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Experimental analysis on performance of high temperature heat pump and desiccant wheel system
Energy and Buildings 66 (2013) 505–513
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Experimental analysis on performance of high temperature heat pump and desiccant wheel system Ying Sheng a,∗ , Yufeng Zhang a , Na Deng a , Lei Fang b , Jinzhe Nie b , Lijun Ma a a b
School of Environmental Science and Engineering, Tianjin University, Tianjin 300072, China International Centre for Indoor Environment and Energy, Technical University of Denmark, Lyngby DK-2800, Denmark
a r t i c l e
i n f o
Article history: Received 10 December 2012 Received in revised form 2 May 2013 Accepted 26 July 2013 Keywords: Desiccant wheel High temperature heat pump Coefficient of performance
a b s t r a c t In order to solve the problem of high energy consumption for regeneration of desiccant wheel in the rotary desiccant system, high temperature heat pump and desiccant wheel (HTHP&DW) system and corresponding air conditioning unit is built and tested in the extensive thermal hygrometric environment. When the mixture refrigerant BY-3 is involved in the air source heat pump, the supply air temperatures are in the range as expected except that when in the extreme hot environment (above 36 ◦ C), dehumidification capability are satisfied and the regeneration temperatures can satisfy the regeneration requirement of desiccant without additional heat. It is also found that outdoor air temperature, humidity ratio and regeneration air flow rate have great impact on the performance of heat pump based on the coefficient of performance (COP) evaluated. COP is not quite high, as the maximum value is 2.26 for heat pump and 2.08 for whole system respectively. Hence several suggestions are made for optimizing the system which is also helpful for facilitating the development of mature product. As a conclusion, HTHP&DW system could be a potential alternative with effective operation which can be promoted in most areas of China based on the result of our experiment. © 2013 Elsevier B.V. All rights reserved.
1. Introduction Sensible load arising from heat transfer into the conditioned space and latent load arising from moisture generated within the space are two main types of the loads that have to be handled in the air conditioning system. In a conventional vapor compression cooling system, air is usually cooled below its dew point and subsequently heated up to the desired supply temperature, which often costs a large amount of energy for overcooling and reheating [1,2]. The desiccant cooling systems are being developed as an alternative to overcome the flaws of vapor compression system [3]. Especially, rotary desiccant air conditioning systems, which are compact and less subject to corrosion and can work continuously, attract more attention [4]. Many configurations of rotary desiccant system are designed and studied and the encouraging results are obtained [5–7]. It is also found that efficient energy saving and low environment impact are achieved when the desiccant wheel is regenerated by means of low temperature thermal energy such as solar heat [8–10], energy from co-generators [11,12], and waste heat [13]. Although the costs of these technologies are near-zero
∗ Corresponding author. Tel.: +86 186 2200 7007. E-mail addresses: [email protected] (Y. Sheng), [email protected] (Y. Zhang), [email protected] (N. Deng), fl@byg.dtu.dk (L. Fang), [email protected] (J. Nie), [email protected] (L. Ma). 0378-7788/$ – see front matter © 2013 Elsevier B.V. All rights reserved. http://dx.doi.org/10.1016/j.enbuild.2013.07.058
[14], the applications of them are limited due to the climate or regional factors. One type of rotary desiccant system is a combination of the rotating desiccant wheel and of the vapor compression system, which can be widely and efficiently used without limitations. The shortcoming is that the thermal heat recovered from the condenser is not sufficient for the desiccant regeneration and a supplementary heater is needed [4], which will increase the energy input and equipment investment. The proposed method in this paper will solve this problem by utilizing the characteristics of high temperature heat pump. Heat pump has a heating capacity at the condenser and a cooling capacity at the evaporator, which is an approach to satisfy simultaneous energy demands for heating and cooling. Many researchers have worked on such systems for simultaneous production in various applications [15–17]. Take advantage of simultaneous production of heating and cooling, heat pump desiccant system as a novel topic has been studied. Lazzarin [18] has investigated a new equipment of self-regenerating liquid desiccant cooling system for supermarket application. Enteria [19] has performed an experimental evaluation of a new solid desiccant heat pump system, the main feature of which is direct impregnation of the desiccant material in the evaporator and condenser of the heat pump system. In Ref. [20], the variable refrigerant volume (VRV) system has been operated in conjunction with a self-regenerating heat pump desiccant unit. And its energy consumption has been analyzed as opposed to the VRV system coupling with heat recovery ventilation
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Nomenclature cp COP HTHP DW V m t h Q E DBT WBT TR PR BY-3 FAR AFR
specific heat of the air (kJ/kg K) coefficient of performance high temperature heat pump desiccant wheel air volumetric flow rate (m3 /h) air quality flow rate (kg/h) air temperature (◦ C) enthalpy of the air (kJ/kg) heat energy (kW) electrical input power (kW) dry bulb temperature (◦ C) wet bulb temperature (◦ C) temperature of refrigerant (◦ C) pressure of refrigerant (MPa) refrigerant of No.3 of BeiYang fresh air ratio (%) air velocity (m/s)
Greek symbols ω air humidity ratio (g/kg) air density (kg/m3 ) Subscripts com compressor sys system proc process air reg regeneration air f fan c, h cooling and heating, respectively W, N states of outdoor air and indoor air, respectively inlet and outlet state of the process air, respectively C, E O state of supply air A, B inlet and outlet state of the regeneration air, respectively
unit. By utilizing the characteristics of high temperature, Hao and Zhang [21] have explored the potential of integrating the moderate or high temperature heat pump subsystem to the desiccant wheel. However the performance of this kind of system is less complex investigated over a wide range of operating conditions. Therefore, a detailed experiment is carried out in this paper with the objectives as follows: • A new high temperature heat pump coupling with solid desiccant wheel (HTHP&DW) system is designed, and the corresponding air conditioning unit is built. • The mixture refrigerant of “BY-3” is involved and tested in order to ascertain that BY-3 can be a candidate for this system. • An experiment is conducted to study the applicability for extensive thermal hygrometric conditions representing most climate zones in China, and the performance of HTHP&DW system. 2. HTHP&DW air conditioning unit design 2.1. HTHP&DW system Fig. 1 shows the schematic diagram of HTHP&DW system. The ventilation system supplies minimum levels of outdoor airflow and re-circulates a large amount of indoor air on the process side. The return air (state N) and outdoor fresh air (state W), with specific ratio of FAR = 20%, are mixed to become the process air (state C) which flows through the desiccant wheel where a large amount of
moisture is removed by the desiccant (state E). The process air is then cooled by the evaporator of HTHP that is controlled for conditioning the room air temperature to a comfortable level (state O). The outdoor fresh air (state W) is used for regenerating the rotating desiccant wheel. The regeneration air is heated up by the condenser of the heat pump (state A) and is expelled after regenerating the desiccant wheel (state B). 2.2. Air conditioning unit design and manufacture The high temperature heat pump, desiccant wheel and fans are the main components of the air conditioning unit. The main specifications of the machines are listed as follows: Desiccant wheel: the desiccant material is silica-gel and the configuration is honeycomb. The wheel is characterized by the following layout: process air passes 75% of the flow area and regeneration air the remaining 25%. The other characteristics of the wheel are: diameter and thickness equal to 450 mm and 200 mm respectively, the nominal rotational speed is 15 r/h. HTHP: The invariable frequency closed piston compressor branded with Copeland made in America is used with the nominal input power of 2 kW. The type of heat exchanger is air source chip tubular. The manual expansion valve with brand of Swagelok is used to adjust the flow rate of the refrigerant. The volume flow rate is in the range of 0–0.084 m3 /h. Fan: the frequency conversion fans are chosen in order to obtain the desired speed. The nominal power of both process and regeneration fans are 0.3 kW and the frequency is 0–50 Hz. Mature commercial products were adopted for assembling this air-conditioning unit. Fig. 2 shows the layout of air conditioning unit and the experiment is carried out on this facility. The following contents should be noted with attention during design and commissioning: (a) Insulation is necessary due to the temperature difference between different air flows. (b) In order to prevent exchange between the process air and regeneration air, elastic material with high temperature resistance should be well sealed on the surface edge of desiccant wheel. Simultaneously harmful gas should not be produced from the elastic material during the operation of unit. (c) The wind resistances of main machines should be taken into consideration during the selection of fans that can ensure the external residual pressure. Besides noise, size and performance of fans should also be considered. (d) The main machines are fixed on a steel frame with scroll wheels on the bottom to meet the requirement of mobile operation. It is expected that our experience with the above unit can facilitate development of commercial product. Compared to the conventional cooling-based dehumidification air conditioning system, this air conditioning unit has the following advantages: • The sensible and latent heat load are handled independently. • This method is used for connecting the desiccant wheel and HTHP to make full use of both heating and cooling from the condenser and evaporator of the heat pump that can provide the thermal energy for desiccant regeneration without additional heat. • The desiccant wheel can be driven only by the heat pump without any heating machine, which will save the cost of equipment investment. • The structure of this air conditioning unit is compact, with the feature of less occupied area.
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Fig. 1. Schematic of HTHP&DW system.
• The system is working in a completely dry environment that will effectively prevent the growth of bacteria. Hence a better indoor air quality can be obtained. 3. Choice of refrigerant For the proposed HTHP, it is assumed that the whole cooling load and regeneration energy would not only provide sufficient energy for heating and cooling, but also meet the demand of high-temperature heat for regeneration. Therefore the choice of refrigerant is critical for the system operation. In this work, the
“BY-3”, a mixed non-azeotropic refrigerant, developed by Tianjin University was chosen. The results of theoretical calculation of this working fluid are listed in Table 1. The REFPROP 8.0-NIST Reference Fluid Properties Database [22] is employed in theoretical calculation. The experiment analysis of BY-3 conducted by Chen and Zhang [23] based on a single-stage vapor compression refrigeration system showed that the highest temperature at outlet of the condenser reached about 85 ◦ C, when the difference between the water temperatures at the condenser outlet and the evaporator inlet was less than 40 ◦ C, COP was larger than 4; and when the difference of that
Fig. 2. Layout of test facility.
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Table 1 Properties of BY-3. Critical temperature (◦ C)
Critical pressure (MPa)
Ozone depletion potential
Global warmth potential
Security level
126.2
4.664
0.04
2 100
Aa 1b
a b
According to TLV-TWA [30], when the refrigerant concentration is equal to or less than 400 × 10−4 %, it is non-toxic. There is no flame spreading in the air if the refrigerant is tested under the condition of 101.3 kPa and 21 ◦ C.
reached 55 ◦ C, the COP still remained 3. The discharge temperature of BY-3 was lower than 100 ◦ C, and the refrigerant vapor pressure kept lower than 1.8 MPa. Based on many experiment results, the rotary desiccant cooling system is more feasible when the required regeneration temperature is low (60–70 ◦ C is sufficient) [24,25]. According to the results of [26], the dehumidification capacity can be 2.5 g/kg when the regeneration temperature is 60 ◦ C, which could satisfy the dehumidification requirement of air conditioning systems. Hence preferably the regeneration temperature is not less than 60 ◦ C. Obviously the air temperature can be raised to over 60 ◦ C if BY3 is involved in the air-source heat pump based on the results of [23]. BY-3 can be a candidate for HTHP&DW system except the pure refrigerant [21]. The detailed discussions on the performance of heat pump using BY-3 are in Section 5.
4. Experimental procedure 4.1. Experimental conditions As shown in Table 2, the experiment is conducted in an extensive thermal hygrometric range that represents most zones in China. The latent loads from indoor air and fresh air are transformed to sensible load during the absorption process, which will increase the inlet air temperature of evaporator of heat pump. More attention should be paid on the question whether the heat pump can assume the large sensible load. Therefore the desired supply air temperature is designed to be 20 ◦ C to verify the cooling capacity of heat pump. Meanwhile the dehumidification capability of this air conditioning unit is also examined and discussed. The volumetric process air flow rate are 270 m3 /h and 540 m3 /h and that of regeneration air flow are 180–350 m3 /h, respectively. Fig. 3 gives an example of air processes on the psychrometric chart at the condition of TW = 34 ◦ C, ωW = 20 g/kg, FAR = 20%. This is also the typical experiment condition designed to meet the extreme weather during the summer in Tianjin of China. The detailed discussion of air properties is in Section 5.1.
4.2. Measurement devices The state of the process air is that of the mixture of outside fresh air and return air with certain ratio. In order to simplify the manufacture of the test facility, the mixed state is achieved by pre-heat and pre-humidification in the treatment tank. The regeneration air is controlled with the adjustable heater and humidifier similarly. The parameters measured in this work include the inlet and outlet air DBT and WDT for both absorption and regeneration, air flow rate, temperature and pressure of refrigerant at inlet and outlet of compressor, input power of compressor and fans. The specifications of the measuring devices are listed in Fig. 2.
4.3. Performance parameter The performance parameters are divided into two categories. The first group deals with the effectiveness of the system in meeting the required indoor conditions. The second category pertains to
efficiency-related indexes, such as COP based on electrical energy input. COP for cooling of heat pump is represented as Qc Ecom
COPcom,c =
(1)
where Qc is the cooling capacity and is defined as Qc = mproc (hE − hO ) = E Vproc (tE − tO )cp
(2)
COP for cooling of the entire air conditioning system is represented as COPsys,c =
Qc E
(3)
In Eq. (3), E is the electrical power used to drive the vapor compressor and fans. Fans are employed to force circulate the regeneration air and process air streams. Hence, E is calculated as E = Ecom + Ef,proc + Ef,reg
(4)
In our experiment study, it is found that the electricity used for fans represents 11% of total electricity consumption for average. Therefore when the performance of this system is evaluated, the electricity input for fans should not be neglected. Heating COP of heat pump is defined as COPcom,h =
Qh Ecom
(5)
In Eq. (5), Qh is the heating capacity, which is calculated as Qh = mreg (hA − hW ) = A Vreg (tA − tW )cp
(6)
The error analysis of the experimental results is performed according to the propagation of the uncertainties. An overall accuracy within the maximum amount of relative uncertainty is 3.73% for COPsys,c , 3.6% for COPcom,c , 4.1% for COPcom,h , 2.01% for Qc and 2.83% for Qh respectively. 5. Results The heat pump has to provide sufficient heating for regeneration and cooling for ventilation air to the conditioned space. In summer, large amount of moisture should be removed from the conditioned space which means that higher regeneration temperature is required. Whether the desired regeneration temperature could be attained becomes the sticking point for whether HTHP&DW system could be a new approach for air conditioning system. Hence the performance of heat pump is the focus in this section. 5.1. Air processes The measurement and calculation report of the air properties for the typical experimental condition of TW = 34 ◦ C, ωW = 20 g/kg, FAR = 20% are shown in Table 3. It is found that air temperature of process air is 20.56 ◦ C which meet the desired supply air temperature (20 ± 1 ◦ C) after the air steam flows through the desiccant wheel. Also the calculated sensible and latent loads for the given processes are listed in Table 4. It is found that the sensible load assumed by DW (process 3) is bigger than that of total sensible load (process 1 + 2 + 4) due to the heat transfer between process air and regeneration air. Furthermore the latent load assumed by DW
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Table 2 Experimental conditions.
Indoor air Outdoor air
Air temperature (◦ C)
Air humidity ratio (g/kg)
Fresh air ratio (%)
Desired supply air temperature (◦ C)
26 28–40
11 8–22
20
20 ± 1
Fig. 3. Air processes on the psychrometric chart.
(process 4) is also bigger than that of total latent load (process 1 + 2). This is because of the mass transferred from the regeneration side to the process side and a minimum leakage of outside air into the air passage. The relative differences of sensible load and latent load are 4% and 5% respectively. In order to explore the dehumidification feasibility of this system used in an extensive climate range and evaluate the performance of the system, a detailed experiment is conducted and
Table 3 Typical measured and calculated state properties of the system shown in Fig. 3. State
DBT (◦ C)
WBT (◦ C)
ω (g/kg)
N W C E O A B
26.00 34.00 27.60 47.33 20.56 63.13 38.17
19.20 27.25 20.83 23.62 15.22 33.23 30.00
11.00 20.00 12.80 8.46 8.46 20.00 23.50
h (kJ/kg) 54.29 85.59 60.53 66.61 42.24 116.08 98.95
Table 4 The calculated sensible and latent load for given processes. No. Processes
1 W→N
2 O→N
3 E→O
4 C→E
Sensible load (kW) Latent load (kg/h) Total sensible load of (1 + 2 + 4) Total latent load of (1 + 2)
0.14 0.58 2.32 1.24
0.39 0.66
2.43 –
1.79 1.31
Note: Vproc = Vreg = 270 m3 /h.
the results are shown in Figs. 4 and 5. Fig. 4 gives the supply air temperature and humidity ratio obtained at different climate conditions, in which it is found that the supply air temperatures (TO ) are in the range as expected (20 ± 1 ◦ C) except when in the extreme hot environment (TW beyond 36 ◦ C). All the supply air humidity ratios (ωO ) are satisfied and the maximum dehumidification efficiency could achieve 38% at the condition of TW = 40 ◦ C, and ωW = 20 g/kg. However, it should be noted that ωO is lower when the outdoor air temperature is higher. This can be attributed to the fact that, when TW is higher, Treg is higher, and dehumidification ability of desiccant wheel increases with the regeneration temperature, resulting in a lower ωO . Fig. 5 shows the regeneration temperature obtained. It is found that most of regeneration temperature (Treg = TA ) are satisfied in the range of 60–66 ◦ C. Therefore this HTHP&DW air conditioning system could be a new approach to provide comfort indoor environment for people. 5.2. Influences of regeneration air flow on the heating capability of heat pump In this section, regardless of supply air temperature (TO ), impact of flow rate of regeneration air (Vreg ) on the behavior of heat pump are investigated under the constant working conditions of Vproc = 540 m3 /h, ωW = 14 g/kg. Fig. 6 shows the variation of regeneration temperature and heating COP against outdoor air temperature at the conditions of different regeneration air flow rate. With the reduction of Vreg , Treg increases that leads to the increase of heat transfer temperature difference between the inlet and outlet air temperature of condenser. Meanwhile the heat transfer coefficient on the air side of
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(b)
the condenser is reduced due to the decrease of the air velocity. As a result, the heating capacity increases slightly. A higher Treg means the condensing temperature should be higher that determines a higher condensing pressure of the compressor. Because of that, the compression ratio is high that increases the input power of compressor. When the air flow decreases, COPcom,h decreases with the rise of input power in the condition of relatively stable heating as confirmed by [27]. It is found that Treg are between 60 ◦ C and 70 ◦ C with a maximum of 74 ◦ C when the volume of air flow equals to 270 m3 /h. Although higher regeneration temperatures are obtained in the condition of Vreg = 180 m3 /h, COPcom,h becomes too low and the compressor will shut down if the regeneration air flow rate is further reduced below 180 m3 /h. Therefore the regeneration air flow rate has significant impact on the performance of heat pump. It is found that a lower regeneration air flow rate results in a lower COPcom,h . 5.3. Global behavior of the system in the extensive climate zone Based on the analysis above, Vproc and Vreg are selected as 270 m3 /h respectively. Apart from the factors analyzed in Section 5.2, outdoor air temperature and humidity ratio do influence the performance of this system.
Fig. 4. Supply air temperature and humidity ratio obtained in the extensive thermal hygrometric environment.
Fig. 5. Regeneration temperature obtained in the extensive thermal hygrometric environment: regeneration temperature equals to the outlet air temperature of condenser.
Fig. 6. Regeneration temperature and heating COP of heat pump against outdoor air temperature.
5.3.1. Effects of outdoor humidity ratio Fig. 7 shows the system performance variations with outdoor air humidity ratio at the constant outdoor air temperature (TW = 34 ◦ C). As ωW increases, cooling capacity of heat pump increases monotonically in the test range of 8–22 g/kg as represented in Fig. 7(a). This is because, with the rise of humidity ratio, the desiccant wheel removes a bigger amount of water vapor, which results in the rise of outlet air temperature (TE ) from the process zone of desiccant wheel and finally results in an increase of the total cooling capacity. In order to ensure the supply air temperature, the opening of expansion valve is increased, which leads to the rise of mass flow of refrigerant. Hence a higher performance is attained as confirmed by [28] when the rise of cooling capacity is more than that of input power of compressor. The input power for fans is constant due to the unchanged regeneration and process air flow rate and therefore COPsys,c shows the same trend as COPcom,c represented in Fig. 7(b). 5.3.2. Effects of outdoor air temperature Fig. 8 shows the variations of cooling capacity, input power of compressor and cooling COP with outdoor air temperature at the constant outdoor air humidity ratio (ωW = 18 g/kg). As represented in Fig. 8(a), with the rise of ambient temperature, cooling capacity increases. This can be attributed to the fact that a higher TW causes a higher Treg , and air could remove more moisture during the regeneration phase. Similarly, during the dehumidification process, the desiccant could adsorb more vapor from the process air, resulting in a rise of total cooling capacity. And the increase of cooling capacity is more than that of input power of compressor. Therefore COPcom,c is higher at a higher outdoor air temperature as confirmed by [28,29]. Fig. 8(b) shows the influence of outdoor air temperature on COP obtained from the experimental study. In order to maintain the normal operation and obtain the desired regeneration temperature, the volume of air flow through the condenser is adjusted higher, which results in the increase of electricity consumption for fans. Therefore COPsys,c does not increase monotonically as COPcom,c because of the rise of input power for fans when TW increases. Then COPsys,c reaches the peak when TW = 34 ◦ C.
Y. Sheng et al. / Energy and Buildings 66 (2013) 505–513
(a)
(b)
Fig. 7. Variation of cooling capacity, input power of heat pump and cooling COP with outdoor air humidity ratio.
511
(a)
(b)
Fig. 8. Variation of cooling capacity, input power of heat pump and cooling COP with outdoor air temperature.
6. Discussions It should be noted that the performances at other tested outdoor temperatures and humidity ratios show the same trends as presented in Figs. 7 and 8. Fig. 9 gives the comparisons of cooling COP between heat pump and whole system. It is found in Fig. 9(a) that COPcom,c calculated at air temperature of 34 ◦ C, 36 ◦ C,38 ◦ C are very close. The reason for this is that the regeneration temperatures are close as shown in Fig. 5 that determines the similar cooling capacity of heat pump for above given explanation for Fig. 8(a). Fig. 9(b) shows the cooling COP of whole system in the extensive thermal hygrometric environment. It is found that COPsys,c get a peak value when TW = 34 ◦ C and ωW = 20 g/kg due to the same reason explained for Fig. 8(b). So it could be concluded that the maximum of COPcom,c is 2.26 when TW = 40 ◦ C and ωW = 20 g/kg; and COPsys,c gets a maximum value of 2.08 when TW = 34 ◦ C and ωW = 20 g/kg. Based on the above analysis, a higher outdoor air temperature and a higher humidity ratio decide a better performance of heat pump. When the operation of heat pump is observed during the experiment, the condensing pressures are below 2.5 MPa in line with the pressure bearing capacity of general air-conditioning unit and the compression ratios are in the range of 5–7 by using the refrigerant of BY-3. There is no reaction between BY-3, refrigerant lubricant, and compressor materials, and this mixture is chemically stable in a wide range of operating temperatures.
Although the feasibility used in extensive climate zone of HTHP&DW air conditioning system is verified, some results are not satisfied such as low COPc . Therefore the author provides some suggestions for optimizing the performance of this system. It is found in Figs. 7(a) and 8(a) that the input power of compressor does not increase in equal proportion with the rise of cooling capacity. This is because the feature of compressor involved in the system is invariable frequency, this kind of compressor cannot change the speed of motor, and then the electricity consumption does not change with the variations of cooling capacity. Therefore the inverter compressor is preferred in order to save energy and improve the performance of heat pump. The precise supply air temperature, humidity ratio and regeneration temperature cannot be obtained without automatic control system. So it is suggested that the automatic control system should be installed in this air conditioning unit. In order to maintain the normal operation of heat pump, large volume of air should flow through the condenser and sequentially pass through the regeneration zone of desiccant wheel, which causes superfluous heat to regenerate the desiccant and therefore the heat transfer between two air streams is strengthened and then increases the burden of heat pump. For this reason the structure of air conditioning unit should be rebuilt so as to expel the unwanted heat from condenser. Furthermore the optimal regeneration air
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Technology of P.R. China in the Sino-Danish collaborated research project: “Activating the Building Construction for Building Environment Control” (Danish International DSF project no. 09-71598, Chinese international collaboration project no. 2010DFA62410).
References
Fig. 9. Comparisons of cooling COP between heat pump and whole system in the extensive thermal hygrometric environment.
velocity, regeneration temperature and proportion of process and regeneration zones should be further studied. 7. Conclusion In this work, HTHP&DW system and corresponding air conditioning unit is built and tested to study the feasibility used in extensive climate zone. Furthermore the performance of heat pump and entire system are analyzed. The property of BY-3 has a strong ability to work in the large temperature difference environment. When BY-3 is involved in the air source heat pump, the supply air temperatures are in the range as expected except when in the extreme hot environment (TW above 36 ◦ C), and the regeneration temperatures can reach more than 60 ◦ C and dehumidification capability are satisfied. HTHP&DW system can be an alternative for the desiccant cooling system that can be used in the extensive thermal hygrometric environment. The ambient thermal hygrometric environment and regeneration air flow rate have great impact on the performance of heat pump. Lower regeneration air flow rate leads to a lower COPcom,h , and higher outdoor air temperature and humidity ratio decide a higher COPcom,c . However, COPc is not quite high with the maximum of 2.26 for heat pump and 2.08 for whole system respectively. Several suggestions provided for optimizing the system are helpful for saving energy and facilitating the development of mature product. Acknowledgements This work is collectively sponsored by the Danish Agency for Science, Technology and Innovation and the Ministry of Science and
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Contents lists available at ScienceDirect
Energy and Buildings journal homepage: www.elsevier.com/locate/enbuild
Experimental analysis on performance of high temperature heat pump and desiccant wheel system Ying Sheng a,∗ , Yufeng Zhang a , Na Deng a , Lei Fang b , Jinzhe Nie b , Lijun Ma a a b
School of Environmental Science and Engineering, Tianjin University, Tianjin 300072, China International Centre for Indoor Environment and Energy, Technical University of Denmark, Lyngby DK-2800, Denmark
a r t i c l e
i n f o
Article history: Received 10 December 2012 Received in revised form 2 May 2013 Accepted 26 July 2013 Keywords: Desiccant wheel High temperature heat pump Coefficient of performance
a b s t r a c t In order to solve the problem of high energy consumption for regeneration of desiccant wheel in the rotary desiccant system, high temperature heat pump and desiccant wheel (HTHP&DW) system and corresponding air conditioning unit is built and tested in the extensive thermal hygrometric environment. When the mixture refrigerant BY-3 is involved in the air source heat pump, the supply air temperatures are in the range as expected except that when in the extreme hot environment (above 36 ◦ C), dehumidification capability are satisfied and the regeneration temperatures can satisfy the regeneration requirement of desiccant without additional heat. It is also found that outdoor air temperature, humidity ratio and regeneration air flow rate have great impact on the performance of heat pump based on the coefficient of performance (COP) evaluated. COP is not quite high, as the maximum value is 2.26 for heat pump and 2.08 for whole system respectively. Hence several suggestions are made for optimizing the system which is also helpful for facilitating the development of mature product. As a conclusion, HTHP&DW system could be a potential alternative with effective operation which can be promoted in most areas of China based on the result of our experiment. © 2013 Elsevier B.V. All rights reserved.
1. Introduction Sensible load arising from heat transfer into the conditioned space and latent load arising from moisture generated within the space are two main types of the loads that have to be handled in the air conditioning system. In a conventional vapor compression cooling system, air is usually cooled below its dew point and subsequently heated up to the desired supply temperature, which often costs a large amount of energy for overcooling and reheating [1,2]. The desiccant cooling systems are being developed as an alternative to overcome the flaws of vapor compression system [3]. Especially, rotary desiccant air conditioning systems, which are compact and less subject to corrosion and can work continuously, attract more attention [4]. Many configurations of rotary desiccant system are designed and studied and the encouraging results are obtained [5–7]. It is also found that efficient energy saving and low environment impact are achieved when the desiccant wheel is regenerated by means of low temperature thermal energy such as solar heat [8–10], energy from co-generators [11,12], and waste heat [13]. Although the costs of these technologies are near-zero
∗ Corresponding author. Tel.: +86 186 2200 7007. E-mail addresses: [email protected] (Y. Sheng), [email protected] (Y. Zhang), [email protected] (N. Deng), fl@byg.dtu.dk (L. Fang), [email protected] (J. Nie), [email protected] (L. Ma). 0378-7788/$ – see front matter © 2013 Elsevier B.V. All rights reserved. http://dx.doi.org/10.1016/j.enbuild.2013.07.058
[14], the applications of them are limited due to the climate or regional factors. One type of rotary desiccant system is a combination of the rotating desiccant wheel and of the vapor compression system, which can be widely and efficiently used without limitations. The shortcoming is that the thermal heat recovered from the condenser is not sufficient for the desiccant regeneration and a supplementary heater is needed [4], which will increase the energy input and equipment investment. The proposed method in this paper will solve this problem by utilizing the characteristics of high temperature heat pump. Heat pump has a heating capacity at the condenser and a cooling capacity at the evaporator, which is an approach to satisfy simultaneous energy demands for heating and cooling. Many researchers have worked on such systems for simultaneous production in various applications [15–17]. Take advantage of simultaneous production of heating and cooling, heat pump desiccant system as a novel topic has been studied. Lazzarin [18] has investigated a new equipment of self-regenerating liquid desiccant cooling system for supermarket application. Enteria [19] has performed an experimental evaluation of a new solid desiccant heat pump system, the main feature of which is direct impregnation of the desiccant material in the evaporator and condenser of the heat pump system. In Ref. [20], the variable refrigerant volume (VRV) system has been operated in conjunction with a self-regenerating heat pump desiccant unit. And its energy consumption has been analyzed as opposed to the VRV system coupling with heat recovery ventilation
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Nomenclature cp COP HTHP DW V m t h Q E DBT WBT TR PR BY-3 FAR AFR
specific heat of the air (kJ/kg K) coefficient of performance high temperature heat pump desiccant wheel air volumetric flow rate (m3 /h) air quality flow rate (kg/h) air temperature (◦ C) enthalpy of the air (kJ/kg) heat energy (kW) electrical input power (kW) dry bulb temperature (◦ C) wet bulb temperature (◦ C) temperature of refrigerant (◦ C) pressure of refrigerant (MPa) refrigerant of No.3 of BeiYang fresh air ratio (%) air velocity (m/s)
Greek symbols ω air humidity ratio (g/kg) air density (kg/m3 ) Subscripts com compressor sys system proc process air reg regeneration air f fan c, h cooling and heating, respectively W, N states of outdoor air and indoor air, respectively inlet and outlet state of the process air, respectively C, E O state of supply air A, B inlet and outlet state of the regeneration air, respectively
unit. By utilizing the characteristics of high temperature, Hao and Zhang [21] have explored the potential of integrating the moderate or high temperature heat pump subsystem to the desiccant wheel. However the performance of this kind of system is less complex investigated over a wide range of operating conditions. Therefore, a detailed experiment is carried out in this paper with the objectives as follows: • A new high temperature heat pump coupling with solid desiccant wheel (HTHP&DW) system is designed, and the corresponding air conditioning unit is built. • The mixture refrigerant of “BY-3” is involved and tested in order to ascertain that BY-3 can be a candidate for this system. • An experiment is conducted to study the applicability for extensive thermal hygrometric conditions representing most climate zones in China, and the performance of HTHP&DW system. 2. HTHP&DW air conditioning unit design 2.1. HTHP&DW system Fig. 1 shows the schematic diagram of HTHP&DW system. The ventilation system supplies minimum levels of outdoor airflow and re-circulates a large amount of indoor air on the process side. The return air (state N) and outdoor fresh air (state W), with specific ratio of FAR = 20%, are mixed to become the process air (state C) which flows through the desiccant wheel where a large amount of
moisture is removed by the desiccant (state E). The process air is then cooled by the evaporator of HTHP that is controlled for conditioning the room air temperature to a comfortable level (state O). The outdoor fresh air (state W) is used for regenerating the rotating desiccant wheel. The regeneration air is heated up by the condenser of the heat pump (state A) and is expelled after regenerating the desiccant wheel (state B). 2.2. Air conditioning unit design and manufacture The high temperature heat pump, desiccant wheel and fans are the main components of the air conditioning unit. The main specifications of the machines are listed as follows: Desiccant wheel: the desiccant material is silica-gel and the configuration is honeycomb. The wheel is characterized by the following layout: process air passes 75% of the flow area and regeneration air the remaining 25%. The other characteristics of the wheel are: diameter and thickness equal to 450 mm and 200 mm respectively, the nominal rotational speed is 15 r/h. HTHP: The invariable frequency closed piston compressor branded with Copeland made in America is used with the nominal input power of 2 kW. The type of heat exchanger is air source chip tubular. The manual expansion valve with brand of Swagelok is used to adjust the flow rate of the refrigerant. The volume flow rate is in the range of 0–0.084 m3 /h. Fan: the frequency conversion fans are chosen in order to obtain the desired speed. The nominal power of both process and regeneration fans are 0.3 kW and the frequency is 0–50 Hz. Mature commercial products were adopted for assembling this air-conditioning unit. Fig. 2 shows the layout of air conditioning unit and the experiment is carried out on this facility. The following contents should be noted with attention during design and commissioning: (a) Insulation is necessary due to the temperature difference between different air flows. (b) In order to prevent exchange between the process air and regeneration air, elastic material with high temperature resistance should be well sealed on the surface edge of desiccant wheel. Simultaneously harmful gas should not be produced from the elastic material during the operation of unit. (c) The wind resistances of main machines should be taken into consideration during the selection of fans that can ensure the external residual pressure. Besides noise, size and performance of fans should also be considered. (d) The main machines are fixed on a steel frame with scroll wheels on the bottom to meet the requirement of mobile operation. It is expected that our experience with the above unit can facilitate development of commercial product. Compared to the conventional cooling-based dehumidification air conditioning system, this air conditioning unit has the following advantages: • The sensible and latent heat load are handled independently. • This method is used for connecting the desiccant wheel and HTHP to make full use of both heating and cooling from the condenser and evaporator of the heat pump that can provide the thermal energy for desiccant regeneration without additional heat. • The desiccant wheel can be driven only by the heat pump without any heating machine, which will save the cost of equipment investment. • The structure of this air conditioning unit is compact, with the feature of less occupied area.
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Fig. 1. Schematic of HTHP&DW system.
• The system is working in a completely dry environment that will effectively prevent the growth of bacteria. Hence a better indoor air quality can be obtained. 3. Choice of refrigerant For the proposed HTHP, it is assumed that the whole cooling load and regeneration energy would not only provide sufficient energy for heating and cooling, but also meet the demand of high-temperature heat for regeneration. Therefore the choice of refrigerant is critical for the system operation. In this work, the
“BY-3”, a mixed non-azeotropic refrigerant, developed by Tianjin University was chosen. The results of theoretical calculation of this working fluid are listed in Table 1. The REFPROP 8.0-NIST Reference Fluid Properties Database [22] is employed in theoretical calculation. The experiment analysis of BY-3 conducted by Chen and Zhang [23] based on a single-stage vapor compression refrigeration system showed that the highest temperature at outlet of the condenser reached about 85 ◦ C, when the difference between the water temperatures at the condenser outlet and the evaporator inlet was less than 40 ◦ C, COP was larger than 4; and when the difference of that
Fig. 2. Layout of test facility.
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Table 1 Properties of BY-3. Critical temperature (◦ C)
Critical pressure (MPa)
Ozone depletion potential
Global warmth potential
Security level
126.2
4.664
0.04
2 100
Aa 1b
a b
According to TLV-TWA [30], when the refrigerant concentration is equal to or less than 400 × 10−4 %, it is non-toxic. There is no flame spreading in the air if the refrigerant is tested under the condition of 101.3 kPa and 21 ◦ C.
reached 55 ◦ C, the COP still remained 3. The discharge temperature of BY-3 was lower than 100 ◦ C, and the refrigerant vapor pressure kept lower than 1.8 MPa. Based on many experiment results, the rotary desiccant cooling system is more feasible when the required regeneration temperature is low (60–70 ◦ C is sufficient) [24,25]. According to the results of [26], the dehumidification capacity can be 2.5 g/kg when the regeneration temperature is 60 ◦ C, which could satisfy the dehumidification requirement of air conditioning systems. Hence preferably the regeneration temperature is not less than 60 ◦ C. Obviously the air temperature can be raised to over 60 ◦ C if BY3 is involved in the air-source heat pump based on the results of [23]. BY-3 can be a candidate for HTHP&DW system except the pure refrigerant [21]. The detailed discussions on the performance of heat pump using BY-3 are in Section 5.
4. Experimental procedure 4.1. Experimental conditions As shown in Table 2, the experiment is conducted in an extensive thermal hygrometric range that represents most zones in China. The latent loads from indoor air and fresh air are transformed to sensible load during the absorption process, which will increase the inlet air temperature of evaporator of heat pump. More attention should be paid on the question whether the heat pump can assume the large sensible load. Therefore the desired supply air temperature is designed to be 20 ◦ C to verify the cooling capacity of heat pump. Meanwhile the dehumidification capability of this air conditioning unit is also examined and discussed. The volumetric process air flow rate are 270 m3 /h and 540 m3 /h and that of regeneration air flow are 180–350 m3 /h, respectively. Fig. 3 gives an example of air processes on the psychrometric chart at the condition of TW = 34 ◦ C, ωW = 20 g/kg, FAR = 20%. This is also the typical experiment condition designed to meet the extreme weather during the summer in Tianjin of China. The detailed discussion of air properties is in Section 5.1.
4.2. Measurement devices The state of the process air is that of the mixture of outside fresh air and return air with certain ratio. In order to simplify the manufacture of the test facility, the mixed state is achieved by pre-heat and pre-humidification in the treatment tank. The regeneration air is controlled with the adjustable heater and humidifier similarly. The parameters measured in this work include the inlet and outlet air DBT and WDT for both absorption and regeneration, air flow rate, temperature and pressure of refrigerant at inlet and outlet of compressor, input power of compressor and fans. The specifications of the measuring devices are listed in Fig. 2.
4.3. Performance parameter The performance parameters are divided into two categories. The first group deals with the effectiveness of the system in meeting the required indoor conditions. The second category pertains to
efficiency-related indexes, such as COP based on electrical energy input. COP for cooling of heat pump is represented as Qc Ecom
COPcom,c =
(1)
where Qc is the cooling capacity and is defined as Qc = mproc (hE − hO ) = E Vproc (tE − tO )cp
(2)
COP for cooling of the entire air conditioning system is represented as COPsys,c =
Qc E
(3)
In Eq. (3), E is the electrical power used to drive the vapor compressor and fans. Fans are employed to force circulate the regeneration air and process air streams. Hence, E is calculated as E = Ecom + Ef,proc + Ef,reg
(4)
In our experiment study, it is found that the electricity used for fans represents 11% of total electricity consumption for average. Therefore when the performance of this system is evaluated, the electricity input for fans should not be neglected. Heating COP of heat pump is defined as COPcom,h =
Qh Ecom
(5)
In Eq. (5), Qh is the heating capacity, which is calculated as Qh = mreg (hA − hW ) = A Vreg (tA − tW )cp
(6)
The error analysis of the experimental results is performed according to the propagation of the uncertainties. An overall accuracy within the maximum amount of relative uncertainty is 3.73% for COPsys,c , 3.6% for COPcom,c , 4.1% for COPcom,h , 2.01% for Qc and 2.83% for Qh respectively. 5. Results The heat pump has to provide sufficient heating for regeneration and cooling for ventilation air to the conditioned space. In summer, large amount of moisture should be removed from the conditioned space which means that higher regeneration temperature is required. Whether the desired regeneration temperature could be attained becomes the sticking point for whether HTHP&DW system could be a new approach for air conditioning system. Hence the performance of heat pump is the focus in this section. 5.1. Air processes The measurement and calculation report of the air properties for the typical experimental condition of TW = 34 ◦ C, ωW = 20 g/kg, FAR = 20% are shown in Table 3. It is found that air temperature of process air is 20.56 ◦ C which meet the desired supply air temperature (20 ± 1 ◦ C) after the air steam flows through the desiccant wheel. Also the calculated sensible and latent loads for the given processes are listed in Table 4. It is found that the sensible load assumed by DW (process 3) is bigger than that of total sensible load (process 1 + 2 + 4) due to the heat transfer between process air and regeneration air. Furthermore the latent load assumed by DW
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Table 2 Experimental conditions.
Indoor air Outdoor air
Air temperature (◦ C)
Air humidity ratio (g/kg)
Fresh air ratio (%)
Desired supply air temperature (◦ C)
26 28–40
11 8–22
20
20 ± 1
Fig. 3. Air processes on the psychrometric chart.
(process 4) is also bigger than that of total latent load (process 1 + 2). This is because of the mass transferred from the regeneration side to the process side and a minimum leakage of outside air into the air passage. The relative differences of sensible load and latent load are 4% and 5% respectively. In order to explore the dehumidification feasibility of this system used in an extensive climate range and evaluate the performance of the system, a detailed experiment is conducted and
Table 3 Typical measured and calculated state properties of the system shown in Fig. 3. State
DBT (◦ C)
WBT (◦ C)
ω (g/kg)
N W C E O A B
26.00 34.00 27.60 47.33 20.56 63.13 38.17
19.20 27.25 20.83 23.62 15.22 33.23 30.00
11.00 20.00 12.80 8.46 8.46 20.00 23.50
h (kJ/kg) 54.29 85.59 60.53 66.61 42.24 116.08 98.95
Table 4 The calculated sensible and latent load for given processes. No. Processes
1 W→N
2 O→N
3 E→O
4 C→E
Sensible load (kW) Latent load (kg/h) Total sensible load of (1 + 2 + 4) Total latent load of (1 + 2)
0.14 0.58 2.32 1.24
0.39 0.66
2.43 –
1.79 1.31
Note: Vproc = Vreg = 270 m3 /h.
the results are shown in Figs. 4 and 5. Fig. 4 gives the supply air temperature and humidity ratio obtained at different climate conditions, in which it is found that the supply air temperatures (TO ) are in the range as expected (20 ± 1 ◦ C) except when in the extreme hot environment (TW beyond 36 ◦ C). All the supply air humidity ratios (ωO ) are satisfied and the maximum dehumidification efficiency could achieve 38% at the condition of TW = 40 ◦ C, and ωW = 20 g/kg. However, it should be noted that ωO is lower when the outdoor air temperature is higher. This can be attributed to the fact that, when TW is higher, Treg is higher, and dehumidification ability of desiccant wheel increases with the regeneration temperature, resulting in a lower ωO . Fig. 5 shows the regeneration temperature obtained. It is found that most of regeneration temperature (Treg = TA ) are satisfied in the range of 60–66 ◦ C. Therefore this HTHP&DW air conditioning system could be a new approach to provide comfort indoor environment for people. 5.2. Influences of regeneration air flow on the heating capability of heat pump In this section, regardless of supply air temperature (TO ), impact of flow rate of regeneration air (Vreg ) on the behavior of heat pump are investigated under the constant working conditions of Vproc = 540 m3 /h, ωW = 14 g/kg. Fig. 6 shows the variation of regeneration temperature and heating COP against outdoor air temperature at the conditions of different regeneration air flow rate. With the reduction of Vreg , Treg increases that leads to the increase of heat transfer temperature difference between the inlet and outlet air temperature of condenser. Meanwhile the heat transfer coefficient on the air side of
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(b)
the condenser is reduced due to the decrease of the air velocity. As a result, the heating capacity increases slightly. A higher Treg means the condensing temperature should be higher that determines a higher condensing pressure of the compressor. Because of that, the compression ratio is high that increases the input power of compressor. When the air flow decreases, COPcom,h decreases with the rise of input power in the condition of relatively stable heating as confirmed by [27]. It is found that Treg are between 60 ◦ C and 70 ◦ C with a maximum of 74 ◦ C when the volume of air flow equals to 270 m3 /h. Although higher regeneration temperatures are obtained in the condition of Vreg = 180 m3 /h, COPcom,h becomes too low and the compressor will shut down if the regeneration air flow rate is further reduced below 180 m3 /h. Therefore the regeneration air flow rate has significant impact on the performance of heat pump. It is found that a lower regeneration air flow rate results in a lower COPcom,h . 5.3. Global behavior of the system in the extensive climate zone Based on the analysis above, Vproc and Vreg are selected as 270 m3 /h respectively. Apart from the factors analyzed in Section 5.2, outdoor air temperature and humidity ratio do influence the performance of this system.
Fig. 4. Supply air temperature and humidity ratio obtained in the extensive thermal hygrometric environment.
Fig. 5. Regeneration temperature obtained in the extensive thermal hygrometric environment: regeneration temperature equals to the outlet air temperature of condenser.
Fig. 6. Regeneration temperature and heating COP of heat pump against outdoor air temperature.
5.3.1. Effects of outdoor humidity ratio Fig. 7 shows the system performance variations with outdoor air humidity ratio at the constant outdoor air temperature (TW = 34 ◦ C). As ωW increases, cooling capacity of heat pump increases monotonically in the test range of 8–22 g/kg as represented in Fig. 7(a). This is because, with the rise of humidity ratio, the desiccant wheel removes a bigger amount of water vapor, which results in the rise of outlet air temperature (TE ) from the process zone of desiccant wheel and finally results in an increase of the total cooling capacity. In order to ensure the supply air temperature, the opening of expansion valve is increased, which leads to the rise of mass flow of refrigerant. Hence a higher performance is attained as confirmed by [28] when the rise of cooling capacity is more than that of input power of compressor. The input power for fans is constant due to the unchanged regeneration and process air flow rate and therefore COPsys,c shows the same trend as COPcom,c represented in Fig. 7(b). 5.3.2. Effects of outdoor air temperature Fig. 8 shows the variations of cooling capacity, input power of compressor and cooling COP with outdoor air temperature at the constant outdoor air humidity ratio (ωW = 18 g/kg). As represented in Fig. 8(a), with the rise of ambient temperature, cooling capacity increases. This can be attributed to the fact that a higher TW causes a higher Treg , and air could remove more moisture during the regeneration phase. Similarly, during the dehumidification process, the desiccant could adsorb more vapor from the process air, resulting in a rise of total cooling capacity. And the increase of cooling capacity is more than that of input power of compressor. Therefore COPcom,c is higher at a higher outdoor air temperature as confirmed by [28,29]. Fig. 8(b) shows the influence of outdoor air temperature on COP obtained from the experimental study. In order to maintain the normal operation and obtain the desired regeneration temperature, the volume of air flow through the condenser is adjusted higher, which results in the increase of electricity consumption for fans. Therefore COPsys,c does not increase monotonically as COPcom,c because of the rise of input power for fans when TW increases. Then COPsys,c reaches the peak when TW = 34 ◦ C.
Y. Sheng et al. / Energy and Buildings 66 (2013) 505–513
(a)
(b)
Fig. 7. Variation of cooling capacity, input power of heat pump and cooling COP with outdoor air humidity ratio.
511
(a)
(b)
Fig. 8. Variation of cooling capacity, input power of heat pump and cooling COP with outdoor air temperature.
6. Discussions It should be noted that the performances at other tested outdoor temperatures and humidity ratios show the same trends as presented in Figs. 7 and 8. Fig. 9 gives the comparisons of cooling COP between heat pump and whole system. It is found in Fig. 9(a) that COPcom,c calculated at air temperature of 34 ◦ C, 36 ◦ C,38 ◦ C are very close. The reason for this is that the regeneration temperatures are close as shown in Fig. 5 that determines the similar cooling capacity of heat pump for above given explanation for Fig. 8(a). Fig. 9(b) shows the cooling COP of whole system in the extensive thermal hygrometric environment. It is found that COPsys,c get a peak value when TW = 34 ◦ C and ωW = 20 g/kg due to the same reason explained for Fig. 8(b). So it could be concluded that the maximum of COPcom,c is 2.26 when TW = 40 ◦ C and ωW = 20 g/kg; and COPsys,c gets a maximum value of 2.08 when TW = 34 ◦ C and ωW = 20 g/kg. Based on the above analysis, a higher outdoor air temperature and a higher humidity ratio decide a better performance of heat pump. When the operation of heat pump is observed during the experiment, the condensing pressures are below 2.5 MPa in line with the pressure bearing capacity of general air-conditioning unit and the compression ratios are in the range of 5–7 by using the refrigerant of BY-3. There is no reaction between BY-3, refrigerant lubricant, and compressor materials, and this mixture is chemically stable in a wide range of operating temperatures.
Although the feasibility used in extensive climate zone of HTHP&DW air conditioning system is verified, some results are not satisfied such as low COPc . Therefore the author provides some suggestions for optimizing the performance of this system. It is found in Figs. 7(a) and 8(a) that the input power of compressor does not increase in equal proportion with the rise of cooling capacity. This is because the feature of compressor involved in the system is invariable frequency, this kind of compressor cannot change the speed of motor, and then the electricity consumption does not change with the variations of cooling capacity. Therefore the inverter compressor is preferred in order to save energy and improve the performance of heat pump. The precise supply air temperature, humidity ratio and regeneration temperature cannot be obtained without automatic control system. So it is suggested that the automatic control system should be installed in this air conditioning unit. In order to maintain the normal operation of heat pump, large volume of air should flow through the condenser and sequentially pass through the regeneration zone of desiccant wheel, which causes superfluous heat to regenerate the desiccant and therefore the heat transfer between two air streams is strengthened and then increases the burden of heat pump. For this reason the structure of air conditioning unit should be rebuilt so as to expel the unwanted heat from condenser. Furthermore the optimal regeneration air
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Technology of P.R. China in the Sino-Danish collaborated research project: “Activating the Building Construction for Building Environment Control” (Danish International DSF project no. 09-71598, Chinese international collaboration project no. 2010DFA62410).
References
Fig. 9. Comparisons of cooling COP between heat pump and whole system in the extensive thermal hygrometric environment.
velocity, regeneration temperature and proportion of process and regeneration zones should be further studied. 7. Conclusion In this work, HTHP&DW system and corresponding air conditioning unit is built and tested to study the feasibility used in extensive climate zone. Furthermore the performance of heat pump and entire system are analyzed. The property of BY-3 has a strong ability to work in the large temperature difference environment. When BY-3 is involved in the air source heat pump, the supply air temperatures are in the range as expected except when in the extreme hot environment (TW above 36 ◦ C), and the regeneration temperatures can reach more than 60 ◦ C and dehumidification capability are satisfied. HTHP&DW system can be an alternative for the desiccant cooling system that can be used in the extensive thermal hygrometric environment. The ambient thermal hygrometric environment and regeneration air flow rate have great impact on the performance of heat pump. Lower regeneration air flow rate leads to a lower COPcom,h , and higher outdoor air temperature and humidity ratio decide a higher COPcom,c . However, COPc is not quite high with the maximum of 2.26 for heat pump and 2.08 for whole system respectively. Several suggestions provided for optimizing the system are helpful for saving energy and facilitating the development of mature product. Acknowledgements This work is collectively sponsored by the Danish Agency for Science, Technology and Innovation and the Ministry of Science and
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