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An experimental study on a novel radiant-convective heating system based on air source heat pump
Accepted Manuscript Title: An experimental study on a novel radiant-convective heating system based on air source heat pump Authors: Jiankai Dong, Long Zhang, Shiming Deng, Bin Yang, Shun Huang PII: DOI: Reference:
S0378-7788(17)32944-4 https://doi.org/10.1016/j.enbuild.2017.10.065 ENB 8085
To appear in:
ENB
Received date: Revised date: Accepted date:
29-8-2017 28-9-2017 20-10-2017
Please cite this article as: Jiankai Dong, Long Zhang, Shiming Deng, Bin Yang, Shun Huang, An experimental study on a novel radiant-convective heating system based on air source heat pump, Energy and Buildings https://doi.org/10.1016/j.enbuild.2017.10.065 This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
An experimental study on a novel radiant-convective heating system based on air source heat pump
Jiankai Donga, Long Zhanga,b, Shiming Dengb, Bin Yangc,*, Shun Huanga
a
Department of Building Thermal Energy Engineering, Harbin Institute of
Technology, Harbin, China b
Department of Building Services Engineering, The Hong Kong Polytechnic
University, Kowloon, Hong Kong SAR, China c
Department of Applied Physics and Electronics, Umeå University, Umeå, Sweden
* Corresponding author: Bin Yang Email address: [email protected]
Abstract Air source heat pump (ASHP) has been widely applied to many parts of the world due to its simple structure and low initial cost. To save energy consumed for spacing heating and enhance the indoor thermal environment, improving the performances of ASHP has become one of the research focus in the relevant field. Currently, the most conventional heating terminal of ASHP system for spacing heating is finned tube heat exchanger coupled with air fan, which may cause strong draught sensation and dry eye problem and make users feel uncomfortable during convective heating. On the other hand, radiant heating is attracting more and more attention due to its 1
comfortable indoor thermal environment. In this paper, a novel radiant-convective heating terminal was presented and coupled into an ASHP system. Both the operating characteristics and heating performances of the novel system were experimentally investigated. The experimental results showed that the novel system took about 28 min to enter a steady operating stage, during which the radiant panel surface temperature and outlet air temperature for the novel heating terminal, and COP of the novel system were 40.9 oC, 32.1 oC, and 3.11, respectively, under a standard heating condition. In addition, all parameters mentioned above saw a linear increase when the outdoor air temperature increased from -4.0 to 10.0 oC, and their respective rising rates were 0.41 oC, 0.28 oC and 0.04 per increased outdoor air temperature. Furthermore, the experimental results also demonstrated that adjusting the indoor air flow rate could effectively allocate the amount of heat generated by different heat transfer modes, which may have significant effects on the indoor thermal environment.
Keywords: Experimental study; Radiant heating; Convective heating; Air source heat pump
Nomenclature A
Outside surface area of the radiant panel [m2]
Ai
Surface area of one of the unheated walls, floor and ceiling [m2]
ASHP
Air source heat pump 2
AUST
Reference temperature [oC]
cp
Constant-pressure specific heat of air [kJ/(kg·oC)]
cs
Specific heat of copper [kJ/(kg·oC)]
COP
Coefficient of performance
G
Mass flow rate of the refrigerant [kg/h]
H
Height [m]
hin
Inlet refrigerant enthalpy for the heating terminal [kJ/kg]
hout
Outlet refrigerant enthalpy for the heating terminal [kJ/kg]
L
Length [m]
P
Input power of the compressor [W]
PID
Proportion-integration-differentiation
Q fc
Forced convective heating capacity [W]
Qnc
Natural convective heating capacity [W]
Qr
Radiant heating capacity [W]
Qref
Refrigerant side heating capacity [W]
Qs
Energy storage capacity [W]
Qtot
Total heating capacity [W]
Ta
Indoor air temperature [oC]
Ti
Surface temperature of one of the unheated walls, floor and ceiling [oC]
Ti ,air
Ts
Inlet air temperature for the heating terminal [oC] Surface temperature of the radiant panel [oC]
To ,air
Outlet air temperature for the heating terminal [oC]
V
Air volume flow rate for the heating terminal [m3/h]
Vc
Copper volume for the heating terminal [m3]
W
Width [m]
Wo, fan
Input power of outdoor air fan [W]
Wi , fan
Input power of indoor air fan [W] 3
Thickness [m]
a
Indoor air density [kg/m3]
c
Copper density [kg/m3]
Tc
Copper temperature change rate for the heating terminal [oC/s]
1. Introduction Building energy consumption is one of the major concerns for global environmental issues and energy crisis [1]. In 2010, about 32% of the total energy consumption was related to buildings, which resulted in 30% of corresponding CO2 emissions [2]. Especially, there was one-third of the related building energy utilized by space heating systems [3]. Among these systems, air source heat pump (ASHP) has been widely applied to many parts of the world due to its simple structure and low initial cost. To save energy consumed for spacing heating and enhance the indoor thermal environment, improving the performances of ASHP has become one of the research focus in the relevant fields. On the one hand, there are a large number of studies have been carried out on investigating the frosting [4-6], defrosting [7-16] and low ambient heating [17-22] performances of ASHP. For example, Song et al. [4] experimentally investigated the impacts of frosting evenness of outdoor coil on the heating performances of an ASHP system. The results showed that the system COP would increase with an increased frosting evenness value. Amer and Wang [7] presented an overview of the defrosting performances of a wide range of systems including ASHP system. They divided the defrosting methods into two categories: passive and active methods. The former was mainly based on surface treatment, while the latter was comprised of many techniques, 4
such as employing high voltage electric field and ultrasonic vibration. Besides, the related defrosting control strategies were also reviewed in their work. Zhang et al. [17] presented a review of the advances in ASHP system applied to cold climate, in which the related advanced systems were divided into single-stage, dual-stage, and multi-stage compression systems. Both the features and heating performances of each kind of advanced systems were reviewed. In addition, the authors concluded that the quasi-two-stage compression system, one type of the dual-stage compression systems, had the greatest development potential in terms of the heating performances and primary investment. On the other hand, there is a certain amount of literature focused on optimizing the system control strategies, functions and components under typical conditions. Tsai and Tsai [23] reported a cascaded fuzzy control strategy for an ASHP system, which had better performance on disturbance rejection than a cascaded proportional and integral control strategy. After that, their research group improved the algorithm of the cascaded fuzzy control strategy with particle swarm optimization and evolutionary programming, and indicated that the ASHP system with the novel control strategy achieved a better performance [24]. In addition, Dong et al. [25] presented a multi-input extremum seeking control strategy for an ASHP system, which could consume the minimum power to satisfy a required heating load under different ambient temperature by adjusting the fan speed of condenser and evaporator, and the suction superheat of compressor. With respect to the system function, it is clear that a conventional ASHP system 5
could only be used for space heating or cooling at one point. However, Ji et al. [26] investigated a novel ASHP system which had three modes: space heating only, water heating only, and space cooling and water heating simultaneous. The system could be regarded as a combination of a conventional ASHP system and ASHP water heater system, and it achieved better performances both in cost and energy saving. In addition, Byrne et al. [27] also investigated a novel ASHP system that could achieve space heating and cooling simultaneously. Furthermore, Dong et al. [28] proposed a novel ASHP system which could achieve space cooling and hot water supply simultaneously. In terms of the system components, the relevant studies focused on refrigerant and heating terminal. Due to the environmental issues, ozone depletion and global warming, the novel environmental friendly alternative refrigerant was investigated by a number of researchers. Han et al. [29] experimentally investigated the performances of an ASHP system with partial optimization using R32 and L41b as refrigerant, respectively, and compared to those of R410A system. Alabdulkarem et al. [30] carried out an experiment on alternative refrigerants: R32, D2Y60 and L41a, and indicated that R32 and L41a were good alternatives to R410A. Besides, they optimized system components to adapt the alternative refrigerants by establishing system model. Devecioglu [31] calculated the seasonal performances of ASHP systems using R446A, R447A, R452B and R454B, and compared to that using R410A. The comparative results showed that all the seasonal energy efficiency ratios of these four alternative refrigerants were higher than that of R410A. In particular, 6
R452B was more suitable to replace R410A for heating and R446A for cooling. Cheng et al. [32] reported that the mixture of R32 and R1234ze(E) was a potential alternative refrigerant for ASHP system. They investigated the impacts of the mass fraction of R32 and R1234ze(E) on system heating capacity and COP, as well as temperature glide, which was also a key factor for energy efficiency. The most conventional heating terminal of ASHP system for spacing heating was finned tube heat exchanger coupled with air fan. This kind of heating mode has good adaptability for dynamic heating load, but may cause strong draught sensation which makes users feel uncomfortable. The ratios of the averaged radiant, such as ceiling and floor heating terminals, are attracting more and more attention in recent years due to their comfortable indoor thermal environment [33-40]. For example, Kang et al. [34] experimentally investigated the heating performances of a ceiling capillary radiant heating terminal and concluded that the indoor thermal comfort met the level of Class A [38] with vertical air temperature difference lowering than 3.2 oC, PMV and PPD reaching 0.014 and 5.26%. Wei et al. [39] experimentally investigated the operating characteristics of a direct radiant floor heating terminal coupled with an ASHP system, and they reported that the novel system can fulfill the heating requirement and have better economic and environmental benefit. In addition, the heat transfer process between the refrigerant and the floor layers was studied by Ma et al. [40], which could provide a theoretical reference for system design. However, this kind of direct radiant heating terminal also has several drawbacks, such as difficult maintenance and poor adaptability for dynamic heating load due to the 7
radiation-oriented heat transfer mode. To combine the advantages of conventional finned tube heat exchanger based on forced convective heat transfer mode and those of direct radiant heating terminal, a novel radiant-convective heating terminal is presented and coupled into an ASHP system in this paper. And then, both the operating characteristics and heating performances of the novel system under standard heating condition are experimentally investigated. Furthermore, the impacts of outdoor air temperature and indoor air flow rate on the performances of the novel system, such as characteristic temperatures and heating capacities, are also experimentally investigated. Finally, a conclusion is given. 2. Experimentations 2.1. Experimental setup The schematic of the experimental setup is shown in Fig. 1. As seen, the setup was comprised of two environmental chambers, one was used for simulating outdoor environment (outdoor-environmental chamber) and the other for simulating indoor environment
(indoor-environmental
chamber),
and
the
experimental
novel
radiant-convective heating system based on ASHP. The dimensions of the outdoor-environmental and indoor-environmental chambers were 2.0 × 2.0 × 2.0 and 4.0 × 4.0 × 2.8 m (L × W × H), respectively. On the
one
hand,
specific
experimental
thermal
condition
inside
the
outdoor-environmental chamber, with air temperature ranging from -8.0 to 20 oC and air relative humidity from 40 to 90%, can be maintained using separate air conditioning system comprised of a cooler (4.5 kW), an electric heater (3.0 kW), a 8
humidifier (2.0 kg/h), a supply fan and a proportion-integration-differentiation (PID) controller. On the other hand, the indoor-environmental chamber was located inside another room, in which a separate air conditioning system (3.5 kW) was installed. The thermal condition inside the indoor-environmental chamber, with air temperature ranging from 0 to 25 oC, can be maintained by controlling the temperature of the air surrounding the chamber. The experimental novel radiant-convective heating system was made up of a compressor, a four-way valve, an outdoor coil, an accumulator, a thermostatic expansion valve, a filter, a gas-liquid separator and a novel indoor coil (radiant-convective heating terminal). The schematic of the radiant-convective heating terminal is shown in Fig. 2. As seen, the novel heating terminal was comprised of an air inlet and outlet, an air fan, copper pipes and fins, a radiant panel and insulating layer. On the one hand, the copper pipes were in contact with the radiant panel tightly, hence the heat of refrigerant passing through the copper pipes can transfer to the radiant panel and then to the indoor space in the form of radiation and natural convection. On the other hand, the copper pipes were finned and these fins combining the panel and the insulating layer constituted a series of air passages. The heat of refrigerant passing through the copper pipes can also transfer to the fins and then to the indoor space in the form of forced convection with the help of air fan. Therefore, the novel heating terminal can heat indoor space in the form of radiation and convection simultaneously. Compared to conventional radiant heating terminals, such as radiant floor and ceiling, the novel heating terminal was mounted more easily, 9
occupied a smaller area and had shortened thermal response time. Furthermore, the novel heating terminal may have more comfortable indoor thermal environment when compared to conventional forced convection heating terminals, such as fan coil unit. The specifications of the novel system are shown in Table 1. During the experiments, the novel radiant-convective heating terminal was installed in the central zone of one vertical wall and 10 cm between its bottom and the floor. The surface temperatures of the radiant panel, four vertical walls, floor and ceiling were measured by temperature sensors (Pt1000, ±0.1 oC). In addition, the temperatures of the refrigerant, indoor and outdoor air were also measured by the temperature sensors (Pt1000, ±0.1 oC). The inlet and outlet air relative humidity of outdoor coil were measured by humidity sensors (±1.5%), the pressures of the refrigerant by pressure sensors (±0.05 bar), the mass flow rate of the refrigerant by mass flowmeter (±0.2%), the outlet air flow velocity of the novel heating terminal by air flow sensor (±0.2 m/s), and the input power by a power module (±0.2%), respectively. The schematics of the measuring points installed in the experimental system are given in Fig. 3. In addition, the experimental setup had a computerized data acquisition unit so that all the signals from sensors and measuring devices were recorded by the data acquisition unit at an interval of 5 s. 2.2. Experimental procedures Before starting experiments, the air temperatures and relative humidity inside the outdoor-environmental and indoor-environmental chambers were maintained at designed values. Firstly, the compressor and the outdoor air fan were turned on. The 10
refrigerant was compressed and sent to the heating terminal, and then throttled by the thermostatic expansion valve, passed through the outdoor coil and evaporated by the outdoor air, and finally flowed back into the compressor. At the same time, the indoor air fan kept closed until the refrigerant temperature at the outlet of the heating terminal reached 25 oC. The whole heating period was set at 60 min. The operating characteristics and heating performances of the experimental system were firstly investigated under the standard heating condition [41], and the corresponding test was named as Test 1. In addition, to obtain the impacts of outdoor air temperatures on the system heating performances, Tests 2 to 9 were carried out, where the outdoor air relative humidity was designed as 60±3% to avoid frosting. Furthermore, Tests 10 to 11 were also performed, which combining Test 1 can present the impacts of indoor air flow rates on the system heating performances. All the experimental conditions of these tests are shown in Table 2. 2.3. Experimental data reductions The refrigerant side heating capacity for the experimental heating terminal, Qref , could be evaluated by:
Qref
G(hin hout ) 3.6
(1)
Where: G is the mass flow rate of the refrigerant, hin and hout the inlet and outlet refrigerant enthalpy for the heating terminal. The radiant heating capacity for the experimental heating terminal [42], Qr ,could 11
be evaluated by:
4 4 Qr 5 108 Ts 273.15 AUST 273.15 A
(2)
Where: Ts is the surface temperature of the radiant panel, AUST the reference temperature evaluated by Equation (3), A the outside surface area of the radiant panel.
5
AUST
AT i 1 5
i i
(3)
Ti i 1
Where: Ai and Ti are the surface area and temperature of one of the unheated walls, floor and ceiling, respectively. The natural convective heating capacity for the experimental heating terminal [38], Qnc ,could be evaluated by:
Qnc 1.78 Ts Ta
1.32
A
(4)
Where: Ta is the indoor air temperature. The forced convective heating capacity for the experimental heating terminal,
Q fc ,could be evaluated by:
12
Q fc
aVc p (To,air Ti ,air ) 3.6
(5)
Where: a is the indoor air density, V the air volume flow rate for the heating terminal, c p the constant-pressure specific heat of air, Ti ,air and To ,air the inlet and outlet air temperature for the heating terminal. Since the amount of copper used in the experimental heating terminal was large, including copper fins, pipes and panel, the amount of energy stored in the heating terminal should not be ignored. In this paper, the change rate of the energy stored in the heating terminal was defined as energy storage capacity, Qs , which could be evaluated by:
Qs cVc cs Tc
(6)
Where: c is the copper density, Vc the copper volume for the heating terminal, cs the specific heat of copper, Tc the copper temperature change rate for the heating terminal. Particularly, the sum of the radiant, natural and forced convective heating capacities, and energy storage capacity was named as non-refrigerant side heating capacity. The total heating capacity of the experimental system, Qtot , was the averaged value of refrigerant side and non-refrigerant side heating capacities, and could be evaluated by:
13
Qtot
1 Qref Qr Qnc Q fc Qs 2
(7)
The coefficient of performance (COP) of the experimental system could be evaluated by:
COP
Qtot P Wo, fan Wi , fan
(8)
Where: P is the input power of the compressor, Wo, fan and Wi , fan the input power of the outdoor and indoor air fans, respectively.
3. Results and discussions The performances of the experimental system operated under a standard heating condition in terms of Test 1 are shown in Figs. 4-7. In addition, the impacts of outdoor air temperature in terms of Tests 2-9 are shown in Figs. 8-11, and those of indoor air flow rate in terms of Test 1 and Tests 10-12 are shown in Figs. 12-15.
3.1 Performances of the experimental system operated under standard heating condition Fig. 4 shows the time-variations of characteristic temperatures for the heating terminal in Test 1, including inlet and outlet refrigerant and air temperatures, condensing temperature, and surface temperature of the radiant panel. As seen, the inlet refrigerant temperature saw a significant increase before 28 min with a rising 14
rate exceeding 0.5 oC /min. Then the rising rate decreased to lower than 0.2 oC /min from 28 to 60 min, and the average temperature of the inlet refrigerant temperature was 71.1 oC during this period. On the other hand, the outlet refrigerant temperature increased to 25.1 oC at 2 min, according to which the indoor air fan was turned on. After that, it consistently increased to 41.1 oC at 28 min and then remained steady at around this value over the following heating period. As for the refrigerant condensing temperature, it was consistently higher than the outlet refrigerant temperature, and the gap between them was at around 3.0 oC during the whole heating period. By contrast, the surface temperature of the radiant panel was almost the same with the outlet refrigerant temperature from 0 to 4 min, and then the former was consistently lower than the latter from 4 to 60 min and the gap between them was about 0.6 oC. With respect to the inlet and outlet air temperatures, the former was at around 20.0 oC during the whole heating period, and the latter increased to 32.1 oC at 28 min and then remained steady at around this value over the following heating period.
According to the time-variations of characteristic temperatures for the heating terminal, the overall heating period was divided into two stages: starting stage from 0 to 28 min and steady operating stage from 28 to 60 min. It can be seen that the starting stage of the experimental system was longer than that of a conventional ASHP system, which was mainly caused by the novel heating terminal combining radiant and convective heat transfer. However, the starting stage may be shorted by 15
optimizing the system control strategy and components, which would be carried out in the future work. Fig. 5 shows the time-variations of characteristic heating capacities for the heating terminal in Test 1, including refrigerant side and non-refrigerant side heating capacities. As seen, the radiant, natural and forced convective heating capacities increased steadily during starting stage, and then remained stable at around 500.2, 363.1 and 2409.4 W, respectively, during steady operating stage. The ratios of the averaged radiant, natural and forced convective heating capacities were 1.4:1.0:6.7 during the whole heating period. The energy storage capacity experienced a sharp rise and reached 854.3 W at 2 min, and then it saw a decrease from 2 to 28 min due to the open of the indoor air fan. During the steady operating stage, the energy storage capacity remained between -50 to 50W. It was noticeable that the refrigerant side heating capacity evaluated by Equation (1) was much larger than the non-refrigerant side heating capacity, comprised of the radiant, natural and forced convective heating capacities, and energy storage capacity, during the initial 8 min into the heating period. This was because the amount of refrigerant flowing into the heating terminal was larger than that outflowing the heating terminal during this period while the refrigerant of the system was redistributed, so that the refrigerant side heating capacity evaluated by the amount of refrigerant flowing into the heating terminal should be larger than the actual value. Furthermore, the gap between the refrigerant side heating capacity and the non-refrigerant side heating capacity narrowed with time and fluctuated between -11% to 10% from 8 to 60 min, which meant that the 16
refrigerant redistribution of the system completed at about 8 min. The averaged values of refrigerant side and non-refrigerant side heating capacities during steady operating stage were at 3016.8 and 3277.8 W, respectively.
Fig. 6 shows the time-variations of suction and discharge pressure for the compressor in Test 1. As seen, the suction pressure decreased from 0 to 2 min dramatically and reached 0.70 MPa. By contrast, the discharge pressure significantly increased to 2.01 MPa during the same period. Then the indoor air fan was turned on, the suction pressure increased slightly and remained steady at around 0.72 MPa during the following heating period. The discharge pressure consistently increased from 2.01 to 2.81 MPa between 2 and 28 min, and then it remained steady at around 2.82 MPa until the end of the heating period.
As seen in Fig. 7, the input power for compressor saw an increase during the starting stage, which was believed to be caused by the increased discharge pressure. During the steady operating period, the averaged value of input power for compressor was at 845.7 W. Furthermore, as mentioned in the discussions of Fig. 5, the refrigerant side heating capacity was larger than the actual value during the initial 8 min into the heating period. Therefore, the COP of the experimental system during this period was evaluated by: 17
COP
Qr Qnc Q fc Qs P Wo, fan Wi , fan
(9)
As for the following heating period, from 8 to 60 min, the COP for the experimental system was evaluated by Equation (8). It can be seen from Fig. 7, the variation trend of the COP was similar to that of input power for compressor, and the averaged value of COP during steady operating stage was at 3.11.
3.2. Impacts of outdoor air temperature on the performances of the experimental system
The impacts of outdoor air temperatures on the performances of the experimental system were analyzed in terms of characteristic temperatures and heating capacities for the heating terminal, suction and discharge pressure, and input power for the compressor, and COP for the experimental system. Particularly, all the parameters were the averaged values during the steady operating stages in Tests 2-9. In Fig. 8, the characteristics temperatures for the heating terminal under different outdoor air temperatures were compared. As seen, all of the inlet and outlet refrigerant, refrigerant condensing, radiant panel surface and outlet air temperatures saw a linear increase when the outdoor air temperature increased from -4.0 to 10.0 oC, provided that the inlet air temperature remained at around 20.0 oC. Besides, the rising rates of 18
the inlet and outlet refrigerant, refrigerant condensing, radiant panel surface and outlet air temperatures were 0.63, 0.55, 0.51, 0.41 and 0.28 oC per increased outdoor air temperature, respectively.
Fig. 9 shows the impacts of outdoor air temperature on the characteristic heating capacities for the heating terminal. It can be seen that the variation trends of the characteristic heating capacities were similar to those of the characteristic temperatures. The rising rates of the refrigerant side, radiant, natural convective and forced convective heating capacities were 68.9, 10.8, 9.7, and 58.0 W per increased outdoor air temperature, respectively. In addition, it was also noticeable that the ratios of the radiant and forced convective heating capacities remained stable at around 0.21 in Tests 2-9, which meant that the variation of the outdoor air temperatures had no effects on the ratio of the radiant to forced convective heating capacities. On the other hand, the ratios of the radiant and natural convective heating capacities decreased from 1.48 to 1.35 when the outdoor air temperature increased from -4.0 to 10.0 oC. Generally, the ratios of the averaged radiant, natural and forced convective heating capacities were around 1.4:1.0:6.9 in Tests 2-9.
As shown in Fig. 10, the suction and discharge pressures increased with an increase in outdoor air temperature, and their variation trends were nearly linear. The 19
rising rate of the former was 0.02 MPa per increased outdoor air temperature, and that of the latter 0.04 MPa per increased outdoor air temperature. Fig. 11 shows the input power for the compressor and COP for the experimental system in Tests 2-9. As for the input power, its averaged value during steady operating stage was at 714.6 W when the outdoor air temperature was -4.0 oC, and increased to 877.7 W when the outdoor air temperature was 10.0 oC. The COP for the experimental system also increased with an increased outdoor air temperature with a rising rate of 0.04, and the value of the COP was at 3.22 when the outdoor air temperature was 10.0 oC. Impacts of indoor air flow rate on the performances of the experimental system
Similar to the analytical method proposed in Section 3.2, the impacts of indoor air flow rate on the performances of the experimental system were analyzed and presented in this section. Fig. 12 shows the impacts of indoor air flow rate on the characteristic temperatures for the heating terminal. As seen, all the characteristic temperatures except inlet air temperature (maintained at around 20 oC during experiments) decreased with an increase in indoor air flow rate, which was believed to be caused by the decreased discharge pressure for the compressor. The values of inlet and outlet refrigerant, refrigerant condensing, radiant panel surface and outlet air temperatures were at 81.0, 47.9, 50.9, 47.8 and 38.9 oC when the indoor air flow was 288 m3/h, and those decreased to 71.1, 41.5, 44.5, 40.9 and 32.1 oC when the indoor air flow was 551 m3/h. 20
In Fig. 13, the characteristic heating capacities for the heating terminal under different indoor air flow rate were compared. As seen, the radiant and natural convective heating capacities decreased with an increased indoor air flow rate. With respect to the indoor air flow rate 288, 456 and 551 m3/h, the values of the former were at 652.7, 558.5 and 500.2 W, and those of the latter at 557.5, 439.5 and 363.1 W. By contrast, the refrigerant side heating capacity increased from 2919.4 to 3016.8 W and the forced convective capacity increased from 1966.7 to 2409.4 W when the indoor air flow rate increased from 288 to 551 m3/h. Furthermore, it was noticeable that the ratio of the natural convective to radiant heating capacities decreased with the increased indoor air flow rate (0.85 at 288 m3/h and 0.73 at 551 m3/h). The same variation trend was observed on the ratio of the radiant to forced convective heating capacities (0.33 at 288 m3/h and 0.21 at 551 m3/h). On the other hand, the ratio of the forced convective to refrigerant side heating capacities increased with the increased indoor air flow rate (0.67 at 288 m3/h and 0.80 at 551 m3/h).
Fig. 14 shows the suction and discharge pressures for the compressor under different indoor air flow rates. It can be seen that the impact of the indoor air flow rate on the suction pressure was insignificant, and the value of suction pressure remained at around 0.72 MPa. By contrast, the discharge pressure decreased from 3.25 to 2.82 21
MPa when the indoor air flow rate increased from 288 to 551 m3/h. Fig. 15 shows the impacts of indoor air flow rate on the input power for the compressor and COP for the experimental system. As seen, when the indoor air flow rated increased from 288 to 551 m3/h, the input power decreased from 948.4 to 845.7 W and the COP increased from 2.86 to 3.11. The above analysis suggested that the indoor air flow rate had significant impacts on the performances of the experimental system. Particularly, adjusting indoor air flow rate could effectively allocate the amount of heat generated through different heat transfer modes, which may have significant effects on the indoor thermal environment.
3. Conclusions To combine the advantages of conventional finned tube heat exchanger based on forced convective heat transfer mode and those of direct radiant heating terminal, an experimental study on a novel radiant-convective heating system based on ASHP was carried out in this paper. Both the operating characteristics and heating performances of the novel system under standard heating condition were experimentally investigated. Furthermore, the impacts of outdoor air temperature and indoor air flow rate on the performances of the novel system, such as characteristic temperatures and heating capacities, were also experimentally investigated. The following conclusions were made: 22
1. The novel system took about 28 min to enter a steady operating stage, which means that it had a longer starting stage than that of a conventional ASHP system due to novel radiant-convective heating terminal. 2. During steady operating stage under the standard heating condition, the radiant panel surface temperature and outlet air temperature for the novel heating terminal were 40.9 and 32.1 oC, respectively. In addition, the radiant, natural convective, forced convective heating capacities were 500.2, 363.1 and 2409.5 W, respectively. The COP of the novel system was 3.11. 3. The radiant panel surface and outlet air temperatures for the heating terminal, and radiant, natural convective and forced convective heating capacities, and COP of the novel system saw a linear increase when the outdoor air temperature increased from -4.0 to 10.0 oC. Their respective rising rates were 0.41 oC, 0.28 oC 10.8 W, 9.7 W, 58.0 W and 0.04 per increased outdoor air temperature. 4. The indoor air flow rate had significant impacts on the performances of the novel system. In particular, adjusting the indoor air flow rate could effectively allocate the amount of heat generated through different heat transfer modes, which may have significant effects on the indoor thermal environment. Generally, the feasibility and stability of the novel heating system have been verified in this work. The investigation on optimizing of the control strategy and components for the novel system, as well as its indoor thermal environment will be carried out in the future work. 23
Acknowledgements The authors are grateful to the financial supports from “National Natural Science Foundation of China (Grant No. 51608146)” for the work reported in this paper. References: [1] Grubler A, Johansson TB, Mundaca L, Nakicenovic N, Pachauri S, Riahi K, et al. Energy primer. In: Global energy assessment - toward a sustainable future. Cambridge, UK and New York, NY, USA: Cambridge University Press; Laxenburg, Austria: International Institute for Applied Systems Analysis. 2012. [2] International Energy Agency. Energy technology perspectives 2012 OECD/IEA, Paris (2012). [3] Ürge-Vorsatz D, Eyre N, Graham P, Harvey D, Hertwich E, Jiang Y, et al. Chapter 10-energy end-use: building. In: Global energy assessment - towarda sustainable future. Cambridge, UK and New York, NY, USA: Cambridge University Press; Laxenburg, Austria: International Institute for Applied Systems Analysis. 2012: 649−760. [4] Song MJ, Xia L, Mao N, Deng SM. An experimental study on even frosting performance of an air source heat pump unit with a multi-circuit outdoor coil. Applied Energy. 2016; 164: 36−44. [5] Tang JC, Gong GC, Su H, Wu FH, Herman C. Performance evaluation of a novel method of frost prevention and retardation for air source heat pumps using the orthogonal experiment design method. Applied Energy. 2016; 169: 696−708. [6] Jiang YQ, Fu HY, Yao Y, Yan L, Gao Q. Experimental study on concentration change of spray solution used for a novel non-frosting air source heat pump system. Energy and Buildings. 2014; 68: 707−712. [7] Amer M, Wang CC. Review of defrosting methods. Renewable and Sustainable Energy Reviews. 2017; 73: 53−74. 24
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Outdoor-environmental chamber
Indoor-environmental chamber
Experimental novel radiant-convective heating system Radiant-convective heating terminal Four-way valve
Electric heater
Outdoor coil Compressor Gas-liquid separator
Humidifier
Accumulator
Thermostatic expansion valve Filter
Cooler
Fig. 1. The schematic of the experimental setup.
28
Air inlet Copper fin Copper pipe Air fan Copper fin pitch Insulating layer
Copper tube spacing
Fin width Radiant panel
Fin length
Refrigerant inlet Refrigerant outlet
Air outlet
Fig. 2. The schematic of the radiant-convective heating terminal.
P T G
Four-way valve
T
T P
Outdoor coil
T
T
T
P T
T H
Radiant-convective heating terminal
T P
Compressor T H
T V
T
Gas-liquid separator
T
T
T T P
Accumulator
Thermostatic expansion valve Filter
P T
T
P
H
V
Temperature
Pressure
Humidity
Velocity
G Mass flow rate
Fig. 3. The schematics of the measuring points installed in the experimental system.
29
75 Indoor air fan was turned on
70
Inlet refrigerant Outlet refrigerant Refrigerant condensing Radiant panel surface Inlet air Outlet air
65 60
Temperature (oC)
55 50 Starting stage 45
Steady operating stage
40 35 30 25 20 15 0
5
10
15
20
25
30
35
40
45
50
55
60
Times (min)
Fig. 4. Time-variations of characteristic temperatures for the heating terminal in Test 1. 3000
2500 Refrigerant redistribution
Capacity (W)
2000
Starting stage
Steady operating stage
1500
Refrigerant side heating Radiant heating Natural convective heating Forced convective heating Energy storage
1000
500
0 0
5
10
15
20
25
30
35
40
45
50
55
60
Times (min)
Fig. 5. Time-variations of characteristic heating capacities for the heating terminal in Test 1.
30
3.0
2.5
Pressure (MPa)
Starting stage
Steady operating stage
2.0
Suction pressure Discharge pressure
1.5
Indoor air fan was turned on
1.0
0.5 0
5
10
15
20
25
30
35
40
45
50
55
60
Times (min)
Fig. 6. Time-variations of suction and discharge pressure for the compressor in Test 1. 900
5
750 Starting stage
4
Steady operating stage
3
COP
Input Power (W)
600
450 2
Input power for compressor COP for experimental system
Evaluated by Equa. (8) 300 Evaluated by Equa. (9)
150
1
0
0 0
5
10
15
20
25
30
35
40
45
50
55
60
Times (min)
Fig. 7. Time-variations of input power for compressor and COP for the experimental system in Test 1.
31
75 70 65
Inlet refrigerant Outlet refrigerant Refrigerant condensing Radiant panel surface Inlet air Outlet air
60
Temperature (oC)
55 50 45 40 35 30 25 20 15 -4 (Test 2)
-2 (Test 3)
0 (Test 4)
2 (Test 5)
4 (Test 6)
6 (Test 7)
8 (Test8)
10 (Test 9)
Outdoor air temperature (oC)
Fig. 8. The characteristic temperatures for the heating terminal in Tests 2-9.
3000
Heating capacity (W)
2500
2000 Refrigerant side Radiant Natural convective Forced convective
1500
1000
500
0 -4 (Test 2)
-2 (Test 3)
0 (Test 4)
2 (Test 5)
4 (Test 6)
6 (Test 7)
8 (Test8)
10 (Test 9)
Outdoor air temperature (oC)
Fig. 9. The characteristic heating capacities for the heating terminal in Tests 2-9.
32
3.0
2.7
Pressure (MPa)
2.4 Scution pressure Discharge pressure
2.1
0.9
0.6
0.3 -4 (Test 2)
-2 (Test 3)
0 2 4 6 (Test 4) (Test 5) (Test 6) (Test 7) Outdoor air temperature (℃ )
8 (Test8)
10 (Test 9)
Fig. 10. The suction and discharge pressures for the compressor in Tests 2-9.
900
4.0
850 3.5
750
3.0
COP
Input Power (W)
800
700 Input power for compressor COP for experimental system
2.5
650
600
2.0 -4 (Test 2)
-2 (Test 3)
0 (Test 4)
2 (Test 5)
4 (Test 6)
6 (Test 7)
8 (Test8)
10 (Test 9)
Outdoor air temperature (oC)
Fig. 11. The input power for the compressor and COP for the experimental system in Tests 2-9.
33
100 90
Inlet refrigerant
Outlet refrigerant
Outlet refrigerant
Radiant panel surface
Inlet air
Outlet air
80
Temperature (oC)
70 60 50 40 30 20 10 0 288 (Test 11)
551 (Test 1)
456 (Test 10) 3
Indoor air flow rate (m /h)
Fig. 12. The characteristic temperatures for the heating terminal in Tests 1, 10-11. 3500
Refrigerant side
Radiant
Natural convective
Forced convective
3000
Heating capacity (W)
2500
2000
1500
1000
500
0 288 (Test 10)
551 (Test 1)
456 (Test 11) 3
Indoor air flow rate (m /h)
Fig. 13. The characteristic heating capacities for the heating terminal in Tests 1, 10-11.
34
Suction pressure
3.5
Discharge pressure
3.0
Pressure (MPa)
2.5
2.0
1.5
1.0
0.5
0.0 288 (Test 10)
551 (Test 1)
456 (Test 11) 3
Indoor air flow rate (m /h)
Fig. 14. The suction and discharge pressures for the compressor in Tests 1, 10-11.
1200
5 Input power for compressor
COP for experimental system
1000
4
3 600
COP
Input power (W)
800
2 400
1
200
0
0 288 (Test 11)
456 (Test 10)
551 (Test 1)
Indoor air flow rate (m3/h)
Fig. 15. The input power for the compressor and COP for the experimental system in Tests 1, 10-11.
35
Table 1 Specifications of the novel system. Items
Parameters
Values
Units
Power type
1-220-50
φ-V-Hz
Input power
885
W
Dimension
700×535×235
mm(L×H×W)
Air flow rate
1750
m3/h
Accumulator
Volume
0.6
L
Thermostatic expansion valve
Type
EMERSON-B
Filter
Type
C-083-S
Refrigerant
Type
R410A
Air flow rate
0-551
m3/h
Copper tube spacing
70
mm
9.52
mm
Copper pipe thickness
0.7
mm
Number of circuits
3
Copper fin pitch
40
mm
Copper fin width
50
mm
Copper fin length
2000
mm
Copper fin thickness
0.3
mm
2000×1600×0.3
mm(L×W×δ)
Compressor
Outdoor coil
Copper pipe external diameter
Novel radiant-convective heating terminal
Radiant panel dimension Radiant panel material
36
Copper
Table 2 Experimental conditions for Tests 1 to 11.
Test
Outdoor air temperature
Outdoor air relative humidity
Inlet air temperature for
Air flow rate for the
the heating terminal
heating terminal
1
7.0±0.3 oC
86±3 %
20.0±0.3 oC
551 m3/h
2
-4.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
3
-2.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
4
0.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
5
2.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
6
4.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
7
6.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
8
8.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
9
10.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
10
7.0±0.3 oC
86±3 %
20.0±0.3 oC
456 m3/h
11
7.0±0.3 oC
86±3 %
20.0±0.3 oC
37
288 3/h
S0378-7788(17)32944-4 https://doi.org/10.1016/j.enbuild.2017.10.065 ENB 8085
To appear in:
ENB
Received date: Revised date: Accepted date:
29-8-2017 28-9-2017 20-10-2017
Please cite this article as: Jiankai Dong, Long Zhang, Shiming Deng, Bin Yang, Shun Huang, An experimental study on a novel radiant-convective heating system based on air source heat pump, Energy and Buildings https://doi.org/10.1016/j.enbuild.2017.10.065 This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
An experimental study on a novel radiant-convective heating system based on air source heat pump
Jiankai Donga, Long Zhanga,b, Shiming Dengb, Bin Yangc,*, Shun Huanga
a
Department of Building Thermal Energy Engineering, Harbin Institute of
Technology, Harbin, China b
Department of Building Services Engineering, The Hong Kong Polytechnic
University, Kowloon, Hong Kong SAR, China c
Department of Applied Physics and Electronics, Umeå University, Umeå, Sweden
* Corresponding author: Bin Yang Email address: [email protected]
Abstract Air source heat pump (ASHP) has been widely applied to many parts of the world due to its simple structure and low initial cost. To save energy consumed for spacing heating and enhance the indoor thermal environment, improving the performances of ASHP has become one of the research focus in the relevant field. Currently, the most conventional heating terminal of ASHP system for spacing heating is finned tube heat exchanger coupled with air fan, which may cause strong draught sensation and dry eye problem and make users feel uncomfortable during convective heating. On the other hand, radiant heating is attracting more and more attention due to its 1
comfortable indoor thermal environment. In this paper, a novel radiant-convective heating terminal was presented and coupled into an ASHP system. Both the operating characteristics and heating performances of the novel system were experimentally investigated. The experimental results showed that the novel system took about 28 min to enter a steady operating stage, during which the radiant panel surface temperature and outlet air temperature for the novel heating terminal, and COP of the novel system were 40.9 oC, 32.1 oC, and 3.11, respectively, under a standard heating condition. In addition, all parameters mentioned above saw a linear increase when the outdoor air temperature increased from -4.0 to 10.0 oC, and their respective rising rates were 0.41 oC, 0.28 oC and 0.04 per increased outdoor air temperature. Furthermore, the experimental results also demonstrated that adjusting the indoor air flow rate could effectively allocate the amount of heat generated by different heat transfer modes, which may have significant effects on the indoor thermal environment.
Keywords: Experimental study; Radiant heating; Convective heating; Air source heat pump
Nomenclature A
Outside surface area of the radiant panel [m2]
Ai
Surface area of one of the unheated walls, floor and ceiling [m2]
ASHP
Air source heat pump 2
AUST
Reference temperature [oC]
cp
Constant-pressure specific heat of air [kJ/(kg·oC)]
cs
Specific heat of copper [kJ/(kg·oC)]
COP
Coefficient of performance
G
Mass flow rate of the refrigerant [kg/h]
H
Height [m]
hin
Inlet refrigerant enthalpy for the heating terminal [kJ/kg]
hout
Outlet refrigerant enthalpy for the heating terminal [kJ/kg]
L
Length [m]
P
Input power of the compressor [W]
PID
Proportion-integration-differentiation
Q fc
Forced convective heating capacity [W]
Qnc
Natural convective heating capacity [W]
Qr
Radiant heating capacity [W]
Qref
Refrigerant side heating capacity [W]
Qs
Energy storage capacity [W]
Qtot
Total heating capacity [W]
Ta
Indoor air temperature [oC]
Ti
Surface temperature of one of the unheated walls, floor and ceiling [oC]
Ti ,air
Ts
Inlet air temperature for the heating terminal [oC] Surface temperature of the radiant panel [oC]
To ,air
Outlet air temperature for the heating terminal [oC]
V
Air volume flow rate for the heating terminal [m3/h]
Vc
Copper volume for the heating terminal [m3]
W
Width [m]
Wo, fan
Input power of outdoor air fan [W]
Wi , fan
Input power of indoor air fan [W] 3
Thickness [m]
a
Indoor air density [kg/m3]
c
Copper density [kg/m3]
Tc
Copper temperature change rate for the heating terminal [oC/s]
1. Introduction Building energy consumption is one of the major concerns for global environmental issues and energy crisis [1]. In 2010, about 32% of the total energy consumption was related to buildings, which resulted in 30% of corresponding CO2 emissions [2]. Especially, there was one-third of the related building energy utilized by space heating systems [3]. Among these systems, air source heat pump (ASHP) has been widely applied to many parts of the world due to its simple structure and low initial cost. To save energy consumed for spacing heating and enhance the indoor thermal environment, improving the performances of ASHP has become one of the research focus in the relevant fields. On the one hand, there are a large number of studies have been carried out on investigating the frosting [4-6], defrosting [7-16] and low ambient heating [17-22] performances of ASHP. For example, Song et al. [4] experimentally investigated the impacts of frosting evenness of outdoor coil on the heating performances of an ASHP system. The results showed that the system COP would increase with an increased frosting evenness value. Amer and Wang [7] presented an overview of the defrosting performances of a wide range of systems including ASHP system. They divided the defrosting methods into two categories: passive and active methods. The former was mainly based on surface treatment, while the latter was comprised of many techniques, 4
such as employing high voltage electric field and ultrasonic vibration. Besides, the related defrosting control strategies were also reviewed in their work. Zhang et al. [17] presented a review of the advances in ASHP system applied to cold climate, in which the related advanced systems were divided into single-stage, dual-stage, and multi-stage compression systems. Both the features and heating performances of each kind of advanced systems were reviewed. In addition, the authors concluded that the quasi-two-stage compression system, one type of the dual-stage compression systems, had the greatest development potential in terms of the heating performances and primary investment. On the other hand, there is a certain amount of literature focused on optimizing the system control strategies, functions and components under typical conditions. Tsai and Tsai [23] reported a cascaded fuzzy control strategy for an ASHP system, which had better performance on disturbance rejection than a cascaded proportional and integral control strategy. After that, their research group improved the algorithm of the cascaded fuzzy control strategy with particle swarm optimization and evolutionary programming, and indicated that the ASHP system with the novel control strategy achieved a better performance [24]. In addition, Dong et al. [25] presented a multi-input extremum seeking control strategy for an ASHP system, which could consume the minimum power to satisfy a required heating load under different ambient temperature by adjusting the fan speed of condenser and evaporator, and the suction superheat of compressor. With respect to the system function, it is clear that a conventional ASHP system 5
could only be used for space heating or cooling at one point. However, Ji et al. [26] investigated a novel ASHP system which had three modes: space heating only, water heating only, and space cooling and water heating simultaneous. The system could be regarded as a combination of a conventional ASHP system and ASHP water heater system, and it achieved better performances both in cost and energy saving. In addition, Byrne et al. [27] also investigated a novel ASHP system that could achieve space heating and cooling simultaneously. Furthermore, Dong et al. [28] proposed a novel ASHP system which could achieve space cooling and hot water supply simultaneously. In terms of the system components, the relevant studies focused on refrigerant and heating terminal. Due to the environmental issues, ozone depletion and global warming, the novel environmental friendly alternative refrigerant was investigated by a number of researchers. Han et al. [29] experimentally investigated the performances of an ASHP system with partial optimization using R32 and L41b as refrigerant, respectively, and compared to those of R410A system. Alabdulkarem et al. [30] carried out an experiment on alternative refrigerants: R32, D2Y60 and L41a, and indicated that R32 and L41a were good alternatives to R410A. Besides, they optimized system components to adapt the alternative refrigerants by establishing system model. Devecioglu [31] calculated the seasonal performances of ASHP systems using R446A, R447A, R452B and R454B, and compared to that using R410A. The comparative results showed that all the seasonal energy efficiency ratios of these four alternative refrigerants were higher than that of R410A. In particular, 6
R452B was more suitable to replace R410A for heating and R446A for cooling. Cheng et al. [32] reported that the mixture of R32 and R1234ze(E) was a potential alternative refrigerant for ASHP system. They investigated the impacts of the mass fraction of R32 and R1234ze(E) on system heating capacity and COP, as well as temperature glide, which was also a key factor for energy efficiency. The most conventional heating terminal of ASHP system for spacing heating was finned tube heat exchanger coupled with air fan. This kind of heating mode has good adaptability for dynamic heating load, but may cause strong draught sensation which makes users feel uncomfortable. The ratios of the averaged radiant, such as ceiling and floor heating terminals, are attracting more and more attention in recent years due to their comfortable indoor thermal environment [33-40]. For example, Kang et al. [34] experimentally investigated the heating performances of a ceiling capillary radiant heating terminal and concluded that the indoor thermal comfort met the level of Class A [38] with vertical air temperature difference lowering than 3.2 oC, PMV and PPD reaching 0.014 and 5.26%. Wei et al. [39] experimentally investigated the operating characteristics of a direct radiant floor heating terminal coupled with an ASHP system, and they reported that the novel system can fulfill the heating requirement and have better economic and environmental benefit. In addition, the heat transfer process between the refrigerant and the floor layers was studied by Ma et al. [40], which could provide a theoretical reference for system design. However, this kind of direct radiant heating terminal also has several drawbacks, such as difficult maintenance and poor adaptability for dynamic heating load due to the 7
radiation-oriented heat transfer mode. To combine the advantages of conventional finned tube heat exchanger based on forced convective heat transfer mode and those of direct radiant heating terminal, a novel radiant-convective heating terminal is presented and coupled into an ASHP system in this paper. And then, both the operating characteristics and heating performances of the novel system under standard heating condition are experimentally investigated. Furthermore, the impacts of outdoor air temperature and indoor air flow rate on the performances of the novel system, such as characteristic temperatures and heating capacities, are also experimentally investigated. Finally, a conclusion is given. 2. Experimentations 2.1. Experimental setup The schematic of the experimental setup is shown in Fig. 1. As seen, the setup was comprised of two environmental chambers, one was used for simulating outdoor environment (outdoor-environmental chamber) and the other for simulating indoor environment
(indoor-environmental
chamber),
and
the
experimental
novel
radiant-convective heating system based on ASHP. The dimensions of the outdoor-environmental and indoor-environmental chambers were 2.0 × 2.0 × 2.0 and 4.0 × 4.0 × 2.8 m (L × W × H), respectively. On the
one
hand,
specific
experimental
thermal
condition
inside
the
outdoor-environmental chamber, with air temperature ranging from -8.0 to 20 oC and air relative humidity from 40 to 90%, can be maintained using separate air conditioning system comprised of a cooler (4.5 kW), an electric heater (3.0 kW), a 8
humidifier (2.0 kg/h), a supply fan and a proportion-integration-differentiation (PID) controller. On the other hand, the indoor-environmental chamber was located inside another room, in which a separate air conditioning system (3.5 kW) was installed. The thermal condition inside the indoor-environmental chamber, with air temperature ranging from 0 to 25 oC, can be maintained by controlling the temperature of the air surrounding the chamber. The experimental novel radiant-convective heating system was made up of a compressor, a four-way valve, an outdoor coil, an accumulator, a thermostatic expansion valve, a filter, a gas-liquid separator and a novel indoor coil (radiant-convective heating terminal). The schematic of the radiant-convective heating terminal is shown in Fig. 2. As seen, the novel heating terminal was comprised of an air inlet and outlet, an air fan, copper pipes and fins, a radiant panel and insulating layer. On the one hand, the copper pipes were in contact with the radiant panel tightly, hence the heat of refrigerant passing through the copper pipes can transfer to the radiant panel and then to the indoor space in the form of radiation and natural convection. On the other hand, the copper pipes were finned and these fins combining the panel and the insulating layer constituted a series of air passages. The heat of refrigerant passing through the copper pipes can also transfer to the fins and then to the indoor space in the form of forced convection with the help of air fan. Therefore, the novel heating terminal can heat indoor space in the form of radiation and convection simultaneously. Compared to conventional radiant heating terminals, such as radiant floor and ceiling, the novel heating terminal was mounted more easily, 9
occupied a smaller area and had shortened thermal response time. Furthermore, the novel heating terminal may have more comfortable indoor thermal environment when compared to conventional forced convection heating terminals, such as fan coil unit. The specifications of the novel system are shown in Table 1. During the experiments, the novel radiant-convective heating terminal was installed in the central zone of one vertical wall and 10 cm between its bottom and the floor. The surface temperatures of the radiant panel, four vertical walls, floor and ceiling were measured by temperature sensors (Pt1000, ±0.1 oC). In addition, the temperatures of the refrigerant, indoor and outdoor air were also measured by the temperature sensors (Pt1000, ±0.1 oC). The inlet and outlet air relative humidity of outdoor coil were measured by humidity sensors (±1.5%), the pressures of the refrigerant by pressure sensors (±0.05 bar), the mass flow rate of the refrigerant by mass flowmeter (±0.2%), the outlet air flow velocity of the novel heating terminal by air flow sensor (±0.2 m/s), and the input power by a power module (±0.2%), respectively. The schematics of the measuring points installed in the experimental system are given in Fig. 3. In addition, the experimental setup had a computerized data acquisition unit so that all the signals from sensors and measuring devices were recorded by the data acquisition unit at an interval of 5 s. 2.2. Experimental procedures Before starting experiments, the air temperatures and relative humidity inside the outdoor-environmental and indoor-environmental chambers were maintained at designed values. Firstly, the compressor and the outdoor air fan were turned on. The 10
refrigerant was compressed and sent to the heating terminal, and then throttled by the thermostatic expansion valve, passed through the outdoor coil and evaporated by the outdoor air, and finally flowed back into the compressor. At the same time, the indoor air fan kept closed until the refrigerant temperature at the outlet of the heating terminal reached 25 oC. The whole heating period was set at 60 min. The operating characteristics and heating performances of the experimental system were firstly investigated under the standard heating condition [41], and the corresponding test was named as Test 1. In addition, to obtain the impacts of outdoor air temperatures on the system heating performances, Tests 2 to 9 were carried out, where the outdoor air relative humidity was designed as 60±3% to avoid frosting. Furthermore, Tests 10 to 11 were also performed, which combining Test 1 can present the impacts of indoor air flow rates on the system heating performances. All the experimental conditions of these tests are shown in Table 2. 2.3. Experimental data reductions The refrigerant side heating capacity for the experimental heating terminal, Qref , could be evaluated by:
Qref
G(hin hout ) 3.6
(1)
Where: G is the mass flow rate of the refrigerant, hin and hout the inlet and outlet refrigerant enthalpy for the heating terminal. The radiant heating capacity for the experimental heating terminal [42], Qr ,could 11
be evaluated by:
4 4 Qr 5 108 Ts 273.15 AUST 273.15 A
(2)
Where: Ts is the surface temperature of the radiant panel, AUST the reference temperature evaluated by Equation (3), A the outside surface area of the radiant panel.
5
AUST
AT i 1 5
i i
(3)
Ti i 1
Where: Ai and Ti are the surface area and temperature of one of the unheated walls, floor and ceiling, respectively. The natural convective heating capacity for the experimental heating terminal [38], Qnc ,could be evaluated by:
Qnc 1.78 Ts Ta
1.32
A
(4)
Where: Ta is the indoor air temperature. The forced convective heating capacity for the experimental heating terminal,
Q fc ,could be evaluated by:
12
Q fc
aVc p (To,air Ti ,air ) 3.6
(5)
Where: a is the indoor air density, V the air volume flow rate for the heating terminal, c p the constant-pressure specific heat of air, Ti ,air and To ,air the inlet and outlet air temperature for the heating terminal. Since the amount of copper used in the experimental heating terminal was large, including copper fins, pipes and panel, the amount of energy stored in the heating terminal should not be ignored. In this paper, the change rate of the energy stored in the heating terminal was defined as energy storage capacity, Qs , which could be evaluated by:
Qs cVc cs Tc
(6)
Where: c is the copper density, Vc the copper volume for the heating terminal, cs the specific heat of copper, Tc the copper temperature change rate for the heating terminal. Particularly, the sum of the radiant, natural and forced convective heating capacities, and energy storage capacity was named as non-refrigerant side heating capacity. The total heating capacity of the experimental system, Qtot , was the averaged value of refrigerant side and non-refrigerant side heating capacities, and could be evaluated by:
13
Qtot
1 Qref Qr Qnc Q fc Qs 2
(7)
The coefficient of performance (COP) of the experimental system could be evaluated by:
COP
Qtot P Wo, fan Wi , fan
(8)
Where: P is the input power of the compressor, Wo, fan and Wi , fan the input power of the outdoor and indoor air fans, respectively.
3. Results and discussions The performances of the experimental system operated under a standard heating condition in terms of Test 1 are shown in Figs. 4-7. In addition, the impacts of outdoor air temperature in terms of Tests 2-9 are shown in Figs. 8-11, and those of indoor air flow rate in terms of Test 1 and Tests 10-12 are shown in Figs. 12-15.
3.1 Performances of the experimental system operated under standard heating condition Fig. 4 shows the time-variations of characteristic temperatures for the heating terminal in Test 1, including inlet and outlet refrigerant and air temperatures, condensing temperature, and surface temperature of the radiant panel. As seen, the inlet refrigerant temperature saw a significant increase before 28 min with a rising 14
rate exceeding 0.5 oC /min. Then the rising rate decreased to lower than 0.2 oC /min from 28 to 60 min, and the average temperature of the inlet refrigerant temperature was 71.1 oC during this period. On the other hand, the outlet refrigerant temperature increased to 25.1 oC at 2 min, according to which the indoor air fan was turned on. After that, it consistently increased to 41.1 oC at 28 min and then remained steady at around this value over the following heating period. As for the refrigerant condensing temperature, it was consistently higher than the outlet refrigerant temperature, and the gap between them was at around 3.0 oC during the whole heating period. By contrast, the surface temperature of the radiant panel was almost the same with the outlet refrigerant temperature from 0 to 4 min, and then the former was consistently lower than the latter from 4 to 60 min and the gap between them was about 0.6 oC. With respect to the inlet and outlet air temperatures, the former was at around 20.0 oC during the whole heating period, and the latter increased to 32.1 oC at 28 min and then remained steady at around this value over the following heating period.
According to the time-variations of characteristic temperatures for the heating terminal, the overall heating period was divided into two stages: starting stage from 0 to 28 min and steady operating stage from 28 to 60 min. It can be seen that the starting stage of the experimental system was longer than that of a conventional ASHP system, which was mainly caused by the novel heating terminal combining radiant and convective heat transfer. However, the starting stage may be shorted by 15
optimizing the system control strategy and components, which would be carried out in the future work. Fig. 5 shows the time-variations of characteristic heating capacities for the heating terminal in Test 1, including refrigerant side and non-refrigerant side heating capacities. As seen, the radiant, natural and forced convective heating capacities increased steadily during starting stage, and then remained stable at around 500.2, 363.1 and 2409.4 W, respectively, during steady operating stage. The ratios of the averaged radiant, natural and forced convective heating capacities were 1.4:1.0:6.7 during the whole heating period. The energy storage capacity experienced a sharp rise and reached 854.3 W at 2 min, and then it saw a decrease from 2 to 28 min due to the open of the indoor air fan. During the steady operating stage, the energy storage capacity remained between -50 to 50W. It was noticeable that the refrigerant side heating capacity evaluated by Equation (1) was much larger than the non-refrigerant side heating capacity, comprised of the radiant, natural and forced convective heating capacities, and energy storage capacity, during the initial 8 min into the heating period. This was because the amount of refrigerant flowing into the heating terminal was larger than that outflowing the heating terminal during this period while the refrigerant of the system was redistributed, so that the refrigerant side heating capacity evaluated by the amount of refrigerant flowing into the heating terminal should be larger than the actual value. Furthermore, the gap between the refrigerant side heating capacity and the non-refrigerant side heating capacity narrowed with time and fluctuated between -11% to 10% from 8 to 60 min, which meant that the 16
refrigerant redistribution of the system completed at about 8 min. The averaged values of refrigerant side and non-refrigerant side heating capacities during steady operating stage were at 3016.8 and 3277.8 W, respectively.
Fig. 6 shows the time-variations of suction and discharge pressure for the compressor in Test 1. As seen, the suction pressure decreased from 0 to 2 min dramatically and reached 0.70 MPa. By contrast, the discharge pressure significantly increased to 2.01 MPa during the same period. Then the indoor air fan was turned on, the suction pressure increased slightly and remained steady at around 0.72 MPa during the following heating period. The discharge pressure consistently increased from 2.01 to 2.81 MPa between 2 and 28 min, and then it remained steady at around 2.82 MPa until the end of the heating period.
As seen in Fig. 7, the input power for compressor saw an increase during the starting stage, which was believed to be caused by the increased discharge pressure. During the steady operating period, the averaged value of input power for compressor was at 845.7 W. Furthermore, as mentioned in the discussions of Fig. 5, the refrigerant side heating capacity was larger than the actual value during the initial 8 min into the heating period. Therefore, the COP of the experimental system during this period was evaluated by: 17
COP
Qr Qnc Q fc Qs P Wo, fan Wi , fan
(9)
As for the following heating period, from 8 to 60 min, the COP for the experimental system was evaluated by Equation (8). It can be seen from Fig. 7, the variation trend of the COP was similar to that of input power for compressor, and the averaged value of COP during steady operating stage was at 3.11.
3.2. Impacts of outdoor air temperature on the performances of the experimental system
The impacts of outdoor air temperatures on the performances of the experimental system were analyzed in terms of characteristic temperatures and heating capacities for the heating terminal, suction and discharge pressure, and input power for the compressor, and COP for the experimental system. Particularly, all the parameters were the averaged values during the steady operating stages in Tests 2-9. In Fig. 8, the characteristics temperatures for the heating terminal under different outdoor air temperatures were compared. As seen, all of the inlet and outlet refrigerant, refrigerant condensing, radiant panel surface and outlet air temperatures saw a linear increase when the outdoor air temperature increased from -4.0 to 10.0 oC, provided that the inlet air temperature remained at around 20.0 oC. Besides, the rising rates of 18
the inlet and outlet refrigerant, refrigerant condensing, radiant panel surface and outlet air temperatures were 0.63, 0.55, 0.51, 0.41 and 0.28 oC per increased outdoor air temperature, respectively.
Fig. 9 shows the impacts of outdoor air temperature on the characteristic heating capacities for the heating terminal. It can be seen that the variation trends of the characteristic heating capacities were similar to those of the characteristic temperatures. The rising rates of the refrigerant side, radiant, natural convective and forced convective heating capacities were 68.9, 10.8, 9.7, and 58.0 W per increased outdoor air temperature, respectively. In addition, it was also noticeable that the ratios of the radiant and forced convective heating capacities remained stable at around 0.21 in Tests 2-9, which meant that the variation of the outdoor air temperatures had no effects on the ratio of the radiant to forced convective heating capacities. On the other hand, the ratios of the radiant and natural convective heating capacities decreased from 1.48 to 1.35 when the outdoor air temperature increased from -4.0 to 10.0 oC. Generally, the ratios of the averaged radiant, natural and forced convective heating capacities were around 1.4:1.0:6.9 in Tests 2-9.
As shown in Fig. 10, the suction and discharge pressures increased with an increase in outdoor air temperature, and their variation trends were nearly linear. The 19
rising rate of the former was 0.02 MPa per increased outdoor air temperature, and that of the latter 0.04 MPa per increased outdoor air temperature. Fig. 11 shows the input power for the compressor and COP for the experimental system in Tests 2-9. As for the input power, its averaged value during steady operating stage was at 714.6 W when the outdoor air temperature was -4.0 oC, and increased to 877.7 W when the outdoor air temperature was 10.0 oC. The COP for the experimental system also increased with an increased outdoor air temperature with a rising rate of 0.04, and the value of the COP was at 3.22 when the outdoor air temperature was 10.0 oC. Impacts of indoor air flow rate on the performances of the experimental system
Similar to the analytical method proposed in Section 3.2, the impacts of indoor air flow rate on the performances of the experimental system were analyzed and presented in this section. Fig. 12 shows the impacts of indoor air flow rate on the characteristic temperatures for the heating terminal. As seen, all the characteristic temperatures except inlet air temperature (maintained at around 20 oC during experiments) decreased with an increase in indoor air flow rate, which was believed to be caused by the decreased discharge pressure for the compressor. The values of inlet and outlet refrigerant, refrigerant condensing, radiant panel surface and outlet air temperatures were at 81.0, 47.9, 50.9, 47.8 and 38.9 oC when the indoor air flow was 288 m3/h, and those decreased to 71.1, 41.5, 44.5, 40.9 and 32.1 oC when the indoor air flow was 551 m3/h. 20
In Fig. 13, the characteristic heating capacities for the heating terminal under different indoor air flow rate were compared. As seen, the radiant and natural convective heating capacities decreased with an increased indoor air flow rate. With respect to the indoor air flow rate 288, 456 and 551 m3/h, the values of the former were at 652.7, 558.5 and 500.2 W, and those of the latter at 557.5, 439.5 and 363.1 W. By contrast, the refrigerant side heating capacity increased from 2919.4 to 3016.8 W and the forced convective capacity increased from 1966.7 to 2409.4 W when the indoor air flow rate increased from 288 to 551 m3/h. Furthermore, it was noticeable that the ratio of the natural convective to radiant heating capacities decreased with the increased indoor air flow rate (0.85 at 288 m3/h and 0.73 at 551 m3/h). The same variation trend was observed on the ratio of the radiant to forced convective heating capacities (0.33 at 288 m3/h and 0.21 at 551 m3/h). On the other hand, the ratio of the forced convective to refrigerant side heating capacities increased with the increased indoor air flow rate (0.67 at 288 m3/h and 0.80 at 551 m3/h).
Fig. 14 shows the suction and discharge pressures for the compressor under different indoor air flow rates. It can be seen that the impact of the indoor air flow rate on the suction pressure was insignificant, and the value of suction pressure remained at around 0.72 MPa. By contrast, the discharge pressure decreased from 3.25 to 2.82 21
MPa when the indoor air flow rate increased from 288 to 551 m3/h. Fig. 15 shows the impacts of indoor air flow rate on the input power for the compressor and COP for the experimental system. As seen, when the indoor air flow rated increased from 288 to 551 m3/h, the input power decreased from 948.4 to 845.7 W and the COP increased from 2.86 to 3.11. The above analysis suggested that the indoor air flow rate had significant impacts on the performances of the experimental system. Particularly, adjusting indoor air flow rate could effectively allocate the amount of heat generated through different heat transfer modes, which may have significant effects on the indoor thermal environment.
3. Conclusions To combine the advantages of conventional finned tube heat exchanger based on forced convective heat transfer mode and those of direct radiant heating terminal, an experimental study on a novel radiant-convective heating system based on ASHP was carried out in this paper. Both the operating characteristics and heating performances of the novel system under standard heating condition were experimentally investigated. Furthermore, the impacts of outdoor air temperature and indoor air flow rate on the performances of the novel system, such as characteristic temperatures and heating capacities, were also experimentally investigated. The following conclusions were made: 22
1. The novel system took about 28 min to enter a steady operating stage, which means that it had a longer starting stage than that of a conventional ASHP system due to novel radiant-convective heating terminal. 2. During steady operating stage under the standard heating condition, the radiant panel surface temperature and outlet air temperature for the novel heating terminal were 40.9 and 32.1 oC, respectively. In addition, the radiant, natural convective, forced convective heating capacities were 500.2, 363.1 and 2409.5 W, respectively. The COP of the novel system was 3.11. 3. The radiant panel surface and outlet air temperatures for the heating terminal, and radiant, natural convective and forced convective heating capacities, and COP of the novel system saw a linear increase when the outdoor air temperature increased from -4.0 to 10.0 oC. Their respective rising rates were 0.41 oC, 0.28 oC 10.8 W, 9.7 W, 58.0 W and 0.04 per increased outdoor air temperature. 4. The indoor air flow rate had significant impacts on the performances of the novel system. In particular, adjusting the indoor air flow rate could effectively allocate the amount of heat generated through different heat transfer modes, which may have significant effects on the indoor thermal environment. Generally, the feasibility and stability of the novel heating system have been verified in this work. The investigation on optimizing of the control strategy and components for the novel system, as well as its indoor thermal environment will be carried out in the future work. 23
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Outdoor-environmental chamber
Indoor-environmental chamber
Experimental novel radiant-convective heating system Radiant-convective heating terminal Four-way valve
Electric heater
Outdoor coil Compressor Gas-liquid separator
Humidifier
Accumulator
Thermostatic expansion valve Filter
Cooler
Fig. 1. The schematic of the experimental setup.
28
Air inlet Copper fin Copper pipe Air fan Copper fin pitch Insulating layer
Copper tube spacing
Fin width Radiant panel
Fin length
Refrigerant inlet Refrigerant outlet
Air outlet
Fig. 2. The schematic of the radiant-convective heating terminal.
P T G
Four-way valve
T
T P
Outdoor coil
T
T
T
P T
T H
Radiant-convective heating terminal
T P
Compressor T H
T V
T
Gas-liquid separator
T
T
T T P
Accumulator
Thermostatic expansion valve Filter
P T
T
P
H
V
Temperature
Pressure
Humidity
Velocity
G Mass flow rate
Fig. 3. The schematics of the measuring points installed in the experimental system.
29
75 Indoor air fan was turned on
70
Inlet refrigerant Outlet refrigerant Refrigerant condensing Radiant panel surface Inlet air Outlet air
65 60
Temperature (oC)
55 50 Starting stage 45
Steady operating stage
40 35 30 25 20 15 0
5
10
15
20
25
30
35
40
45
50
55
60
Times (min)
Fig. 4. Time-variations of characteristic temperatures for the heating terminal in Test 1. 3000
2500 Refrigerant redistribution
Capacity (W)
2000
Starting stage
Steady operating stage
1500
Refrigerant side heating Radiant heating Natural convective heating Forced convective heating Energy storage
1000
500
0 0
5
10
15
20
25
30
35
40
45
50
55
60
Times (min)
Fig. 5. Time-variations of characteristic heating capacities for the heating terminal in Test 1.
30
3.0
2.5
Pressure (MPa)
Starting stage
Steady operating stage
2.0
Suction pressure Discharge pressure
1.5
Indoor air fan was turned on
1.0
0.5 0
5
10
15
20
25
30
35
40
45
50
55
60
Times (min)
Fig. 6. Time-variations of suction and discharge pressure for the compressor in Test 1. 900
5
750 Starting stage
4
Steady operating stage
3
COP
Input Power (W)
600
450 2
Input power for compressor COP for experimental system
Evaluated by Equa. (8) 300 Evaluated by Equa. (9)
150
1
0
0 0
5
10
15
20
25
30
35
40
45
50
55
60
Times (min)
Fig. 7. Time-variations of input power for compressor and COP for the experimental system in Test 1.
31
75 70 65
Inlet refrigerant Outlet refrigerant Refrigerant condensing Radiant panel surface Inlet air Outlet air
60
Temperature (oC)
55 50 45 40 35 30 25 20 15 -4 (Test 2)
-2 (Test 3)
0 (Test 4)
2 (Test 5)
4 (Test 6)
6 (Test 7)
8 (Test8)
10 (Test 9)
Outdoor air temperature (oC)
Fig. 8. The characteristic temperatures for the heating terminal in Tests 2-9.
3000
Heating capacity (W)
2500
2000 Refrigerant side Radiant Natural convective Forced convective
1500
1000
500
0 -4 (Test 2)
-2 (Test 3)
0 (Test 4)
2 (Test 5)
4 (Test 6)
6 (Test 7)
8 (Test8)
10 (Test 9)
Outdoor air temperature (oC)
Fig. 9. The characteristic heating capacities for the heating terminal in Tests 2-9.
32
3.0
2.7
Pressure (MPa)
2.4 Scution pressure Discharge pressure
2.1
0.9
0.6
0.3 -4 (Test 2)
-2 (Test 3)
0 2 4 6 (Test 4) (Test 5) (Test 6) (Test 7) Outdoor air temperature (℃ )
8 (Test8)
10 (Test 9)
Fig. 10. The suction and discharge pressures for the compressor in Tests 2-9.
900
4.0
850 3.5
750
3.0
COP
Input Power (W)
800
700 Input power for compressor COP for experimental system
2.5
650
600
2.0 -4 (Test 2)
-2 (Test 3)
0 (Test 4)
2 (Test 5)
4 (Test 6)
6 (Test 7)
8 (Test8)
10 (Test 9)
Outdoor air temperature (oC)
Fig. 11. The input power for the compressor and COP for the experimental system in Tests 2-9.
33
100 90
Inlet refrigerant
Outlet refrigerant
Outlet refrigerant
Radiant panel surface
Inlet air
Outlet air
80
Temperature (oC)
70 60 50 40 30 20 10 0 288 (Test 11)
551 (Test 1)
456 (Test 10) 3
Indoor air flow rate (m /h)
Fig. 12. The characteristic temperatures for the heating terminal in Tests 1, 10-11. 3500
Refrigerant side
Radiant
Natural convective
Forced convective
3000
Heating capacity (W)
2500
2000
1500
1000
500
0 288 (Test 10)
551 (Test 1)
456 (Test 11) 3
Indoor air flow rate (m /h)
Fig. 13. The characteristic heating capacities for the heating terminal in Tests 1, 10-11.
34
Suction pressure
3.5
Discharge pressure
3.0
Pressure (MPa)
2.5
2.0
1.5
1.0
0.5
0.0 288 (Test 10)
551 (Test 1)
456 (Test 11) 3
Indoor air flow rate (m /h)
Fig. 14. The suction and discharge pressures for the compressor in Tests 1, 10-11.
1200
5 Input power for compressor
COP for experimental system
1000
4
3 600
COP
Input power (W)
800
2 400
1
200
0
0 288 (Test 11)
456 (Test 10)
551 (Test 1)
Indoor air flow rate (m3/h)
Fig. 15. The input power for the compressor and COP for the experimental system in Tests 1, 10-11.
35
Table 1 Specifications of the novel system. Items
Parameters
Values
Units
Power type
1-220-50
φ-V-Hz
Input power
885
W
Dimension
700×535×235
mm(L×H×W)
Air flow rate
1750
m3/h
Accumulator
Volume
0.6
L
Thermostatic expansion valve
Type
EMERSON-B
Filter
Type
C-083-S
Refrigerant
Type
R410A
Air flow rate
0-551
m3/h
Copper tube spacing
70
mm
9.52
mm
Copper pipe thickness
0.7
mm
Number of circuits
3
Copper fin pitch
40
mm
Copper fin width
50
mm
Copper fin length
2000
mm
Copper fin thickness
0.3
mm
2000×1600×0.3
mm(L×W×δ)
Compressor
Outdoor coil
Copper pipe external diameter
Novel radiant-convective heating terminal
Radiant panel dimension Radiant panel material
36
Copper
Table 2 Experimental conditions for Tests 1 to 11.
Test
Outdoor air temperature
Outdoor air relative humidity
Inlet air temperature for
Air flow rate for the
the heating terminal
heating terminal
1
7.0±0.3 oC
86±3 %
20.0±0.3 oC
551 m3/h
2
-4.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
3
-2.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
4
0.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
5
2.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
6
4.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
7
6.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
8
8.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
9
10.0±0.3 oC
60±3 %
20.0±0.3 oC
551 m3/h
10
7.0±0.3 oC
86±3 %
20.0±0.3 oC
456 m3/h
11
7.0±0.3 oC
86±3 %
20.0±0.3 oC
37
288 3/h