all applications.
3.1.1. The system COP Two systems were simulated in this part: the PCMC system as shown in Fig. 1 and the basic system without PCMC heat exchanger. The plant geometric size is the same as discussed in Section 1.3. For this analysis, a 1630 g R22 initial charge and 5 K superheat was used for the PCMC system, and a 1530 g R22 charge and 10 K superheat was used for the basic system. Compare to basic system, more refrigerant is charged and less superheat is assumed in PCMC system, because when PCMC is integrated into the system results an extra volume charged with gas, this will affect the results if the initial charge does not change, this has been explained in this paper, Part 1. Less superheat is used; it is the main benefit when PCMC is introduced into the system. The air inlet temperatures used for both the two condensers and the evaporators are assumed to be consistent. The air inlet temperature of the condenser was fixed at 30 °C, however, the input air temperature to the evaporator was an average 26 °C. To simulate the effect of load variation the inlet air temperature to the evaporator was allowed to fluctuate as a sine wave with amplitude of 4°. Fig. 4 shows that the PCMC system produced an average COP which is about 4% higher than that of basic system. The superheat in PCMC system is about 5 K lower than that of the basic system. Lower superheat allows higher COP, because it allows the evaporator to operate with a boiling heat transfer coefficient for more of its surface. This improves heat transfer and allows higher evaporator pressure and this reduces the work input requirement. Lower superheat also reduces the work of compression since the constant entropy lines are less acute closer to the saturation line.
2.2. Environment temperature influence on system performance For vapour compression refrigeration system, the higher the environmental temperature is, the lower the system COP produced. However, for the PCM system, the situation is slightly different. When a higher environmental temperature is applied to the PCM system, the higher condenser temperature will tend to increase the temperature difference between the refrigerant and the PCM phase change temperature, allowing greater energy to be transferred to the PCM heat exchanger. The sub-cooling effect of the PCMB will therefore be greater at higher ambient temperatures, which will attenuate the negative effect due to higher condenser temperature. Table 2 shows the simulation results. It can be clearly seen that the COP of the PCM system decreases by 14.56%, i.e. from 2.49 to 2.13, when the environment temperature increases from 16 °C to 28 °C. However, the COP of the basic system (without PCM) decreases by 19.96% from a value of 2.28 to 1.82. This shows that the higher the environment temperature, the greater are the savings that the PCM system brings.
3. PCMC minimizing the superheat Basically there are two expansion devices-thermostatic expansion valve and orifice in refrigeration system; therefore the two systems were simulated in this section. The geometry of the heat exchangers at position C (as shown in Fig. 1) is the same as that at B. The physical properties of PCMC used in this simulation are listed in Table 1. 4.2 °C was assumed in TEV system and 3.5 °C was used in orifice system.
3.1.2. PCMC stabilization The PCMC system also shows the benefits in terms of better system stability. Fig. 5 shows the attenuating effect the PCM has on the refrigerant temperature as it leaves the PCM and enters the compressor. As stated in previous paragraph PCMC allows the system a lower
Table 2 Cycle parameters varied with air inlet temperature of condenser Basic Air inlet temperature of condenser (°C) Inlet temperature of compressor (°C) Outlet temperature of compressor (°C) Evaporating temperature (°C) Condensing temperature (°C) Outlet temperature of condenser (°C) Outlet temperature of PCM (°C) Superheat (K) Sub-cooling (K) COP
16 16.5 125.5 10.4 39.6 32.3 6.1 7.3 2.38
PCMB 20 15.6 128.4 12.4 43.5 36 3.2 7.5 2.11
24 16.5 134.8 13.6 46.5 39.3 2.9 7.2 1.95
28 17.3 140.5 14.7 49.3 42.5 2.7 6.8 1.82
16 16.4 124.0 8.9 37.2 31.5 21.9 7.5 15.3 2.49
20 14.7 123.5 10.3 39.7 35.1 23.6 4.4 16.1 2.42
24 14.8 128.4 11.6 42.7 38.5 25.1 3.2 17.6 2.28
28 15.6 135.6 12.7 45.9 41.8 26.3 2.9 19.5 2.13
F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918
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2.7 2.65 2.6
COP
2.55 2.5 2.45 2.4
PCMC Orifice Basic PCMC_TEV
2.35 2.3 0
500
1000
1500
2000
2500
3000
3500
4000
Time (Sec)
Fig. 4. The simulated COP of the three systems.
evaporating temperature PCMC outlet temperature
6
the evaporator outlet temperature
5
Temperature (˚C)
4 3 2 1 0 0
500
1000
1500
-1
2000
2500
3000
3500
4000
Time (sec)
Fig. 5. PCMC stabilization.
superheat, and makes the temperature at the evaporator outlet has less amplitude than that produced by the basic system. This is because the gas in the superheated region is easily influenced by variations in the outside thermal load as its thermal capacity is much less than that of a liquid. Since the PCMC system operates with less gas in the evaporator the system is more stable. This attenuating effect could also be found with the compressor outlet temperature compared to the basic system because of less fluctuating inlet temperature to the compressor. 3.2. The investigation on PCMC heat exchanger— Orifice system In addition for use with TEV expansion devices the PCMC heat exchanger is potentially useful with orifice, capillary or short tube restriction systems which are found in many domestic systems. By positioning the PCMC heat exchanger as shown in Fig. 1, the heat
exchanger can be used to evaporate any liquid that exits in the evaporator under part load conditions. The theory of fixed orifice restrictors is that they compensate for part load conditions by restricting flow in response to thermal load. However their response lags that of the system with the penalty of low efficiency caused by artificially high running conditions during the lag period. The introduction of additional capacitance via a PCM heat exchanger allow the expansion device to operate in phase with the system and improve the overall system efficiency. To investigate this performance, the orifice system with the PCMC storage system has been simulated using the mathematical model. The geometric parameters inputted to the model are the same as the previous section with the exception of using an orifice restrictor in place of the TEV. The air inlet temperatures inputted to the model are similar to that used in the previous investigation although the amplitude of the thermal load fluctuation was increased slightly.
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F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918 PCMC_superheat
Basic_superheat
3.2 PCMA
3.1
10
3
8
2.9 COP
Superheat (°C)
12
6
Basic
2.8 2.7
4
2.6
2
2.5
0
2.4
0
500
1000
1500
2000 2500 Time (sec)
3000
3500
4000
0
500
Fig. 6. The superheat of the two systems.
4. PCMA reducing the condenser pressure 4.1. System performance Experimental investigation shows that the system COP was improved by 6% when the PCMA heat exchanger was incorporated into the basic system [2]. The mathematical model was also used to simulate the PCMA under various conditions. The geometric input parameters to the model are the same as in the previous Section 2.1. The air inlet temperature to the evaporator was fixed at 24 °C and to the condenser was an average of 25 °C with sine function of amplitude of 6 °C. The properties of PCMA are the same with PCMB properties. Fig. 7 shows a COP comparison for the two systems. The result analysis shows that the PCMA system COP improved by about 5.8%, and that both the COPs varied with the ambient temperatures. Fig. 8 shows the refrigerant temperatures at the PCMA inlet and outlet, and the temperature difference across the PCM heat exchanger. The temperature difference increases for the initial 200 s, then tends to decrease with time. For the initial 200 s, the refrigerant is changing phase and exchanges its latent heat with PCM. This can be clearly seen by
1500
2000 2500 Time (sec)
3000
3500
4000
Fig. 7. The COP of two systems.
120
Temperature (°C)
100 inlet temp. of PCMA
outlet temp. of PCMA
temp. difference of PCMA
condensing temp.
80 60 40 20 0 0
600
1200
1800 Time (sec)
2400
3000
3600
Fig. 8. The simulated temperatures at PCMA inlet and outlet.
the refrigerant outlet temperature being equal to the condensing temperature as shown in the diagram. After 200 s, the PCM wall temperature is higher than the condensing temperature and the refrigerant exchanges sensible heat with the PCM. The PCMA outlet temperature can be seen becoming progressively warmer as the PCM phase interface moves away from the wall and the thermal resistance to heat transfer increases. The reason for the improvement in COP is explained by Fig. 9 which shows the condenser pressures in PCMA system is lower than the basic system. This is because most of the desuperheating is done in the PCMA heat exchanger and this enables the condenser to operate more efficiently lowering the condensing temperature.
2100000 1800000 Pressure (Pa)
Fig. 4 also shows the effect of the PCMC heat exchanger on the COP of an orifice system. This diagram shows that the PCMC gives a greater potential improvement for the orifice system over the TEV system. The average superheat (2 K) in the orifice system is less than the superheat in the TEV system (3.8 K). This is because the TEV system uses the superheat to control the valve position and a finite superheat is always necessary. However, in the orifice system, a larger size restrictor can be used to control zero superheat as shown in Fig. 6. The lower superheat gives a higher COP (7% improvement) for the orifice system. Lower superheat also allows the compression process to follow a more acute entropy line which further reduces the work of compression. The combined effect of these factors is higher COP.
1000
1500000 1200000 900000
Pe_PCMA
Pc_PCMA
Pe_basic
Pc_basic
600000 300000 0
500
1000
1500
2000 2500 Time (sec)
3000
3500
Fig. 9. The condenser pressures of the two systems.
4000
F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918
4.2. Condenser pressure control
air inlet temp. of evap
air inlet temp. of cond_Fan
air inlet temp. of cond_no control
air inlet temp. of cond_PCM control
Temperature (°C)
25 20 15 10 5 0 500
1500
2500 3500 Time (sec)
4500
5500
Fig. 11. The inlet air temperatures of the condenser and the evaporator for three systems.
trolled system the condenser boundaries are the same except the lower ambient temperature region. When the fan-controlled system assumes the fan speed can be controlled to keep the condenser temperature constant. For the PCM-controlled system the condenser air inlet temperature is shown in Fig. 11. When the ambient temperature is very low, the condenser is closed and the second PCM is in operation, using the equivalent boundary the model now uses a modified air inlet temperature which accounts for the charging effect of the PCM heat exchanger.
Pc_fan_control Pc_PCMA_control Pc_no_control
1600000
COP_fan_contol COP_no control PCM_control
6
1500000
5.5
1400000
5
1300000
COP
1200000 1100000
4.5 4
1000000
3.5
900000
Time (sec)
Fig. 12. The condenser pressures and COPs for three systems.
Fig. 10. PCMA system investigation.
5000
4000
3000
2000
1000
5000
4000
3000
2000
0
Time (sec)
0
3
800000 1000
Pressure (Pa)
Many conventional refrigeration systems use mechanical condenser pressure controls, which are used to maintain the condenser pressure or condensing temperature above approximately 20 °C when the ambient temperature is below a finite value. This is necessary to ensure sufficient flow of liquid through the evaporator and is achieved using head pressure control valves or by cycling the condenser fans. The effect of mechanical pressure control is to limit the operation at lower head pressures by artificially elevating the condenser pressure. In UK this will normally occur when the ambient condition is at a temperature of 10 °C or below and this occurs for nearly 50% of the year. This artificial elevation of condenser pressure produces a large energy penalty. An alternative means of producing stability without inefficient mechanical delivery pressure controls is by using thermal storage. Fig. 10 shows the PCMA scheme used for this purpose. The first PCM heat exchanger (PCMA1) is located after the compressor. A second PCM heat exchanger is connected in parallel with the condenser in the system. Under low ambient conditions this system will maintain minimum condenser conditions by recharging the PCM heat exchanger. This will be achieved by diverting the flow via the PCMA2 heat exchanger. The existing mathematical model which includes the basic four components and a PCM model does not allow the condenser to be replaced by the PCMA2 in the middle of simulation. This is because it caused the condenser parameters to be lost and resulted in iteration divergence. A new derivation of the mathematical model is required to investigate this scheme by changing the boundary conditions in place of container changing. When the condenser temperature is less than the minimum set point temperature, the condenser boundary conditions were equivalently replaced by the PCMA2 heat exchanger boundary conditions. The simulations were carried on the same system with three different condenser boundaries. The condenser and the evaporator air inlet temperatures are shown in Fig. 11. For the uncontrolled system and the fan-con-
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F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918
Fig. 12 shows the results simulated by the method of boundary equivalence. The results clearly show that the COP for the PCM control option is considerably higher (20%) than the average COP for that without PCM control. This analysis assumes the daily diurnal variation occurs over 1 h. This simplification was made to reduce computation time and show in principal the savings that could be produced. In practice the average increase in COP will be limited by the storage capacity available and the thermal resistance between the PCM and refrigerant over the discharge and charge cycles. An alternative way of producing PCM condenser control is using a PCM heat exchanger within the condenser to control condensing temperatures around the PCM transition temperature. Further work is necessary to investigate the performance of this option.
5. Conclusions The novel application of PCMs in refrigeration systems at position A, B, C (as illustrated in Fig. 1) with a shell and tube structure have been investigated extensively by the novel mathematical model developed in authorsÕ previous work. For energy savings, PCMB is the suitable application and improves the system COP upon the phase transition temperature, which is determined by the climate. This improvement is up to 8% in the UK climate. At position C, the system COPs for the TEV and orifice systems are also improved by 4% and 7% respectively under any climate conditions. A better improvement for the orifice system was obtained because of a lower average superheat achieved. At position A and C further benefits such as system stabilization were also observed. For the PCMC system, the lower superheat minimizes the gas heat transfer region in the evaporator and therefore attenuates the fluctuation of the evaporator outlet temperature. Putting the PCM heat exchanger at position C further stabilizes the refrigerant temperature entering the compressor and reduces the peak compressor inlet tempera-
ture. The PCM heat exchanger at position A also helps the system to reduce and control the condenser pressure, thereby saving energy through a lower pressure ratio. Furthermore, condenser pressure control is also achieved by using a bypass PCM heat exchanger to control the minimum condenser temperature at low ambient temperatures. As this is usually achieved with mechanical condenser pressure controls, the application of PCMA to control head pressures will also save energy. The optimum size and structure of the PCM heat exchanger will depend upon the application. When used at position A, PCMA is depleted more quickly when used in other positions. The large relative size of PCM heat exchanger makes recharging take longer and be more difficult. A novel form of PCM heat exchange that will enhance recharging has been proposed [2]. With PCMC the additional suction line pressure drop that needs to be minimized whilst maximizing heat transfer between the PCM and the refrigerant. The optimum structure of this is the subject of further work. In particular it may be possible to integrate the PCM directly with the suction piping and insulation.
Acknowledgement The financial support of the Engineering Physical Science Research Council (EPSRC) is gratefully acknowledged.
References [1] F. Wang, G.G. Maidment, J.F. Missenden, R. Tozer, A review of research concerning the use of PCMs in air conditioning and refrigeration engineering, ABT 2002, Hong Kong, 2002. [2] F. Wang, The passive use of phase change materials (PCMs) in refrigeration system, Ph.D Thesis, London South Bank University, 2003. [3] K. Cornwell, Thermal optimisation of refrigeration evaporators, in: Proc. of the institute of refrigeration, 1992, pp. 1–11. [4] Z. Ure, Positive temperature eutectics TES system, AIRAH J. (2001) 31–35.
The novel use of phase change materials in refrigeration plant. Part 3: PCM for control and energy savings Fuqiao Wang a
a,*
, Graeme Maidment a, John Missenden a, Robert Tozer
b
London South Bank University, Faculty of Engineering, Science and the Built Environment, 103 Borough Road, SE1 0AA, London, UK b Waterman Gore-Mechanical and Electrical Consulting Engineers, Versailles Court, 3 Paris Garden, SE1 8ND, London, UK Received 23 August 2004; accepted 17 June 2005 Available online 10 May 2007
Abstract Phase change materials (PCMs) have been recognized as energy storage tanks since the 1980s. The ÔtankÕ has been introduced into the refrigeration system to enable its capacitance to take account of fluctuations in the daily cooling load. However, this part of the paper will present a novel control purpose of using PCM in refrigeration systems. The novel application of PCMs in refrigeration systems at different positions in the refrigeration cycle circuit with a shell and tube structure has been investigated extensively by the novel mathematical model presented in part 2 of the paper. The results show that for energy savings, PCMs at different positions give coefficient of performances (COPs) up to 8% through lowering the sub-cooling. PCMs also improve the system COPs up to 4% and 7% for the thermostatic expansion valve (TEV) and orifice systems, respectively, by minimizing the superheat. Further benefits such as system stabilization were also observed in this investigation. Ó 2005 Elsevier Ltd. All rights reserved. Keywords: Phase change materials; Refrigeration; Air conditioning; Control; Thermal energy storage
1. Introduction 1.1. The different purpose of using PCM from traditional usage The traditional use of PCM in refrigeration system is of a cold storage tank. During Ôoff-peakÕ periods, the refrigeration system produces cooling for the PCM storage system. The stored cooling energy will be released during peak periods to assist the refrigeration system to meet the peak load. PCM storage system used as a storage tank achieves smaller size chiller, lower running cost and higher efficiency. A recent development in the application of PCMs in refrigeration cycles is shown in Fig. 1 [1]. This improves *
Corresponding author. Tel.: +44 20 78157634; fax: +44 20 78157699. E-mail address: [email protected] (F. Wang). 1359-4311/$ - see front matter Ó 2005 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2005.06.010
cycle control stability and also results in large energy savings. Compared with a conventional system higher COP can be achieved by increasing sub-cooling by PCMB, minimizing superheat by PCMC and reducing condenser pressure by PCMA. These novel methods of using PCMs for controlling purpose can be introduced together in traditional vapour compressor refrigeration systems. 1.2. The benefits to the system The primary benefit of increased sub-cooling is higher COP. For example, a 15 K increase in sub-cooling levels has been reported to lead to approximately a 10% improvement in peak plant efficiency [4]. A refrigeration system incorporating a phase change material in contact with the liquid lines (as shown in Fig. 1) enables the refrigeration system to benefit from lower night time ambient conditions. At night lower temperatures
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F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918
Nomenclature COP Pc Pe PCM PCMA
coefficient of performance pressure in the condenser (Pa) pressure in the evaporator (Pa) phase change material PCM heat exchanger located between the condenser and the compressor
PCMB PCM heat exchanger between the condenser and the TEV PCMC PCM heat exchanger between the evaporator and the compressor t temperature (°C) TEV thermal expansion valve
Fig. 1. PCM heat exchanger used individually at different positions in the refrigeration cycle.
generate lower condenser pressures (for the case of PCMA) as well as increased liquid sub-cooling (for the case of PCMB). The use of a PCM in the liquid line then provides an artificial cooling load on the condenser during daytime, which will avoid the necessity for inefficient condenser control, as well as allowing the night time cold energy to be stored for release during daytime operation. At the exit from the evaporator the refrigerant vapour must be dry to avoid droplet damage in the compressor and most systems are designed to produce superheated refrigerant to ensure this. The degree of superheat is usually controlled by a thermostatic expansion valve (TEV), which is a proportional controller responding to the control signal, the degree of superheated vapour at the evaporator outlet. In practice an undesirable oscillating behaviour of the TEV control known as ÔhuntingÕ occurs when a low superheat is applied. In extreme cases, the TEV operates as an on/off controller. The use of higher superheat can reduce the unstable operating range. However, this reduces the efficiency of the evaporator. A superheat setting of 5 K is often cited as an efficient compromise, but this fluctuates with load, and many refrigeration systems require higher superheat settings for stable operation. As a consequence, up to 20% of the heat exchanger surface area is required to transfer heat to the dry vapour thus using it ineffectively [3]. The higher the superheat is, the lower the efficiency of the system is. This is in part due to a lower evaporation temperature which increases the compression pressure ratio and work of compression, however it also causes the compression process to begin
further in the superheat zone and this causes the isentropic enthalpy change across the compressor to be greater. The use of a suction line PCM heat exchanger overcomes this by minimizing the degree of superheat in the suction line. The purpose of the suction line PCM exchanger is to fix the superheat in the suction line by providing thermal capacitance and enabling the evaporator to operate approaching wet, which will further improve system efficiency. In order to investigate quantitatively the benefits achieved by the novel use of PCM in the system, a validated mathematical model was used. The details of the mathematical model can be found in part 2 of the paper. Here the results of these novel ideas which include the use of PCM heat exchangers in different position as illustrated in Fig. 1 are introduced and discussed. The mathematical model was based on the following specific plant. 1.3. System to be simulated The test plant is based on an IMI Marstair impact cooling split system model C60/E. The indoor unit comprises an evaporator (IMI impact CU6E) and a fan. The outdoor unit (IMI CUE60) is comprised of a condenser, a compressor and a propeller fan. The details of the plant geometric parameters can be referred from part 1 of the paper. This paper will exam the benefits of the novel uses of PCM heat exchangers as shown in Fig. 1 individually by using the model that is developed in part 2 of this paper, and also discuss some parameters which affect the per-
F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918 3.2
COP
3
2. PCMB increasing the sub-cooling
2.9 2.8
During the simulation, PCM heat exchanger B was assumed to be charged with E21 solution that is made by EPS Ltd Company. This is Grauber salt eutectic solutions complete with nucleating agent, gelling agent and stabilizers. However, the bulk of the solution is Grauber salt. Its physical properties are listed in Table 1. The heat exchanger effective length is 4.32 m and its diameter is 110 mm. In the simulation, only the melt temperature was considered to change from 16 °C to 28 °C, the other parameters were fixed. Mathematical investigation shows that different phase change temperatures have different influences on the system. The results are shown in Fig. 2 which assumes the air inlet temperature to the condenser was fixed at 30 °C, and the air inlet temperature to the evaporator was fixed at 24 °C. The system was charged of
Table 1 PCM materials physical properties Fusion heat (kJ/kg)
Transition temperature (°C)
Thermal conductivity (W/m K)
Density (kg/m3)
150 140
21 4.2/3.5
0.43 0.44
1480 1469
35
tpcmb=16 tpcmb=19 tpcmb=22 tpcmb=25 tpcmb=28
30 25 20 15 10
2.7
2.1. Different phase transition temperatures with different sub-cooling
5 0
2.6 0
1000
2000 3000 Time (sec)
4000
0
0.04
0.03
0.02
0.01
16_mid_discharge 28_mid_discharge 16_mid_charge 28_mid_charge
16_end_discharge 28_end_discharge 16_end_charge 28_end_charge
0 0
500
1000
1500
2000 3000 Time (sec)
4000
3200 g of Freon22(R22) in the model, and the geometric sizes of the heat exchangers including the PCMB heat exchanger and TEV system were fixed. Fig. 2 shows the COP result from this investigation and identifies how the system COP varies with the PCMB temperature and time. It can be clearly seen that a higher COP can be achieved by lowering the phase change temperature. This is because greater sub-cooling is achieved with lower phase change temperature as also shown in Fig. 2. However, the phase change temperature will also affect the ability to recharge the PCM. Fig. 3 compares two PCMB types with different phase transition temperatures, namely 16 °C and 28 °C. Their phase interfaces moving against time are shown for the discharge (melt) process. The lower temperature PCMB can be seen to discharge faster than the higher temperature version. Fig. 3 also compares charging (solidification) of the same two PCMB systems in an environment of 12 °C, it can be seen that the lower temperature PCMB recharges more slowly than the higher PCMB option. It can be seen that the lower temperature PCMB takes approximately 4 times as long to recharge
0.05
16_entrance_discharge 28_entrance_discharge 16_entrance_charge 28_entrance_charge
1000
Fig. 2. System COP and the sub-cooling various with PCMB temperature and time.
0.06
Phase-interface position diameter (m)
PCMA&B PCMC
tpcmb=28 tpcmb=25 tpcmb=22 tpcmb=19 tpcmb=16
3.1
Sub-cooling (˚C)
formance of these three systems. The numerical investigation initially starts with the PCM heat exchanger at position B, then at C, and finally at A as shown in Fig. 1.
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2000
2500
3000
3500
4000
Time (sec)
Fig. 3. PCMB phase interface moving against time during charging/discharging period.
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F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918
3.1. The investigation on PCMC system—TEV system
than to discharge. The recharge usually takes place at night when low ambient temperatures occur. The ability of the PCM to recharge will depend upon the night time temperature, recharge duration and phase change temperature. Lower phase change temperature may improve COP but may limit recharge and therefore careful selection is important for all applications.
3.1.1. The system COP Two systems were simulated in this part: the PCMC system as shown in Fig. 1 and the basic system without PCMC heat exchanger. The plant geometric size is the same as discussed in Section 1.3. For this analysis, a 1630 g R22 initial charge and 5 K superheat was used for the PCMC system, and a 1530 g R22 charge and 10 K superheat was used for the basic system. Compare to basic system, more refrigerant is charged and less superheat is assumed in PCMC system, because when PCMC is integrated into the system results an extra volume charged with gas, this will affect the results if the initial charge does not change, this has been explained in this paper, Part 1. Less superheat is used; it is the main benefit when PCMC is introduced into the system. The air inlet temperatures used for both the two condensers and the evaporators are assumed to be consistent. The air inlet temperature of the condenser was fixed at 30 °C, however, the input air temperature to the evaporator was an average 26 °C. To simulate the effect of load variation the inlet air temperature to the evaporator was allowed to fluctuate as a sine wave with amplitude of 4°. Fig. 4 shows that the PCMC system produced an average COP which is about 4% higher than that of basic system. The superheat in PCMC system is about 5 K lower than that of the basic system. Lower superheat allows higher COP, because it allows the evaporator to operate with a boiling heat transfer coefficient for more of its surface. This improves heat transfer and allows higher evaporator pressure and this reduces the work input requirement. Lower superheat also reduces the work of compression since the constant entropy lines are less acute closer to the saturation line.
2.2. Environment temperature influence on system performance For vapour compression refrigeration system, the higher the environmental temperature is, the lower the system COP produced. However, for the PCM system, the situation is slightly different. When a higher environmental temperature is applied to the PCM system, the higher condenser temperature will tend to increase the temperature difference between the refrigerant and the PCM phase change temperature, allowing greater energy to be transferred to the PCM heat exchanger. The sub-cooling effect of the PCMB will therefore be greater at higher ambient temperatures, which will attenuate the negative effect due to higher condenser temperature. Table 2 shows the simulation results. It can be clearly seen that the COP of the PCM system decreases by 14.56%, i.e. from 2.49 to 2.13, when the environment temperature increases from 16 °C to 28 °C. However, the COP of the basic system (without PCM) decreases by 19.96% from a value of 2.28 to 1.82. This shows that the higher the environment temperature, the greater are the savings that the PCM system brings.
3. PCMC minimizing the superheat Basically there are two expansion devices-thermostatic expansion valve and orifice in refrigeration system; therefore the two systems were simulated in this section. The geometry of the heat exchangers at position C (as shown in Fig. 1) is the same as that at B. The physical properties of PCMC used in this simulation are listed in Table 1. 4.2 °C was assumed in TEV system and 3.5 °C was used in orifice system.
3.1.2. PCMC stabilization The PCMC system also shows the benefits in terms of better system stability. Fig. 5 shows the attenuating effect the PCM has on the refrigerant temperature as it leaves the PCM and enters the compressor. As stated in previous paragraph PCMC allows the system a lower
Table 2 Cycle parameters varied with air inlet temperature of condenser Basic Air inlet temperature of condenser (°C) Inlet temperature of compressor (°C) Outlet temperature of compressor (°C) Evaporating temperature (°C) Condensing temperature (°C) Outlet temperature of condenser (°C) Outlet temperature of PCM (°C) Superheat (K) Sub-cooling (K) COP
16 16.5 125.5 10.4 39.6 32.3 6.1 7.3 2.38
PCMB 20 15.6 128.4 12.4 43.5 36 3.2 7.5 2.11
24 16.5 134.8 13.6 46.5 39.3 2.9 7.2 1.95
28 17.3 140.5 14.7 49.3 42.5 2.7 6.8 1.82
16 16.4 124.0 8.9 37.2 31.5 21.9 7.5 15.3 2.49
20 14.7 123.5 10.3 39.7 35.1 23.6 4.4 16.1 2.42
24 14.8 128.4 11.6 42.7 38.5 25.1 3.2 17.6 2.28
28 15.6 135.6 12.7 45.9 41.8 26.3 2.9 19.5 2.13
F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918
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2.7 2.65 2.6
COP
2.55 2.5 2.45 2.4
PCMC Orifice Basic PCMC_TEV
2.35 2.3 0
500
1000
1500
2000
2500
3000
3500
4000
Time (Sec)
Fig. 4. The simulated COP of the three systems.
evaporating temperature PCMC outlet temperature
6
the evaporator outlet temperature
5
Temperature (˚C)
4 3 2 1 0 0
500
1000
1500
-1
2000
2500
3000
3500
4000
Time (sec)
Fig. 5. PCMC stabilization.
superheat, and makes the temperature at the evaporator outlet has less amplitude than that produced by the basic system. This is because the gas in the superheated region is easily influenced by variations in the outside thermal load as its thermal capacity is much less than that of a liquid. Since the PCMC system operates with less gas in the evaporator the system is more stable. This attenuating effect could also be found with the compressor outlet temperature compared to the basic system because of less fluctuating inlet temperature to the compressor. 3.2. The investigation on PCMC heat exchanger— Orifice system In addition for use with TEV expansion devices the PCMC heat exchanger is potentially useful with orifice, capillary or short tube restriction systems which are found in many domestic systems. By positioning the PCMC heat exchanger as shown in Fig. 1, the heat
exchanger can be used to evaporate any liquid that exits in the evaporator under part load conditions. The theory of fixed orifice restrictors is that they compensate for part load conditions by restricting flow in response to thermal load. However their response lags that of the system with the penalty of low efficiency caused by artificially high running conditions during the lag period. The introduction of additional capacitance via a PCM heat exchanger allow the expansion device to operate in phase with the system and improve the overall system efficiency. To investigate this performance, the orifice system with the PCMC storage system has been simulated using the mathematical model. The geometric parameters inputted to the model are the same as the previous section with the exception of using an orifice restrictor in place of the TEV. The air inlet temperatures inputted to the model are similar to that used in the previous investigation although the amplitude of the thermal load fluctuation was increased slightly.
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Basic_superheat
3.2 PCMA
3.1
10
3
8
2.9 COP
Superheat (°C)
12
6
Basic
2.8 2.7
4
2.6
2
2.5
0
2.4
0
500
1000
1500
2000 2500 Time (sec)
3000
3500
4000
0
500
Fig. 6. The superheat of the two systems.
4. PCMA reducing the condenser pressure 4.1. System performance Experimental investigation shows that the system COP was improved by 6% when the PCMA heat exchanger was incorporated into the basic system [2]. The mathematical model was also used to simulate the PCMA under various conditions. The geometric input parameters to the model are the same as in the previous Section 2.1. The air inlet temperature to the evaporator was fixed at 24 °C and to the condenser was an average of 25 °C with sine function of amplitude of 6 °C. The properties of PCMA are the same with PCMB properties. Fig. 7 shows a COP comparison for the two systems. The result analysis shows that the PCMA system COP improved by about 5.8%, and that both the COPs varied with the ambient temperatures. Fig. 8 shows the refrigerant temperatures at the PCMA inlet and outlet, and the temperature difference across the PCM heat exchanger. The temperature difference increases for the initial 200 s, then tends to decrease with time. For the initial 200 s, the refrigerant is changing phase and exchanges its latent heat with PCM. This can be clearly seen by
1500
2000 2500 Time (sec)
3000
3500
4000
Fig. 7. The COP of two systems.
120
Temperature (°C)
100 inlet temp. of PCMA
outlet temp. of PCMA
temp. difference of PCMA
condensing temp.
80 60 40 20 0 0
600
1200
1800 Time (sec)
2400
3000
3600
Fig. 8. The simulated temperatures at PCMA inlet and outlet.
the refrigerant outlet temperature being equal to the condensing temperature as shown in the diagram. After 200 s, the PCM wall temperature is higher than the condensing temperature and the refrigerant exchanges sensible heat with the PCM. The PCMA outlet temperature can be seen becoming progressively warmer as the PCM phase interface moves away from the wall and the thermal resistance to heat transfer increases. The reason for the improvement in COP is explained by Fig. 9 which shows the condenser pressures in PCMA system is lower than the basic system. This is because most of the desuperheating is done in the PCMA heat exchanger and this enables the condenser to operate more efficiently lowering the condensing temperature.
2100000 1800000 Pressure (Pa)
Fig. 4 also shows the effect of the PCMC heat exchanger on the COP of an orifice system. This diagram shows that the PCMC gives a greater potential improvement for the orifice system over the TEV system. The average superheat (2 K) in the orifice system is less than the superheat in the TEV system (3.8 K). This is because the TEV system uses the superheat to control the valve position and a finite superheat is always necessary. However, in the orifice system, a larger size restrictor can be used to control zero superheat as shown in Fig. 6. The lower superheat gives a higher COP (7% improvement) for the orifice system. Lower superheat also allows the compression process to follow a more acute entropy line which further reduces the work of compression. The combined effect of these factors is higher COP.
1000
1500000 1200000 900000
Pe_PCMA
Pc_PCMA
Pe_basic
Pc_basic
600000 300000 0
500
1000
1500
2000 2500 Time (sec)
3000
3500
Fig. 9. The condenser pressures of the two systems.
4000
F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918
4.2. Condenser pressure control
air inlet temp. of evap
air inlet temp. of cond_Fan
air inlet temp. of cond_no control
air inlet temp. of cond_PCM control
Temperature (°C)
25 20 15 10 5 0 500
1500
2500 3500 Time (sec)
4500
5500
Fig. 11. The inlet air temperatures of the condenser and the evaporator for three systems.
trolled system the condenser boundaries are the same except the lower ambient temperature region. When the fan-controlled system assumes the fan speed can be controlled to keep the condenser temperature constant. For the PCM-controlled system the condenser air inlet temperature is shown in Fig. 11. When the ambient temperature is very low, the condenser is closed and the second PCM is in operation, using the equivalent boundary the model now uses a modified air inlet temperature which accounts for the charging effect of the PCM heat exchanger.
Pc_fan_control Pc_PCMA_control Pc_no_control
1600000
COP_fan_contol COP_no control PCM_control
6
1500000
5.5
1400000
5
1300000
COP
1200000 1100000
4.5 4
1000000
3.5
900000
Time (sec)
Fig. 12. The condenser pressures and COPs for three systems.
Fig. 10. PCMA system investigation.
5000
4000
3000
2000
1000
5000
4000
3000
2000
0
Time (sec)
0
3
800000 1000
Pressure (Pa)
Many conventional refrigeration systems use mechanical condenser pressure controls, which are used to maintain the condenser pressure or condensing temperature above approximately 20 °C when the ambient temperature is below a finite value. This is necessary to ensure sufficient flow of liquid through the evaporator and is achieved using head pressure control valves or by cycling the condenser fans. The effect of mechanical pressure control is to limit the operation at lower head pressures by artificially elevating the condenser pressure. In UK this will normally occur when the ambient condition is at a temperature of 10 °C or below and this occurs for nearly 50% of the year. This artificial elevation of condenser pressure produces a large energy penalty. An alternative means of producing stability without inefficient mechanical delivery pressure controls is by using thermal storage. Fig. 10 shows the PCMA scheme used for this purpose. The first PCM heat exchanger (PCMA1) is located after the compressor. A second PCM heat exchanger is connected in parallel with the condenser in the system. Under low ambient conditions this system will maintain minimum condenser conditions by recharging the PCM heat exchanger. This will be achieved by diverting the flow via the PCMA2 heat exchanger. The existing mathematical model which includes the basic four components and a PCM model does not allow the condenser to be replaced by the PCMA2 in the middle of simulation. This is because it caused the condenser parameters to be lost and resulted in iteration divergence. A new derivation of the mathematical model is required to investigate this scheme by changing the boundary conditions in place of container changing. When the condenser temperature is less than the minimum set point temperature, the condenser boundary conditions were equivalently replaced by the PCMA2 heat exchanger boundary conditions. The simulations were carried on the same system with three different condenser boundaries. The condenser and the evaporator air inlet temperatures are shown in Fig. 11. For the uncontrolled system and the fan-con-
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F. Wang et al. / Applied Thermal Engineering 27 (2007) 2911–2918
Fig. 12 shows the results simulated by the method of boundary equivalence. The results clearly show that the COP for the PCM control option is considerably higher (20%) than the average COP for that without PCM control. This analysis assumes the daily diurnal variation occurs over 1 h. This simplification was made to reduce computation time and show in principal the savings that could be produced. In practice the average increase in COP will be limited by the storage capacity available and the thermal resistance between the PCM and refrigerant over the discharge and charge cycles. An alternative way of producing PCM condenser control is using a PCM heat exchanger within the condenser to control condensing temperatures around the PCM transition temperature. Further work is necessary to investigate the performance of this option.
5. Conclusions The novel application of PCMs in refrigeration systems at position A, B, C (as illustrated in Fig. 1) with a shell and tube structure have been investigated extensively by the novel mathematical model developed in authorsÕ previous work. For energy savings, PCMB is the suitable application and improves the system COP upon the phase transition temperature, which is determined by the climate. This improvement is up to 8% in the UK climate. At position C, the system COPs for the TEV and orifice systems are also improved by 4% and 7% respectively under any climate conditions. A better improvement for the orifice system was obtained because of a lower average superheat achieved. At position A and C further benefits such as system stabilization were also observed. For the PCMC system, the lower superheat minimizes the gas heat transfer region in the evaporator and therefore attenuates the fluctuation of the evaporator outlet temperature. Putting the PCM heat exchanger at position C further stabilizes the refrigerant temperature entering the compressor and reduces the peak compressor inlet tempera-
ture. The PCM heat exchanger at position A also helps the system to reduce and control the condenser pressure, thereby saving energy through a lower pressure ratio. Furthermore, condenser pressure control is also achieved by using a bypass PCM heat exchanger to control the minimum condenser temperature at low ambient temperatures. As this is usually achieved with mechanical condenser pressure controls, the application of PCMA to control head pressures will also save energy. The optimum size and structure of the PCM heat exchanger will depend upon the application. When used at position A, PCMA is depleted more quickly when used in other positions. The large relative size of PCM heat exchanger makes recharging take longer and be more difficult. A novel form of PCM heat exchange that will enhance recharging has been proposed [2]. With PCMC the additional suction line pressure drop that needs to be minimized whilst maximizing heat transfer between the PCM and the refrigerant. The optimum structure of this is the subject of further work. In particular it may be possible to integrate the PCM directly with the suction piping and insulation.
Acknowledgement The financial support of the Engineering Physical Science Research Council (EPSRC) is gratefully acknowledged.
References [1] F. Wang, G.G. Maidment, J.F. Missenden, R. Tozer, A review of research concerning the use of PCMs in air conditioning and refrigeration engineering, ABT 2002, Hong Kong, 2002. [2] F. Wang, The passive use of phase change materials (PCMs) in refrigeration system, Ph.D Thesis, London South Bank University, 2003. [3] K. Cornwell, Thermal optimisation of refrigeration evaporators, in: Proc. of the institute of refrigeration, 1992, pp. 1–11. [4] Z. Ure, Positive temperature eutectics TES system, AIRAH J. (2001) 31–35.