Experimental investigations on the use of capillary tube and thermostatic expansion valve in storage-enhanced heat recovery room air-conditioner

Energy and Buildings 101 (2015) 76–83

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Experimental investigations on the use of capillary tube and thermostatic expansion valve in storage-enhanced heat recovery room air-conditioner Jie Jia, W.L. Lee ∗ Department of Building Services Engineering, The Hong Kong Polytechnic University, Hung Hom, Hong Kong

a r t i c l e

i n f o

Article history: Received 13 December 2014 Received in revised form 26 March 2015 Accepted 3 May 2015 Available online 11 May 2015 Keywords: Heat recovery Air-conditioner Capillary tube Thermostatic expansion valve Space cooling Water heating

a b s t r a c t The use of storage-enhanced heat recovery room air-conditioner (SEHRAC) to supply hot water and provide space cooling for energy saving has been confirmed effective, but the heat recovery process can lead to significant refrigerant pressure fluctuations. Whether an expansion device can function properly despite the fluctuations is decisive to the successful use of SEHRAC. Capillary tube (CT) and thermostatic expansion valve (TEV) are the two most used expansion devices in SEHRAC, but a suitable type has never been identified in extant literature. In this study, laboratory experiments were conducted to enable a side-by-side comparison of the two expansion devices. A prototype SEHRAC, which could be switched between the CT and TEV systems, was designed and setup in a test facility for two identical sets of experiments. For each set of experiments, a series of tests under different outdoor temperatures were conducted. It was found that an overall better performance was achieved by the TEV system. The space cooling and water heating capacities were on average 16.3–19.4% and 18.5–23.4% larger than the CT system. The overall coefficient of performance was also found 12.5–20.9% higher. The results confirmed that TEV is preferred for use in SEHRAC. © 2015 Elsevier B.V. All rights reserved.

1. Introduction Air-conditioning and domestic water heating are two dominant electricity end-uses in residential buildings. In Hong Kong, out of the total energy consumption of the residential sector, 23% goes to air-conditioning and 19% goes to water heating [1]. These figures clearly show that a technology that can simultaneously reduce energy use for space cooling and water heating is highly promising. Recent research efforts therefore focused on investigating the energy saving potential of a storage-enhanced heat recovery room air-conditioner (SEHRAC). SEHRAC is a conventional air-conditioner provided with an additional tank-immersed refrigerant-to-water heat exchanger between the compressor and the air-cooled condenser for recovery of condenser heat for water heating. Whether the use of SEHRAC can supply hot water and provide space cooling for energy saving was investigated and confirmed by Ying [2] in Singapore and Monerasinghe et al. [3] in Malaysia. Its

∗ Corresponding author. Tel.: +852 27665852. E-mail address: [email protected] (W.L. Lee). http://dx.doi.org/10.1016/j.enbuild.2015.05.007 0378-7788/© 2015 Elsevier B.V. All rights reserved.

use in Hong Kong was examined by the authors based on a series of simulations and experimental tests. In the simulation study [4], the effectiveness of SEHRAC in satisfying the daily space cooling and water heating demands of households residing in a typical residential estate in Hong Kong was confirmed. In the experimental study [5], the energy performance of SEHRAC in comparison with conventional air-conditioner under different outdoor temperature and indoor cooling load conditions was examined. A mathematical model for evaluating the performance of SEHRAC was also developed. Another mathematical model for achieving the overall system optimization was constructed by Techarungpaisan et al. [6]. Apart from mathematical models, recent researches also focused on incorporating SEHRAC with other control devices to enable its all-year round operation [7,8]. In a SEHRAC, same as other refrigeration systems, expansion device is one of the most important components. It functions to regulate the refrigerant flow to the evaporator and to maintain a pressure difference between the high and low pressure sides. Failure to function will be detrimental including refrigerant flooding to damage the compressor, refrigerant starving to lower the evaporator effectiveness, and hunting to cause large fluctuations in evaporator superheat. As operation of SEHRAC can lead to significant

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Nomenclature cp C COP h m P Q t T V w W x

specific heat capacity (kJ/kg K) regression coefficient coefficient of performance enthalpy (kJ/kg) mass flow rate (kg/s) pressure (Pa) space cooling capacity/water heating capacity (kW) temperature (◦ C) absolute temperature (K) volume (m3 ) moisture content (kg/kg dry air) power consumption (kW) refrigerant vapor content (kg/kg)

Greek symbol ϕ relative humidity density (kg/m3 )   operating time (s) Subscript a air b atmospheric cl space cooling e evaporator fg latent heat ir inlet refrigerant overall oa r return s supply sl saturated liquid refrigerant saturated water vapor sw w water wh water heating

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Farzad and O’Neal [14]. It was found that TEV can better regulate the refrigerant flow to meet the cooling demand at off-design conditions, but CT also performs fairly well in most conditions. It was also reported that the feasible use of TEV is restricted to some systems because of the valve instability problem. On this issue, Maidment et al. [15] evaluated the use of CT and TEV in a CRAC serving a supermarket’s display cabinet. It was found that CT is better than TEV in energy performance and in providing a more stable operating condition. It is noted from the above that whether CT or TEV is better depends largely on the associated operation characteristics. Unfortunately, these studied results are not applicable to SEHRAC because its operation characteristics differ largely from HPHW and CRAC. In a HPHW, the water delivered to the condenser is usually tap water with very little variation in temperature throughout the year [16]. Similar condition also applies to CRAC in the cooling season. While for SEHRAC, due to the heat recovery occurred in the storage tank, the condenser water temperature is highly fluctuating, varying between 25 ◦ C and 55 ◦ C. Given relevant research works are virtually none, to identify whether CT or TEV is preferred for use in SEHRAC, a prototype equipped with the two expansion devices was setup for experimental studies. In the prototype, the CT and TEV could be alternatively activated by the gate valves provided, thus enabling the direct comparison of the two systems (CT and TEV systems) under the same experimental conditions. In the experiments, the operating parameters on the refrigerant, air and water sides of the systems were monitored. Based on the experimental results, the performances of the two systems were determined and compared. Given the equipment capacity is small and the refrigerant properties are highly volatile, monitoring of the refrigerant side performance is difficult. Thus the results presented in this study and the experimental protocol established will also be useful for future development of SEHRAC for energy saving.

2. Experimental study pressure fluctuations on the refrigerant side [5], the selection of a suitable type of expansion device that can function properly despite the fluctuations is very important. Within the capacity range of SEHRAC, capillary tube (CT) and thermostatic expansion valve (TEV) are most used. CT is a length of small diameter tube and is therefore considered as a fixed type expansion device. As opposed to CT, TEV can actively regulate the refrigerant flow by varying its opening extent according to feedback from a sensing bulb. Thus a question is raised whether CT or TEV is preferred for use in SEHRAC. However, previous relevant works mainly focused on evaluating the use of CT and TEV in heat pump water heaters (HPWHs) and conventional room air-conditioners (CRACs). Choi and Kim [9] investigated the influence of different expansion devices on the performance of a HPWH under a range of operating conditions. The experimental results concluded that the TEV system has an overall performance better than the CT system. Similar conclusion was also drawn by Kim and O’Neal [10] and Zhang et al. [11]. However, Agrawal and Bhattacharyya [12] came up with a different conclusion. They adopted a detailed steady state simulation model to evaluate the performance of a transcritical CO2 HPWH. It was found that with the use of a proper CT, the HPWH can operate efficiently under a wide range of condenser inlet water temperatures. It was also concluded that the performance of the CT system is on a par with the TEV system. For the use of CT and TEV in CRACs, Stoecker et al. [13] found that by replacing the CT with a TEV, the overall energy efficiency of a CRAC can be improved by 2–3%. A similar study was conducted by

2.1. Experimental setup The experimental study was conducted in a test facility located at the Hong Kong Polytechnic University. It comprised two environmental chambers resembling indoor and outdoor conditions, which were completely insulated and separated. Each chamber measured 3.9 m in length, 3.8 m in width and 2.2 m in height, which was constructed in conformity with ASHRAE 16 [17] for rating of room air-conditioners. Given the same test facility was used in an earlier study by the authors on evaluating the energy performance of SEHRAC in comparison with conventional air-conditioner [5], to avoid duplication, only brief descriptions of the test facility are presented below. The outdoor chamber was conditioned by a built-in airconditioning system (BIACS) and was provided with a set of sensible and latent load generation units (LGUs). The indoor chamber was provided with another set of LGUs. A computer supervisory program was adopted to regulate the outputs of the LGUs to maintain the pre-set temperatures and humidities in the two chambers. The experimental conditions and equipment capacities were set according to findings of an earlier study [4]. The schematic of the prototype SEHRAC is shown in Fig. 1. It mainly comprised a split-type air-cooled air-conditioner, a refrigerant-to-water heat exchanger and a cylindrical water storage tank. The indoor unit was basically a direct expansion evaporator with copper tubes and aluminum fins. The outdoor unit consisted of an air-cooled condenser and a rotary type compressor. There was no capacity control provided for the compressor. The CT and TEV were

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Fig. 1. Schematic of the prototype SEHRAC (T: Thermocouple; P: Pressure transmitter; H: Hygrometer; HX: Heat exchanger; SG: Sight glass; V: Gate valve).

parallel-connected on a manifold feeding the evaporator. They could be alternatively activated by the gate valves provided. This design enabled carrying out comparative tests on their influence on the system performance. The heat exchanger was a tailor-made helical coil. It was immersed vertically in the storage tank with top inlet to cater for the density change of vapor-liquid refrigerant. In the prototype, the evaporator functions to enable heat absorption of liquid refrigerant from the indoor chamber to become vapor refrigerant. An accumulator is provided at the evaporator outlet to ensure that only pure vapor refrigerant is supplied to the compressor where the pressure and temperature of the vapor refrigerant is elevated. The high pressure high temperature refrigerant, on discharged from the compressor, is supplied to the heat exchanger inside the storage tank for water heating. On leaving the heat exchanger, the refrigerant is further cooled in the condenser located in the outdoor chamber and then dried in a drier before it flows through the expansion devices (CT or TEV), where its pressure and temperature are reduced to become liquid refrigerant for another cycle of cooling in the indoor chamber. The detailed specifications of the prototype are summarized in Table 1. The air-conditioner had a nominal cooling capacity of 2.5 kW. The original CT in the air-conditioner was used in this study. The tested TEV was a commercially available Danfoss TX2 internally equalized expansion valve with an orifice number of 01. Its sensing bulb was attached to a horizontal section of the suction line near the evaporator outlet. The CT and TEV as well as the connecting piping were insulated from surroundings by using non-hydroscopic insulation material. The amount of refrigerant charged into the system was determined according to the actual piping length and the manufacturer’s recommendations. The total volume of water to be heated in the storage tank was 130 L (same as an earlier study [5]). The tank was insulated with rock wool wrapped with a layer of aluminum foil with a total thickness of 50 mm.

2.2. Experimental measurements The specifications of the measuring instruments used in this study are summarized in Table 2. In the experiments, the operating

Table 1 Specifications of the prototype SEHRAC. Split-type air-cooled air-conditioner Model Nominal cooling capacity Nominal power consumption Compressor Evaporator and condenser Expansion device

Working fluid

Mitsubishi SRK09CMP 2.5 kW 0.925 kW Hermetic and rotary type Aluminum fins and copper tubes Capillary tube (ID = 1.4 mm; OD = 2.5 mm; Length = 720 mm) Thermostatic expansion valve (Danfoss TX2; Orifice number: 01) R22

Heat exchanger Material Number of turns Total length

Copper 16 15.6 m

Water storage tank Size Insulation layer

550 mm × 590 mm (Diameter × Height) Rock wool (Thickness = 50 mm; Thermal conductivity = 0.038 W/m K)

parameters on the refrigerant, air and water sides of the systems were monitored in conformity to the measurement methods recommended. The refrigerant temperatures were measured by calibrated thermocouples (K type). The refrigerant high and low side pressures were measured by pressure transmitters with different ranges to minimize the measurement uncertainty. The refrigerant flow rate was measured by a variable area flow meter installed between the condenser and the expansion devices. The pressure drop across the flow meter was measured (=2.1 kPa), which is much lower than the allowable limit (82.7 kPa) recommended by ASHRAE 116 [22] to fulfill the accuracy requirement. The air conditions entering and leaving the evaporator were measured by hygrometers. The supply air flow rate was measured by thermal anemometers positioned according to the Log Tchebycheff rule. The water temperatures at different axial and radial locations of the storage tank were

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Table 2 Summary of measuring instruments and corresponding measurement standards. Instrument

Model

Range

Accuracy

Measurement standard

Hygrometer

ROTRONIC HygroFlex

±0.1 ◦ C±1% RH

BS 5141 [18]

Thermal anemometer Pt100 RTD Thermocouple Pressure transmitter

E+E 70-VT62B5 Omega 1/10 DIN Omega TT-K-24 Danfoss AKS 32 and 33

±2% F.S. ±0.05 ◦ C ±0.1 ◦ C ±0.3% F.S.

ASHRAE 41.2 [19] ASHRAE 41.1 [20] ASHRAE 41.1 [20] ASHRAE 41.3 [21]

Variable area flow meter Power quality tester

H250/RR/M9 ISO-TECH IPM-3005

−40 to 85 ◦ C 0–100% RH 0–5 m/s −100 to 400 ◦ C 0–1250 ◦ C −0.1 to 1.2 MPa −0.1 to 3.4 MPa 0–3.2 L/min 0.01–9.999 kW

±1.6% F.S. ±20 W

ASHRAE 116 [22] ISO 5151 [23]

measured by Pt100 RTDs. The power consumption of the prototype was measured by a power quality tester. All the measuring instruments were connected to two data loggers (Agilent 34970A), which allow a high sampling rate of the measurements.

2.3. Experimental procedures Before the commencement of the actual experiments, a preliminary test was carried out to check the proper functioning of the TEV and to determine an optimal superheat setting. This is important because the setting can significantly affect the performance of a TEV system. A high evaporator superheat can lead to underutilization of the evaporator surface to degrade the system performance. While if the superheat is set too low, the TEV may lose control on the refrigerant flow and cause “hunting” (refrigerant flow shows sustained oscillations). In this preliminary test, the superheat setting of the TEV was gradually increased to a position when the sustained oscillations started to disappear. After the preliminary test, two identical sets of experiments with the prototype operated with CT and TEV were performed. There were three test conditions for each set of experiments. The outdoor temperature was varied from 25 ◦ C to 35 ◦ C with an interval of 5 ◦ C (same as an earlier study [5]), while the indoor temperature was set at a low temperature of 18.5 ◦ C to enable the continuous operation of the compressor, and to facilitate evaluation of proper operation of the expansion device under low evaporating temperature. The desired conditions in the chambers were achieved by the use of PID controllers to provide real-time adjustments of the LGUs outputs and were maintained unchanged throughout each test. For each test, the outdoor chamber was first conditioned to the pre-set condition by the use of the BIACS and LGUs. The water in the storage tank was heated to an initial temperature of 25 ◦ C. Upon achieving the desired conditions, the prototype, the LGUs in the indoor chamber and all the measuring instruments were powered on to maintain the indoor chamber at 18.5 ◦ C and 65%RH. Under this condition, the indoor cooling load was actively adjusted to exactly offset the cooling output of the prototype. The test was continued until the mean water temperature in the storage tank reached 50 ◦ C. Due to the unsteady system behavior as the water temperature increases, all the measurements in this study were taken at a short interval of 5 s to better capture the dynamic variations.

temperature of 30 ◦ C. The results obtained for other test conditions will also be analyzed and will be given where necessary. 3.1. Refrigerant side Proper operation on the refrigerant side can be revealed by checking if under pressure fluctuations; the expansion device can still feed a proper amount of refrigerant to the evaporator, can keep a suitable degree of evaporator superheat, and can minimize the amount of vapor refrigerant formed in the expansion process to maximize the system capacity. To illustrate the pressure fluctuations during the heat recovery process, Fig. 2 compares the variations in compressor suction and discharge pressures for the two systems. It can be seen that the discharge pressures for both systems display a continuous increase with operating time. This is due to the fact that as the water temperature in the storage tank increases, the heat rejection performance of the prototype drops to increase the condensing pressure and thus discharge pressure. Between the two systems, the TEV system has a slightly higher discharge pressure (0.7% on average) due to the gradual reduction in opening extent of the TEV. While for the suction pressure, as explained in a previous study [24], it is less influenced by the heat recovery, thus a very small change is observed for both systems. Fig. 3 shows the changes in refrigerant flow rate with operating time, illustrating that the flow rate for the CT system increases with the discharge pressure. This is reasonable because the resistance across is a CT is factory fixed so that the refrigerant flow rate at any one time depends on the pressure difference across it [25]. On the contrary, the refrigerant flow rate for the TEV system is fairly stable despite the discharge pressure increase. This is due to the active control provided by the TEV. At a particular valve opening position,

3. Results and discussion This section discusses the performance of the prototype operated with TEV as compared to CT under various outdoor temperatures. The performance is evaluated based on the operation characteristics on the refrigerant side, the space cooling capacity, the water heating capacity, and the system energy efficiency. Due to the large amount of experimental data collected, performance comparisons of the two systems presented below are for an outdoor

Fig. 2. Compressor suction and discharge pressures comparison.

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Fig. 3. Refrigerant flow rates comparison.

the pressure increase also tends to increase the refrigerant flow, but that subsequently lowers the degree of superheat at the evaporator outlet. To restore the pre-set level of superheat, the TEV will reduce its opening extent and throttle the refrigerant flow. The flow rate thus does not vary much during the heat recovery process. Fig. 4 compares the degrees of evaporator superheat for the two systems, which are taken as the temperature difference between the saturated refrigerant and the refrigerant at the evaporator outlet. For the TEV system, the degree of superheat can be maintained constant as the discharge pressure increases due to the active control provided by the TEV, but the same cannot be done by the CT. For the CT system, the degree of superheat is high initially and decreases significantly with operating time. The high initial superheat value can lower the evaporator effectiveness and thus system performance and efficiency, and the decrease in superheat value with operating time is caused by the rise in refrigerant flow (Fig. 3) which will increase the risk of compressor flooding. The results in Fig. 4 were obtained under a lower than comfort indoor temperature (=18.5 ◦ C) and high water temperature (=50 ◦ C) which is considered the worst scenario where the evaporating temperature was low [26,27], and the compression ratio as well as the refrigerant flow rate were high to increase the risk of compressor flooding. A positive degree of superheat for both systems (8.1–3.4 ◦ C for the CT system; 5.9 ◦ C for the TEV system) therefore

Fig. 5. Comparison of xe,ir .

confirms the feasible use of CT and TEV in SEHRAC under higher indoor temperature conditions. The presence of vapor refrigerant at the evaporator inlet will reduce the refrigeration effect and is considered a loss to the system capacity [25]. This vapor is called flash gas, which is formed in the expansion process due to a drop in saturated temperature with reduced refrigerant pressure. The amount of flash gas at the evaporator inlet is represented by its mass fraction (vapor content, xe,ir ). xe,ir cannot be directly measured. It is estimated by the following equation: xe,ir =

he,ir − hsl hfg

(1)

where he,ir is the enthalpy of refrigerant at the evaporator inlet (kJ/kg); hsl is the saturated enthalpy of liquid refrigerant (kJ/kg); hfg is the refrigerant latent heat of evaporation (kJ/kg). The enthalpies in Eq. (1) can be evaluated by the use of a set of empirical equations [28]. By assuming an isenthalpic expansion process, he,ir is equal to the refrigerant enthalpy before the expansion device, which can subsequently be determined based on the corresponding pressure and temperature. Fig. 5 compares xe,ir for the two systems, illustrating xe,ir increases with operating time for both systems. This can be explained by the fact that as the water temperature increases, the amount of heat rejection drops to lower the subcooling effect. As a result, more flash gas will be generated to cool down the liquid refrigerant to the required evaporating temperature. Between the two systems, xe,ir for the TEV system is constantly smaller, especially after the initial stage. This is attributed to the higher heat rejection amount (Fig. 7) for the TEV system to increase the subcooling effect, as compared to the CT system [25]. In the above, the pressure fluctuations, refrigerant flow rates, degrees of evaporator superheat and refrigerant vapor contents at the evaporator inlet for the CT and TEV systems are explained and compared. 3.2. Space cooling capacity In this study, the space cooling capacity (Qcl ) of the prototype is determined based on the enthalpy drop across the evaporator (Eq. (2)), while the air enthalpy (ha ) is quantified by using Eqs. (3)–(5).

Fig. 4. Degrees of superheat comparison.

Qcl = me,a (hr,a − hs,a )

(2)

ha = 1.006ta + w(2501 + 1.86ta )

(3)

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Fig. 6. Space cooling capacities comparison. Fig. 7. Water heating capacities comparison.

ϕPsw w = 0.622 Pb − ϕPsw ln Psw =

C1 /Ta + C2 + C3 Ta + C4 Ta2

(4) + C5 Ta3

+ C6 ln Ta

(5)

where me,a is the evaporator supply air flow rate (kg/s); hr,a and hs,a are the return and supply air enthalpies (kJ/kg), respectively, of the evaporator; ta is the air temperature (◦ C); w is the moisture content (kg/kg dry air); ϕ is the relative humidity; Psw is the saturated pressure of water vapor (Pa); Pb is the atmospheric pressure (Pa); C1 to C6 are regression coefficients recommended by ASHRAE Handbook [29]; Ta is the absolute air temperature (K). Fig. 6 compares the space cooling capacities for the CT and TEV systems at an operating time interval of 10 min. It is noted that despite the TEV can operate to maintain a stable refrigerant flow and evaporator superheat in response to the refrigerant pressure fluctuations, the cooling capacity experiences a drop because of the increase in xe,ir , resulting in less amount of liquid refrigerant available for evaporation. This indicates the inadequacy of TEV when used in SEHRAC but is reasonable because TEV is typically a linear controller operating simply in response to a change in evaporator superheat [30]. It is also reasonable to see that the cooling capacity for the TEV system is on average 19.0% larger than the CT system. This is due to the better control of evaporator superheat (as explained in the previous section) and the interactive influence of refrigerant flow rate and xe,ir on the cooling capacity. In the initial stage, the larger cooling capacity for the TEV system can be explained by the higher refrigerant flow and the slightly lower xe,ir . After that, despite the refrigerant flow rate for the CT system increases rapidly to exceed the TEV system, the rise in xe,ir outweighs to give a smaller cooling capacity.

the heating capacities for both systems decrease continuously with operating time. This can be explained by the water temperature increase in the storage tank to decrease the temperature difference between the refrigerant and water. The heat transfer performance of the heat exchanger is thus adversely affected. As a result, the amount of vapor refrigerant condensed in the heat exchanger is reduced to lower the amount of heat rejection in the storage tank. For the same reason, it is noted that the heating capacities for both systems drop at a much lower rate in the first 40 min due to a higher temperature difference initially. Furthermore, it is reasonable to note that the heating capacity for the TEV system is on average 18.7% larger than the CT system because of the higher amount of recoverable heat resulted from larger space cooling capacity. It can also be explained by the higher compressor discharge pressure for the TEV system (Fig. 2) and thus higher refrigerant temperature at the storage tank inlet to enhance the heat transfer performance of the heat exchanger [27]. From the above analysis, it can be concluded that the use of TEV can result in a larger space cooling capacity and water heating capacity due to its active control characteristics. 3.4. Energy performance Fig. 8 compares the power consumptions for the CT and TEV systems. Due to the rise in compressor discharge pressure as shown

3.3. Water heating capacity In this study, the water heating capacity (Qwh ) of the prototype is deduced from the increase rate of mean water temperature in the storage tank: Qwh = cp,w w Vw

dtw d

(6)

where cp,w is the specific heat capacity of water (kJ/kg K); w is the water density (kg/m3 ); Vw is the total volume of water in the storage tank (m3 ); tw is the mean water temperature (◦ C);  is the operating time (s). Fig. 7 compares the water heating capacities for the CT and TEV systems at an operating time interval of 10 min. It can be seen that

Fig. 8. Power consumptions and overall COPs comparison.

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J. Jia, W.L. Lee / Energy and Buildings 101 (2015) 76–83 Table 3 Uncertainties in the calculated system performance. Performance parameter

Relative uncertainty (%)

Space cooling capacity Water heating capacity Power consumption Overall COP

CT system

TEV system

6.8 8.8 2.2 5.8

5.7 5.2 2.1 4.6

the result (ıR) can be estimated by using Kline and McClintock’s method (Eq. (9)) [31,32]: R = f (v1 , v2 , · · ·, vn )

(8)

  n  2  ∂R ıR =  ıvi Fig. 9. Energy performance comparisons for various outdoor temperatures.

in Fig. 2, the power consumptions for both systems experience a continuous increase with operating time. It is also noted that owing to the higher discharge pressure for the TEV system (Fig. 2), it has slightly larger power consumption than the CT system. In this study, the system energy efficiency is quantified by the use of overall coefficient of performance (COP), which takes into account the space cooling and water heating capacities as well as the power consumption of the prototype. The overall COP (COPoa ) is mathematically presented as: COPoa

Q + Qwh = cl W

(7)

where W is the power consumption of the prototype. The overall COP comparisons are also shown in Fig. 8, illustrating that the system efficiency decreases gradually with operating time. According to the previous discussions, this can be ascribed to the corresponding decrease in cooling and heating capacities and the increase in power consumption. Moreover, it is unveiled from Fig. 8 that the TEV system, despite the larger power consumption, still has an overall COP on average 18.6% higher than the CT system. To review if the above findings are also valid for other test conditions, the time-weighted average cooling and heating capacities, power consumptions and overall COPs for both systems are determined and compared for different outdoor temperatures in Fig. 9. It is noted that a rise in outdoor temperature increases the water heating capacity and simultaneously decreases the space cooling capacity to result in a cancelation effect in system output (sum of cooling and heating capacities). Therefore, the system output shows only a slight decrease with an increase in outdoor temperature from 25 ◦ C to 35 ◦ C, which combined with the increase in power consumption, leads to a drop in overall COP for both systems. It is also noted that similar to the results for an outdoor temperature of 30 ◦ C, the overall COP for the TEV system is 12.5–20.9% higher than the CT system, primarily owing to the larger system output.

3.5. Uncertainty assessment The uncertainties in the calculated system performance can be determined based on the instrument accuracies summarized in Table 2. For a result R needs to be determined from a number of independent variables (vi ) (Eq. (8)), if the uncertainties in the measured values of these variables can be ascertained, the uncertainty in

(9)

∂vi

i=1

Therefore, Eqs. (10)–(12), derived on the basis of Eqs. (2), (6) and (7), can be used to determine the uncertainties in the calculated space cooling and water heating capacities as well as overall COP, while that in the power consumption can simply be taken as the corresponding instrument uncertainty. ıQcl = Qcl

 

ıQwh = Qwh ıCOPoa = COPoa

ıme,a me,a

2 +

ıh2r,a + ıh2s,a (hr,a − hs,a )

(10)

2

2 2 ıtw,+ + ıtw,−

(11)

(tw,+ − tw,− )2



ıQcl Qcl + Qwh

2

+

ıQwh Qcl + Qwh

2

+

ıW W

2 (12)

Because of the unsteady system behavior as shown in the previous sections, the maximum possible uncertainties are calculated and summarized in Table 3. The results are ranging from 2.1% to 8.8%, which are judged to be acceptable as they are less than the ISO’s recommendation of 10% [23]. Performance comparison results for the two systems are summarized in Table 4. It is noted that a consistent result is obtained for the percentage increase in space cooling capacity (19.0%), water heating capacity (18.7%) and overall COP (18.6%) for the TEV system as compared to the CT system to confirm the results obtained are reasonable. In addition, an acceptable level of uncertainty is obtained, confirming that the TEV system has an overall performance better than the CT system. Table 4 Summary of performance comparisons of the two systems in response to the refrigerant pressure fluctuations. Performance

CT system

TEV system

Proper regulation of refrigerant flow Maintaining a stable evaporator superheat

No (increasing)

Yes (stable)

No (decreases from 8.1 ◦ C to 3.4 ◦ C) Significant

Yes (maintained at 5.9 ◦ C) Moderate

Decreases from 2.0 kW to 1.5 kW Decreases from 2.1 kW to 0.6 kW Decreases from 4.5 to 1.9

Average 19.0% highera Average 18.7% highera Average 18.6% highera

Rise in refrigerant vapor content at the evaporator inlet Space cooling capacity Water heating capacity Overall COP a

As compared to the CT system.

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4. Conclusions The use of CT and TEV in SEHRAC as the expansion device was side-by-side compared by experimental investigations. To facilitate the investigations, a prototype, which could be switched between the CT and TEV systems, was designed and setup in a test facility for two identical sets of experiments. For each set of experiments, a series of tests under different outdoor temperatures were conducted. It was found that the TEV system could better regulate the refrigerant flow, maintain a stable evaporator superheat, and minimize the rise in refrigerant vapor content at the evaporator inlet. The space cooling and water heating capacities for the TEV system were on average 16.3–19.4% and 18.5–23.4% larger than the CT system. Its overall COP was also found 12.5–20.9% higher for the range of outdoor temperatures. However, for both systems, due to a drop in heat transfer performance of the heat exchanger with operating time, a corresponding drop in cooling and heating capacities was observed. Uncertainty assessment was conducted. It was found that the measured and calculated results in this study were within an acceptable range of accuracy. On this basis, the study confirmed that TEV is preferred for use in SEHRAC as the expansion device. This conclusion, together with the experimental protocol established, will also be useful for further development of SEHRAC for energy saving. Acknowledgement This work is supported by Hong Kong Research Grants Council General Research Fund No. 519412. References [1] Electrical and Mechanical Services Department, Hong Kong Energy End-Use Data, 2013, Hong Kong SAR Government, 2013. [2] W.M. Ying, Performance of room air conditioner used for cooling and hot water heating, ASHRAE Trans. 95 (1989) 441–444. [3] N.J. Monerasinghe, R. Ratnalingam, B.S. Lee, Conserved energy from room air-conditioners for water heating, Energy Convers. Manage. 22 (1982) 171–173. [4] J. Jia, W.L. Lee, Applying storage-enhanced heat recovery room air-conditioner (SEHRAC) for domestic water heating in residential buildings in Hong Kong, Energy Build. 78 (2014) 132–142. [5] J. Jia, W.L. Lee, Experimental study of the application of intermittently operated SEHRAC (storage-enhanced heat recovery room air-conditioner) in residential buildings in Hong Kong, Energy (2015), http://dx.doi.org/10.1016/j.energy. 2015.02.075 (in press). [6] P. Techarungpaisan, S. Theerakulpisut, S. Priprem, Modeling of a split type air conditioner with integrated water heater, Energy Convers. Manage. 48 (2007) 1222–1237. [7] M.L. Jiang, J.Y. Wu, Y.X. Xu, R.Z. Wang, Transient characteristics and performance analysis of a vapor compression air conditioning system with condensing heat recovery, Energy Build. 42 (2010) 2251–2257.

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[8] S.W. Wang, Z.Y. Liu, Y. Li, K.K. Zhao, Z.G. Wang, Experimental study on split air conditioner with new hybrid equipment of energy storage and water heater all year round, Energy Convers. Manage. 45 (2005) 3047–3059. [9] J.M. Choi, Y.C. Kim, The effects of improper refrigerant charge on the performance of a heat pump with an electronic expansion valve and capillary tube, Energy 27 (2002) 391–404. [10] B.H. Kim, D.L. O’Neal, Effect of refrigerant flow control on the heating performance of a variable-speed heat pump operating at low outdoor temperature, J. Solar Energy Eng. 127 (2005) 277–286. [11] J. Zhang, R.Z. Wang, J.Y. Wu, System optimization and experimental research on air source heat pump water heater, Appl. Therm. Eng. 27 (2007) 1029–1035. [12] N. Agrawal, S. Bhattacharyya, Optimized transcritical CO2 heat pumps: performance comparison of capillary tubes against expansion valves, Int. J. Refrig. 31 (2008) 388–395. [13] W.F. Stoecker, L.D. Smith, B.N. Emde, Influence of the expansion device on the seasonal energy requirements of a residential air conditioner, ASHRAE Trans. 87 (1981) 349–360. [14] M. Farzad, D.L. O’Neal, Influence of the expansion device on air conditioner system performance characteristics under a range of charging conditions, ASHRAE Trans. 99 (1993) 3–13. [15] G.G. Maidment, J.F. Missenden, R.W. James, R.M. Tozer, Analysis of the expansion valves used for controlling refrigerant feed into delicatessen cabinets in supermarkets, J. Food Eng. 40 (1999) 47–58. [16] H. Chen, W.L. Lee, Combined space cooling and water heating system for Hong Kong residences, Energy Build. 42 (2010) 243–250. [17] ASHRAE Standard 16, Method of testing for rating room air conditioners and packaged terminal air conditioners, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, 1983. [18] British Standard 5141, Specification for air heating and cooling coils, 1975. [19] ASHRAE Standard 41.2, Standard method for laboratory airflow measurement, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, 1987. [20] ASHRAE Standard 41.1, Standard method for temperature measurement, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, 1986. [21] ASHRAE Standard 41.3, Standard method for pressure measurement, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, 1989. [22] ASHRAE Standard 116, Methods of testing for rating seasonal efficiency of unitary air conditioners and heat pumps, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, 2010. [23] ISO Standard 5151, Non-ducted air conditioners and heat pumps – testing and rating for performance, 2010. [24] T.Y. Bong, M.N.A. Hawlader, W. Mahmood, Influence of expansion device on the performance of an air-conditioner with a desuperheater, ASHRAE Trans. 94 (1988) 661–672. [25] R.J. Dossat, Principles of Refrigeration, fourth ed., Prentice-Hall, 1997. [26] A. Moreno-Rodríguez, A. González-Gil, M. Izquierdo, N. Garcia-Hernando, Theoretical model and experimental validation of a direct-expansion solar assisted heat pump for domestic hot water applications, Energy 45 (2012) 704–715. [27] A. Moreno-Rodriguez, N. Garcia-Hernando, A. González-Gil, M. Izquierdo, Experimental validation of a theoretical model for a direct-expansion solarassisted heat pump applied to heating, Energy 60 (2013) 242–253. [28] A.C. Cleland, Computer subroutines for rapid evaluation of refrigerant thermodynamic properties, Int. J. Refrig. 9 (1986) 346–351. [29] ASHRAE Handbook, Fundamentals, American Society of Heating Refrigerating and Air-Conditioning Engineers, Atlanta, 2009. [30] P.M.T. Broersen, M.F.G. Van der Jagt, Hunting of evaporators controlled by a thermostatic expansion valve, J. Dyn. Syst. Meas. Control 102 (1980) 130–135. [31] E.O. Doebelin, Measurement Systems Application and Design, third ed., McGraw-Hill, 1983. [32] J.P. Holman, Experimental Methods for Engineers, fifth ed., McGraw-Hill, 1989.