Applied Thermal Engineering 27 (2007) 430–441 www.elsevier.com/locate/apthermeng
A novel rainwater–ground source heat pump – Measurement and simulation Guohui Gan *, Saffa B. Riffat, C.S.A. Chong Institute of Building Technology, School of the Built Environment, University of Nottingham, University Park, Nottingham NG7 2RD, United Kingdom Received 20 July 2005; accepted 10 July 2006 Available online 18 September 2006
Abstract This paper presents the results of experimental measurement and numerical simulation of the performance of a heat pump system designed to make use of rainwater and ground as heat sources/sinks. The system was tested under laboratory conditions. A refrigerant was circulated through a closed loop heat exchanger to transfer heat between the heat pump and rainwater in a storage tank and another heat exchanger made of solid bars or heat pipes to transfer heat between the stored rainwater and surrounding soil. Physical and thermal properties of soil such as water content, density, specific heat, thermal diffusivity and thermal conductivity were determined. Numerical simulations were also carried out for a rainwater storage tank installed under ground for domestic application of the heat pump with different operating modes, heating loads and the sizes and types of heat exchanger. 2006 Elsevier Ltd. All rights reserved. Keywords: Ground-source heat pump; Heat exchanger; Rainwater storage; Computational fluid dynamics
1. Introduction Buildings account for approximately 40% of the total energy consumption in the developed countries. With the increasing concern over global warming due to CO2 emissions from the use of fossil fuels, efforts are being made to develop energy efficient and environmentally friendly systems. Ground-source heat pumps (GSHPs), which make use of earth or ground water as the heat source or sink often through borehole heat exchangers, are suitable for heating and cooling of buildings and so could play a significant role in reducing CO2 emissions. The main benefit of using GSHPs is that the temperature of the subsurface is not subject to large variations experienced by air, currently the most common thermal energy source for heat pumps, and so would allow construction of more efficient systems with superior performance. GSHPs do not need large cool-
*
Corresponding author. Tel.: +44 115 9514876. E-mail address:
[email protected] (G. Gan).
1359-4311/$ - see front matter 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2006.07.011
ing towers and their running costs are lower than conventional heating and air conditioning systems. As a result, GSHPs have increasingly been used for building heating and cooling with an annual rate of increase of 10% in recent years. By 2003, there were about 800,000 GSHP installations worldwide [1]. Most of the installations are found in the USA and Europe. In the UK, there are only a few hundred installations [2]. One of the barriers to the wider use of GSHPs is the cost of drilling deep boreholes for the installation of ground heat exchanger loops. However, this obstacle could be overcome by making use of a rainwater collection system that employs underground storage tanks. The maximum excavation depth required for installing these tanks is only 2–3 m. This could be accomplished using ordinary excavators rather than specialised drilling machines required for deep boreholes. Collection of rainwater for use in buildings has been practised since ancient times. Rainwater has a number of advantages over the mains water supply. For example, it is soft water and so eliminates the problems associated with hard water. Collection of rainwater also reduces loading on
G. Gan et al. / Applied Thermal Engineering 27 (2007) 430–441
watercourses during rainfall and encourages conservative use of water. In a typical modern application, rainwater collected from the roof of a building is filtered and then stored in storage tanks. Water stored in this way could be a suitable energy source for heat pumps because its temperature is more stable than air temperature. Measurements of the rainwater collection system installed at the Eco-house, University of Nottingham by the authors in March 2003 showed that the temperature of stored rainwater at an average level of 1.8 m below ground remained nearly constant at 6.8 C, which was very close to the soil temperature (7 C) at the same level, when the outdoor air temperature varied from 0 C to 16 C. The dual use of rainwater as a supplementary water supply and as a heat transfer medium would enhance the economics of GSHPs. Over the last 20 years, numerical and experimental investigations into GSHPs have been mainly focussed on vertical borehole heat exchangers [3–11] and, to a lesser extent, horizontal trench heat exchangers [12–14]. In recent years, innovative techniques such as water tank coils [15,16] have been used to enhance the performance of horizontal heat exchangers and an analytical model has been developed for a hemispherical surface water tank as the ground heat source/sink for a heat pump system [17,18]. Research into the use of rainwater as a heat transfer medium for GSHPs has not been carried out. This paper describes an experimental and computational investigation into a GSHP system that utilises rainwater as a heat source/sink by employing a heat exchanger integrated into a water storage tank and surrounding soil. Deployment of the system would reduce the need for heating/cooling of buildings as well as mains water supply and so reduce energy costs and CO2 emissions.
431
2. Laboratory testing of a rainwater–GSHP system A rainwater–GSHP system was designed and constructed for laboratory testing under controlled conditions. Fig. 1 shows the schematic diagram of the test rig. A thermostatically controlled water heater (1.4 kW) supplied hot water which was circulated between the hot water supply tank and cylindrical hot water storage tank (1575 mm in diameter and 1190 mm deep, made of LMDPE) using a pump to keep the surface temperature of the cylindrical soil storage tank (1300 mm in diameter and 1200 mm deep, made of galvanised steel) at a desired level. A heat exchanger (heat exchanger A) was used to transfer heat from soil in the soil tank to the cold water in the cylindrical cold water tank (570 mm in diameter and 780 mm deep, made of MDPE) from which heat was further transferred to the cooling device (heat pump or chiller) through another heat exchanger (heat exchanger B, a double spiral copper coil of 12.7 mm in diameter and with a total heat transfer area of 1.02 m2) using a mixture of 40% monoethylene glycol and 60% water. The chiller was able to supply the fluid circulating through heat exchanger B at controlled temperatures between 4 C and 13 C. The heat pump was originally connected to the looped ground source using a working fluid (refrigerant) to provide heating/cooling of the laboratory building. It was controllable for power output and supply air temperature but not adjustable for the temperature of the working fluid and so when it was used as the heat sink the temperature in the cold water tank would drop below the freezing point. This was useful to determine, e.g., how long the system could be operated continuously but it was not feasible to obtain data for a long period of operation.
Fig. 1. Schematic diagram of the test rig.
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Heat exchanger A was one of the key components of the system investigated. The following two types of heat exchanger A were designed and tested: • Solid bar heat exchanger. It consisted of 24 cylindrical aluminium bars in three rows. Each row had eight bars of 63.5 mm in diameter and 400 mm long arranged radially. Seven aluminium fins of 120 mm long, 200 mm high and 2 mm thick were used on the water side and four fins of 200 mm long, 200 mm high and 2 mm thick on the soil side of each solid bar. • Heat pipe heat exchanger. It consisted of two rows of four radially arranged copper heat pipes, each 38.1 mm in diameter and 400 mm long. Four aluminium fins of 120 mm long, 200 mm high and 4 mm thick were used on the water side. For the soil side, four fins of different sizes were used, the height and thickness of which were the same as those on the water side but the length varied from 400 mm near the cold water tank to 760 mm
near the hot water storage tank. The fins for heat pipes were thicker than those for solid bars to a degree to compensate for low efficiency of long fins. The heat pipes were designed and manufactured by Thermacore Europe Ltd. [19] based on the requirements for the laboratory testing. Methanol was used as the working fluid for the heat pipes. Because the total surface area of fins for part of the (heat pipe) heat exchanger in the cold water tank was much smaller than that in the soil due to limited space, a small pump was used to circulate the cold water within the tank to increase the heat transfer between the fins and cold water for all the tests. This also helped to maintain a more uniform water temperature in the cold water tank. Fig. 2 shows the two assembled heat exchangers. Soil temperatures at three vertical sections and 11 horizontal points along the heat flow path from hot to cold water storage tanks were measured using thermocouples
Fig. 2. Two types of heat exchanger in soil and water tanks. (a) Solid bar heat exchanger and (b) heat pipe heat exchanger.
G. Gan et al. / Applied Thermal Engineering 27 (2007) 430–441
• Soil without heat exchanger A. • Soil with solid bar heat exchanger. • Soil with heat pipe heat exchanger. For each set of tests, both the heat pump and chiller were used as the heat sink. Fig. 3 shows an example of measured soil temperatures at selected points along the heat flow path from the hot water storage tank to the cold water tank (T1 for soil near hot water storage tank to T6 for soil near cold water tank) and mean value (Tmean), with the solid bar heat exchanger and using the chiller as the heat sink. The mean soil temperature was taken to be the mass weighted average of the measured soil temperatures for all the points along the radius. The soil temperature (T) was normalised by the hot (Th) and cold (Tc) water temperatures. The soil temperature ratios for some points were not equal or close
T1
T2
T3
T4
T5
T6
Hot water tank
to one at the start of the test because the cold water was at a similar temperature to the soil temperature which was not exactly uniform for the whole tank. Unlike traditional transient heating/cooling tests such as the hot plate method for determining the thermal conductivity of materials where the heat source/sink was controlled at a constant rate or temperature, the cold water for this investigation was part of the energy source. Therefore, the cold water temperature, like the soil temperature, decreased continuously during the entire test period as shown in Fig. 4(a). The decrease in the temperature was most pronounced during the first hour of the test by then the temperature had been reduced by 50% from its initial to final temperatures. When the soil temperature was normalised by the temperature of the cooling fluid in heat exchanger B, which was nearly constant from the chiller, instead of the cold water temperature, the temperature ratio (denoted by Tmean_coil in Fig. 3) decreased continuously from approximately one at the beginning. The cooling power input from the chiller also decreased during the test (see Fig. 4(b)) as heat stored in the water and soil was removed. The decrease roughly followed the pattern for the cold water temperature. Without a heat exchanger in the soil, heat transfer from hot water through soil to cold water was found to be slow due to the low soil thermal conductivity and due to the limited contact area between soil and the cold water tank. For example, the normalised temperature drop was approximately 0.17 in 6 h of operation. Deployment of the solid bar heat exchanger increased the heating/cooling rate and capacity of the system considerably – the temperature drop
1.0
(T-Tfinal)/(Tinitial-Tfinal)
(type K). Thermocouples were also used to measure temperatures of hot and cold water and the fluid mixture in heat exchanger B. The flow rate of the fluid in heat exchanger B was measured using a rotameter. The cooling power provided by the heat pump or chiller was calculated from the measured inlet and outlet temperatures and the flow rate together with the specific heat of the fluid in heat exchanger B. The density and water content of soil were measured using standard methods [20]. The specific heat of soil was then calculated from the solid and water components of the soil [21]. The measured density and water content for one batch of soil were 1771 kg/m3 and of 17.4%, respectively. The calculated specific heat of soil was 1334 J/kg K. The thermal diffusivity and conductivity of soil were determined using the numerical method together with experimental measurement of the temperatures as described in the next section. The following three sets of tests were carried out under different hot and cold water temperatures:
433
0.8 0.6 0.4 0.2
Cold water tank
0.0 Soil tank, Tmean
0
4
8
0
4
8
Soil tank, Tmean_coil
1.0
Tmean_coil
Chiller cooling power (W)
(T-Tc)/(Th-Tc)
T3
T2
0.6 Tmean 0.4
T4 T5
0.2
T6
0.0 0
4
8
12 Time (hr)
16
16
20
24
12
16
20
24
5000
T1 0.8
12 Time (hr)
4000 3000 2000 1000 0
20
24
Fig. 3. Measured soil temperatures with the solid bar heat exchanger and using the chiller as the heat sink.
Time (hr)
Fig. 4. Variations of cold water temperature and power input from chiller. (a) Cold water temperature and (b) power input from chiller.
G. Gan et al. / Applied Thermal Engineering 27 (2007) 430–441
increased to 0.31 for the same cooling period and the heat transfer rate increased by over 80%. Heat pipes are generally superior to solid bars for heat transfer because of the latent heat transfer in the partially evacuated pipes. However, it was found that the heat pipe heat exchanger was not as effective as the solid bar heat exchanger for this application. This was due to the much lower efficiency of fins for the heat pipes (varied from 25% to 54%) than that for solid bars (86%) because much larger fins were used for 2 · 4 heat pipes than those for 3 · 8 solid bars in order to compare the performance under the same total fin surface area and for the same amount of heat transfer through the heat pipes and solid bars. If the same number and size were used for heat pipes and solid bars, i.e., same fin efficiency and heat transfer area, the heat pipes would be expected to be more effective.
30
Soil temperature (ºC)
434
25
20
Measured Diffusivity = 0.81 mm2/s Diffusivity = 0.45 mm2/s
15
10 1
2
3
4
5 6 7 Measuring point
8
9
10
11
Fig. 5. Measured and predicted (with specified diffusivity) soil temperatures without heat exchanger A at 6th hour.
3.2. Simulation of the rainwater storage tank and heat exchanger
3. Computer modelling of rainwater–GSHP systems A finite difference method was used to solve a set of conduction and convection equations for the heat transfer between the hot water, cold water, heat exchanger (A) and surrounding soil as illustrated in Fig. 1. In addition, a commercial computational fluid dynamics (CFD) software FLUENT [22] was used to predict the performance of the GSHP system involving fluid flow in the water tank, conduction heat transfer in soil and convective heat transfer between the interface of soil and water tank as well as soil and atmosphere. Modelling was carried out both under laboratory test conditions to determine the thermal properties of soil and under simulated rainwater tank installations for domestic use. 3.1. Determination of soil thermal diffusivity and conductivity The determination of the thermal diffusivity and thermal conductivity of soil was based on the predicted and measured soil temperature histories and physical properties. Predictions were performed for the transient temperature distribution without heat exchanger A in the soil tank using the measured surface temperatures as boundary conditions. The thermal diffusivity of soil was then determined through fitting the predicted temperatures to the measured temperatures, in terms of either histories for specific points (e.g., temperature histories for the centreline between hot and cold water storage tanks or mass average) or variations along the heat flow path from the hot water storage tank to the cold water tank at specific times. An example of the latter is shown in Fig. 5. The ‘measured’ thermal diffusivity was 0.45 mm2/s for soil with a density of 1771 kg/m3 and water content of 17.4%. Fig. 5 also shows that use of a higher value (0.81 mm2/s) for the thermal diffusivity based on data from CIBSE [23] would over-predict the rate of heat transfer. The soil thermal conductivity was given by the product of the thermal diffusivity, density and specific heat.
Simulations were carried out for a cylindrical rainwater storage tank installed one meter below the ground surface for different heat extraction rates and operations by the heat pump, different materials and configurations of the heat exchanger between the rainwater tank and surrounding earth. For example, assuming that a 2-storey house of 10 m long · 6 m wide · 3 m ceiling height required heating fresh air at one air change rate by a 15 K temperature increase, using a heat pump with a coefficient of performance of three would require a heat extraction rate from the ground of 1.2 kW. This is approximately the same as for a temperature drop of 5 K of rainwater stored in a 5 m3 tank in one day. In the examples illustrated below, FLUENT [22] was used to predict the soil and water temperatures for the cylindrical (i.e., axisymmetric) computational domain. The heat exchanger in the water tank (heat exchanger B) was simplified as seven uniformly spaced flat circular plates. It was assumed that the initial soil temperature was 10 C. The ambient air was assumed at 5 C temperature and 3 m/s wind speed. To model the heat transfer between the water tank and soil, heat exchanger A was also simplified as flat plates for both solid and heat pipe heat exchangers. 3.2.1. Solid heat exchangers The solid heat exchanger was modelled as seven large annular plates (made of copper or aluminium). Each plate had an inner radius and an outer radius of 1 m and 10 m, respectively, for modelling the heat exchanger in soil of 9 m long. The predicted temperature of stored water varied with space and time. For heating operation, water in the vicinity of heat exchanger B would have the lowest temperature. Fig. 6 shows the variations of the lowest water temperature with time for continuous and intermittent operations. Fig. 7 shows two examples of predicted temperatures of soil and stored water with and without a solid heat exchanger to transfer heat between the two media for continuous operation. The performance of the heat exchanger was assessed from the duration that the heat
G. Gan et al. / Applied Thermal Engineering 27 (2007) 430–441
Table 1 Predicted duration of operating the GSHP with and without solid heat exchangers
10
No HX
Al HX
Cu HX
Temperature ( º C)
8 6
Heat extraction (kW)
Operation mode
Duration (days) for different heat exchangers No HX
Aluminium HX
Copper HX
1.2
Continuous 12-h on, 12-h off 8-h on, 16-h off
2 34 7 10 34
5 26 66
20 23 99 >99
2.4
Continuous 12-h on, 12-h off 8-h on, 16-h off
21 h 2 23 4 12
25 h 4 11
55 h 16 46
4 2 0 0
3
6
9 12 Time (Day)
15
18
435
21
10
No HX
Al HX
Cu HX
Temperature ( ºC)
8 6 4 2 0 0
20
40 60 Time (Day)
80
100
Fig. 6. Predicted lowest water temperature using a solid heat exchanger (heat extraction rate = 1.2 kW). (a) Continuous operation and (b) intermittent (12-h on, 12-h off) operation.
pump could be operated for a given mode (continuous or intermittent) before the water in the storage tank would begin to freeze, i.e., the lowest water temperature dropped to 0 C in Fig. 6. The predicted duration of operating the heat pump is given in Table 1. As seen from Table 1 and Fig. 6(a), without heat exchanger A, operating the heat pump continuously would lead to the stored water at an initial temperature of 10 C to freeze in less than three days and completely frozen in three days (white colour in the water tank in Fig. 7 indicating temperatures below the freezing point, i.e., ice). When a horizontal solid heat exchanger was used to transfer heat from soil to the stored water, the heat pump could be operated continuously for five days with the aluminium heat exchanger or 20 23 days (20 days and 16 h) with the copper heat exchanger before
the water surrounding the evaporator coil of the heat pump began to freeze. For intermittent operation, the stored rainwater could be used for heat transfer for longer periods. For example, for 12-h-on (heating) and 12-h-off operation as shown in Fig. 6(b) (where the variation in the lowest water temperature for intermittent operation is shown only for the first week and the lowest temperature during heating is shown for the remaining period), the system without heat exchanger A could be run for seven days intermittently compared with less than three days for continuous operation before the water in the tank would freeze. The operating period for the system integrated with an aluminium heat exchanger could be extended from five days continuously to 26 days intermittently. Using a copper heat exchanger, the system could be operated intermittently for practically an unlimited period under the same constant ambient conditions. Figs. 6(b) and 8 show that the lowest water temperature would still be about 2 C above the freezing point after 60-day intermittent operation of the system with the copper heat exchanger. This means that the heat exchanger would be capable of transferring heat from soil to water fast enough to provide the specified heating load. If heating is required for a shorter period, e.g., 8 h a day (i.e., 8-h-on and 16-h-off operation), the heat pump could be operated for an even longer period. However, in practice, the ambient conditions would change during such a long period. The system operating performance can still be predicted if variations in ambient air temperature and wind velocity are known.
Fig. 7. Predicted water and soil temperatures for continuous operation.
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G. Gan et al. / Applied Thermal Engineering 27 (2007) 430–441
Fig. 8. Predicted water and soil temperatures for intermittent operation.
Temperature ( º C)
10
No HX Al HX Cu HX
8 6
No HX Al HX Cu HX
4 2 0 0
1
2
3 Time (Day)
4
6
5
10
Temperature ( ºC)
Design and operation of the rainwater–GSHP system would also depend on heating/cooling loads and other factors such as the rainwater storage tank size and surrounding soil properties. When the heat extraction rate was doubled to 2.4 kW, it was predicted that the rainwater in the storage tank without heat exchanger A would begin to freeze within 21 h. Integrating the solid heat exchanger would not significantly increase the time for continuous operation. Using the aluminium or copper heat exchanger would allow operation of the system continuously only for 25 or 55 h, respectively, before water in the tank began to freeze. Essentially, this means that in order to meet a higher level of continuous heating demand the system with a preset lowest water temperature of 0 C would require a larger rainwater storage tank and/or faster heat transfer from/to the heat exchanger particularly on the section in the water tank by such means as efficient fins or water circulation within the tank using a small pump as used in the laboratory testing. The circulation or mixing of the stored water would enable the extension of the operating period of the heat pump because the predicted mean water temperature (i.e., average for the stored water) was higher than the lowest water temperature in the tank. For example, when the lowest water temperature reached the freezing point, the mean water temperature was between 2.8 C and 3.4 C for the heat extraction rate of 2.4 kW (Fig. 9a). If the stored water had a uniform temperature as a result of mixing, the mean water temperature could replace the lowest temperature for assessing the system performance. Based on the mean water temperature, the operating period of the heat pump could be extended from 21 to 32 h, from 25 to 53 h and from 55 to 130 h for the three cases without the heat exchanger in soil, with the aluminium and copper heat exchangers, respectively. For cases with longer permissible operating periods, the difference between the mean and lowest water temperatures was smaller. For example, for the system with the copper heat exchanger which could be operated at the heat extraction rate of 1.2 kW for 20 23 days before the stored water began to freeze, the difference was 1.5 C approximately and the mean water temperature would reach the freezing point after 30 days (Fig. 9b). In addition to the extension of the operating period resulting directly from the reduced spatial variation in water temper-
No HX Al HX
8
Cu HX No HX Al HX Cu HX
6 4 2 0 0
5
10
15
20
25
30
Time (Day)
Fig. 9. Predicted lowest and mean water temperatures. (a) Heat extraction rate = 2.4 kW and (b) heat extraction rate = 1.2 kW.
ature, water circulation would enhance the heat transfer between the stored water, heat exchangers in the tank and the tank surface in contact with water, thus indirectly increasing the available energy for extraction. GSHPs for building applications often involve intermittent operation rather than continuous operation and so the rainwater–GSHP system integrated with heat exchanger A could in practice cope with higher heating loads, though for shorter operating periods. For the 12-h-on and 12-hoff intermittent operation, e.g., using the aluminium heat exchanger would enable the system to extract 2.4 kW heating power for four days and the operating period could be increased further to 16 days using the copper heat exchanger. If heating is required only for 8 h a day, i.e., 8-h-on and 16-h-off operation, the system could sustain heat extraction at 2.4 kW for 4.5, 11 and 46 days without the
G. Gan et al. / Applied Thermal Engineering 27 (2007) 430–441 10
9m Temperature ( º C)
8
4m 3m 2m 1m
6 4 2 0 0
3
6
9 12 Time (Day)
15
18
21
10
1m
2m
3m
4m
Difference (%)
8 6 4 2 0 0
3
6
9 12 Time (Day)
15
18
21
Fig. 10. Effect of the length of copper heat exchanger on the lowest water temperature (heat extraction rate = 1.2 kW). (a) Lowest water temperature and (b) relative difference from the original 9 m long HX.
heat exchanger in soil, with the use of aluminium and copper heat exchangers, respectively. The water temperature and duration of operating the heat pump would also be affected by the size of the heat exchanger. Fig. 10a shows the effect on the lowest water temperature of the length of the copper heat exchanger in soil (length = difference between outer and inner radii) at a heat extraction rate of 1.2 kW. The predicted temperature was more or less the same for all the lengths during the early stage (about three days) and very close for lengths of 3, 4 and 9 m for the whole period presented in the figure (21 days). When the length of the heat exchanger was reduced from original 9 m to 1, 2, 3 and 4 m, the duration that the heat pump could be operated before the stored water began to freeze was reduced from 20 23 days to 10.5, 15.5, 18.5 and 20 days, respectively. To elucidate the difference in the water temperature profiles using the heat exchanger of different lengths clearly, a relative temperature difference for a shorter than 9-m long heat exchanger is defined as the ratio of the difference between the lowest water temperature using the 9-m long heat exchanger and
One heat-pipe plate
Fin
Two heat-pipe plates
Fin
437
that using a shorter heat exchanger to the temperature drop from the initial value (i.e., 10 C) using a shorter heat exchanger. Fig. 10b shows that the relative temperature difference was negligible in the early period of operation. The difference from the original 9 m long heat exchanger was still within 1% after operating the heat pump for 2.5, 5.5, 12 and 18.5 days with heat exchangers of 1, 2, 3 and 4 m long, respectively. Therefore, the heat exchanger could practically be simulated with a length of 4 m or even 3 m (in the soil section) without much loss of accuracy. Put it another way, for continuous operation of the heat pump for a period of two to three weeks at a heat extraction rate of 1.2 kW, a heat exchanger of 4–5 m radius in soil would be sufficient. 3.2.2. Heat pipes In order to compare the performance of the heat pipe heat exchanger with solid heat exchangers, simulation was also carried out using the effective thermal conductivity for the heat pipes. The effective thermal conductivity of a heat pipe can be calculated from the thermal resistances [24,25]. However, because the heat exchanger was modelled as flat plates and the internal structure of the heat pipes was not involved, the effective thermal conductivity (ke) could not be determined but was assumed to be 100 times the thermal conductivity of copper (kp), i.e., ke = 100kp. A much higher value (ke = 500kp) was also used later to investigate if the external resistance could be the limiting factor for heat transfer. Thyrum and Cruse [26] varied the effective thermal conductivity of heat pipes from 10,000 to 100,000 W/mK (about 26–260 times copper thermal conductivity) for modelling heat-pipe assisted heat sinks. Their results indicated that such large differences in the estimation of the heat-pipe thermal conductivity resulted in relatively small changes to the predicted temperature of the heat sink due to the external resistance between the fins and ambient. Simulation was performed for four arrangements of heat-pipe plates – 1, 2, 3 and 7 plates – as shown in Fig. 11. A vertical plate (annular in three dimensions) was in place to model the fins on heat-pipe plates for heat transfer between heat exchanger A, stored water and heat exchanger B when the number of plates for the two heat exchangers was different. Without this, heat pipes would have no effect on the performance of the system in terms of the operating period because heat transferred from heat pipes would not be able to flow to areas away from heat pipes in time to prevent the stored water close to heat exchanger B freezing up due to the low thermal conductivity
Three heat-pipe plates
HX A: Seven HX B heat-pipe plates
Fin
Fig. 11. Arrangements of heat-pipe plates for simulation.
G. Gan et al. / Applied Thermal Engineering 27 (2007) 430–441
of water. For real design or 3-D simulation, the heat pipes and fins in the water tank could be sandwiched between coils or other configurations representing heat exchanger B but this would not be possible for a 2-D axisymmetric model as fins would have cut through heat exchanger B. A smaller number of heat-pipe plates than seven was simulated because seven heat-pipe plates were found to be capable of transferring heat between the soil and stored water for an unlimited period for the heat extraction rates (1.2 kW and 2.4 kW) investigated. Even if the number of heat-pipe plates was reduced to three, the heat pump could still be run continuously without the risk of the stored water freezing. Therefore, discussion will be focussed on the results for one and two heat-pipe plates. The variations in the lowest water temperature using one and two heatpipe plates for a heat extraction rate of 1.2 kW are shown in Fig. 12. The ratio of two plates to the total number, i.e., 2/7, is comparable to the ratio of the heat pipes to solid bars for laboratory testing, 8/24. Tables 2 and 3 show the duration of operating the heat pump simulated using the effective thermal conductivity for the heat pipe heat exchanger. It is seen from Fig. 12 that using one heat-pipe plate would enable operation of the heat pump to extract heat at a rate of 1.2 kW continuously for 38.5 days before the
10
Temperature (º C)
one plate
two plates
8 6 4 2 0 0
5
10
15
20 25 Time (Day)
30
35
40
Fig. 12. Predicted lowest water temperature using a heat pipe heat exchanger (heat extraction rate = 1.2 kW).
Table 2 Predicted duration of operating the GSHP with one heat-pipe plate Heat extraction (kW)
1.2
2.4
Thermal conductivities of heat pipes and fins
Heat extraction (kW)
Thermal conductivity of fins
Duration (days)
1.2
kf = kp kp/9 kp/4.5
Unlimited 7 61
2.4
kp
6
stored water began to freeze and using two heat-pipe plates could prevent the stored water freezing. When the heat extraction rate was doubled to 2.4 kW, however, the operating period of the heat pump with one heat-pipe plate would be reduced drastically to less than two days (40 h) (see Fig. 13 and also Table 2 for ke = 100kp). The duration would not be much longer than that using one copper plate (ke = kp) of 25 h which coincidently was the same as that using seven aluminium plates (Table 1), suggesting that the external resistance between the heat pipes and soil had become the limiting factor for heat transfer. This was confirmed by a further simulation using a higher effective thermal conductivity (ke = 500kp). The higher effective thermal conductivity increased the operating period by two hours only, from 40 h to 42 h. Even when the number of heat-pipe plates was doubled, the heat pump could be operated for six days at the higher heat extraction rate (Table 3). To enable the operation of the heat pump for a longer period under the simulated conditions, more heat-pipe plates would be required. Similar to the solid heat exchanger, the water temperature and duration of operating the heat pump would be affected by the size of heat pipes. Fig. 14a shows the effect on the lowest water temperature of the length of heat pipes in soil at a heat extraction rate of 1.2 kW. When the length of heat pipes was reduced from original 9 m to 1, 2, 3 and 4 m, the duration that the heat pump could be operated before the stored water began to freeze was reduced from 38.5 days to 15, 21.5, 27 and 31.5 days, respectively. The relative difference in the temperature drop from the original 9 m long heat pipes was less than 1% after operating the
10
ke=500kp
Duration
ke
kf
100 kp
kp kp/36 kp/6 kp/3
38 12 days 3 14 days 4 14 days 5 12 days
kp
40 h 42 h 25 h
100kp 500kp kp
Table 3 Predicted duration of operating the GSHP with two heat-pipe plates (ke = 100kp)
ke – effective thermal conductivity of heat pipes; kf – thermal conductivity of fins; kp – thermal conductivity of copper.
Temperature ( º C)
438
ke=100kp
ke=kp
8 6 4 2 0 0
6
12
18
24 30 Time (Hour)
36
42
48
Fig. 13. Effect of the equivalent thermal conductivity of heat pipes on the predicted lowest water temperature (one heat-pipe plate, heat extraction rate = 2.4 kW).
G. Gan et al. / Applied Thermal Engineering 27 (2007) 430–441 10
9m 4m
8
Lowest temperature (ºC)
Lowest temperature (ºC)
10
3m 2m 1m
6 4 2 0
kf=kp kf=kp/36
8
kf=kp/6 kf=kp/3
6 4 2 0
0
5
10
10
15
2m
30
3m
35
40
0
5
10
15
20 25 Time (Day)
30
35
40
10
4m
Lowest temperature (ºC)
1m
20 25 Time (Day)
8 Difference (%)
439
6 4 2 0
kf=kp
8
kf =kp/9 kf =kp/4.5
6 4 2 0
0
5
10
15
20 25 Time (Day)
30
35
40
Fig. 14. Effect of the length of heat pipes on the lowest water temperature (heat extraction rate = 1.2 kW). (a) Lowest water temperature and (b) relative difference from the original 9 m long HX.
heat pump with 1, 2, 3 and 4 m long heat pipes for 5, 10, 17, 27 days, respectively (Fig. 14b). Because less than seven heat-pipe plates were used, the size of fins would in reality have been increased in order to transfer heat from the soil to stored water and the flat plates modelled as the heat exchanger inside the storage tank. As a result, the fin efficiency would have been reduced considerably. Assuming that plate fins were used on the heat-pipe plates, the fin efficiency can be estimated from the following equation when the heat transfer through the fin tip is negligible [27]: qffiffiffiffiffi tanh b k2h fd qffiffiffiffiffi g¼ b k2h fd where h is the convective heat transfer coefficient, kf is the fin thermal conductivity, d is the fin thickness and b is the fin height. The height of fins on one heat-pipe plate would have been increased by six times in order to cover the same area as provided by smaller fins for seven heat-pipe plates. To maintain the same efficiency of the large fins, the thickness or the thermal conductivity or the product of both would have to be increased by 36 times according to the above equation. Because the geometry of individual fins was not involved due to the excessive time requirement for transient heat transfer modelling, to simulate the effect of the reduced fin efficiency resulting from the increased fin height (same thickness), the thermal conductivity for the modelled
0
5
10
15 20 25 Time (Day)
30
35
40
Fig. 15. Effect of the fin height for heat pipes on the lowest water temperature (heat extraction rate = 1.2 kW). (a) One plate and (b) two plates.
fin (kf) (vertical plate in Fig. 11) was reduced by 36 times for one heat-pipe plate. Fig. 15a shows that the lowest water temperature decreased much faster with the reduced efficiency of large fins on one heat-pipe plate than that of the original small fins on seven plates. The operating period was reduced by over 10 times to 3.25 days from original 38.5 days (Table 2). Assuming that the large fins were six times as thick as the small fins so that the thermal conductivity of the large fins would only be reduced to 1/6 of the small fins to maintain the same fin efficiency, the operating period of the system would not be increased significantly (from 3.25 to 4.25 days). Even if the thickness was increased to 12 times of that for the small fins (kf = kp/3), the operating period was extended to only 5.5 days. For two heat-pipe plates, the fin height would have been tripled to cover the same area and the thermal conductivity was therefore assumed to be 1/9 of that for the small fins to maintain the same fin efficiency. The predicted operating period was no longer unlimited as for kf = kp but reduced to seven days (Fig. 15b and Table 3), less than the duration for the system with seven copper plates (20 23 days). This confirms the experimental finding that the heat pipe heat exchanger was less effective than the solid bar heat exchanger due to the lower fin efficiency. Assuming that the thickness of the large fins was doubled, as for fins on heat pipes for laboratory testing compared with those on solid bars, the thermal conductivity of the large fins would then be 2/9 of the small fins to maintain the same fin efficiency, the operating period of the system could be increased to
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61 days. This is longer than that for seven copper plates but in the laboratory testing the number of fins on heat pipes was less than that for solid bars due to the limited space available to accommodate the thicker fins. The problem would also occur in real fin design as using very thick fins, e.g., six times as thick assumed for a simulation for one heat-pipe plate, would present difficulty in accommodating the same number of fins, thus decreasing the overall hear transfer rate. 3.2.3. Further considerations The above predictions were performed for static water in the storage tank without taking into account supply of rainwater from the roof catchment or possibly top-up from mains water. If the flow of fresh water into the tank was taken into account, the system could be run for longer periods than predicted before the stored water began to freeze, depending on the patterns of rainfall and water consumption as well as the temperature of supply water (rain or mains water). The heat exchanger used to transfer heat between rainwater and soil could take different forms from heat conductors used in the simulations. For example, tank coils similar to those tested by Doherty et al. [16] could be employed instead. For this type of application, rainwater in the storage tank could be circulated directly through open coils or a refrigerant circulated in a closed loop between rainwater and surrounding soil as an alternative to the tank coil system where a refrigerant was circulated between a heat pump and tank coils [16]. Such a system could be designed and controlled to make use of stored rainwater as the first heat source/sink and to circulate water or refrigerant in the coils if and when more heat would be needed. It would also reduce the risk of water freezing in the storage tank if the coils could cover a sufficient large area of soil. Furthermore, the system would provide an opportunity for integration of a solar collector unit to form a combined solar–rainwater–GSHP. The combined system would allow solar energy to be used for hot water heating and/or stored in the rainwater tank using a closed loop to circulate a heat transfer fluid between the domestic hot water cylinder, rainwater storage tank and solar collector, thus increasing the overall coefficient of performance. It would also be possible to design the rainwater tank and solar collector for the heat pump system without the use of the underground heat exchanger between rainwater and soil. 4. Conclusions A unique facility has been developed to test the performance of individual components and the integrated system of a rainwater–GSHP involving a heat pump, water and soil and the storage tanks as well as heat exchangers under controlled conditions. Two types of heat exchanger – solid bars and heat pipes – have been designed and tested for the heat pump system to transfer heat between stored rainwater and soil. The solid
bar heat exchanger was found to be more effective due to the larger cross section and higher fin efficiency. CFD has been used to determine the thermal diffusivity/ conductivity of soil and predict the thermal performance of a rainwater–GSHP for a domestic application with different operating conditions. Results suggest that it is necessary to employ a heat exchanger to transfer heat between rainwater and surrounding soil in order to operate such a GSHP system at a design heating/cooing rate for a long period, unless a very large rainwater storage tank or other means such as solar energy is utilised to provide the required heat source/sink for the heat pump. Otherwise, the heating or cooling capacity of the system would decrease after a short period of operation if water temperatures have to be controlled within certain range e.g. between 1 C to prevent ice building up and 20 C to prevent rapid growth of bacteria such as the Legionella. The numerical method can be used for optimum design of rainwater storage tank and heat exchangers. The performance of the rainwater–GSHP in terms of the operating period before the temperature of the stored water reaches a limit depends on the operating mode, heating/cooling load and the size and type of heat exchanger. The longer the period of operation and/or the higher the load, the larger the heat exchanger would be required to transfer heat between soil and stored water. Use of heat pipes in place of solid heat exchangers can reduce the number/size of heat transfer elements but the heat transfer rate may be limited by the external resistances particularly for the section in the water tank where space is often limited. Acknowledgement The research project has been funded by the UK Engineering and Physical Sciences Research Council. References [1] J.W. Lund, B. Sanner, L. Rybach, R. Curtis, G. Hellstorm, Geothermal heat pumps – a world overview, Renewable Energy World 6 (4) (2003) 218–227. [2] G. Ellis, GSHP club seminar – ground source heat pumps from heat extraction to delivery, The National Energy Foundation, Milton Keynes, 12th May 2005. Available from: . [3] V.C. Mei, S.K. Fisher, Vertical concentric tube ground coupled heat exchangers, ASHRAE Transactions 89 (2) (1983) 391–406. [4] J.D. Deerman, S.P. Kavanaugh, Simulation of vertical U-tube ground-coupled heat pump systems using the cylindrical heat source solution, ASHRAE Transactions 97 (1) (1991) 287–295. [5] T.K. Lei, Development of a computational model for a groundcoupled heat exchanger, ASHRAE Transactions 99 (1) (1993) 149– 159. [6] N.K. Muraya, D.L. O’Neal, W.M. Heffington, Thermal interference of adjacent legs in a vertical U-tube heat exchanger for a groundcoupled heat pump, ASHRAE Transactions 102 (2) (1996) 12–21. [7] C. Yavuzturk, J.D. Spitler, S.J. Rees, Transient two-dimensional finite volume model for the simulation of vertical U-tube ground heat exchangers, ASHRAE Transactions 105 (2) (1999) 465–474.
G. Gan et al. / Applied Thermal Engineering 27 (2007) 430–441 [8] D. Pahud, B. Matthey, Comparison of the thermal performance of double U-pipe borehole heat exchangers measured in situ, Energy and Buildings 33 (5) (2001) 503–507. [9] Y. Bi, L. Chen, C. Wu, Ground heat exchanger temperature distribution analysis and experimental verification, Applied Thermal Engineering 22 (2) (2002) 183–189. [10] A. Hepbasli, O. Akdemir, E. Hancioglu, Experimental study of a closed loop vertical ground source heat pump system, Energy Conversion and Management 44 (4) (2003) 527–548. [11] H. Zeng, N. Diao, Z. Fang, Heat transfer analysis of boreholes in vertical ground heat exchangers, Heat and Mass Transfer 46 (23) (2003) 4467–4481. [12] V.C. Mei, Theoretical heat pump ground coil analysis with variable ground farfield boundary conditions, AIChE Journal 32 (7) (1986) 1211–1215. [13] O.J. Svec, J.H.L. Palmer, Performance of a spiral ground heat exchanger for heat pump application, Energy Research 13 (5) (1989) 503–510. [14] M. Inalli, H. Esen, Experimental thermal performance evaluation of a horizontal ground-source heat pump system, Applied Thermal Engineering 24 (14–15) (2004) 2219–2232. [15] R.R. Johnson, J.A. Edwards, J.C. Mulligan, Y. Mohammad-Zadeh, P. Safemazandarani, Experimental evaluation of surface tank ground-coupled heat pump system, Proceedings of the Intersociety Energy Conversion Engineering Conference 4 (1988) 53–58. [16] P.S. Doherty, S. Al-Huthaili, S.B. Riffat, N. Abodahab, Ground source heat pump – description and preliminary results of the Eco House system, Applied Thermal Engineering 24 (17–18) (2004) 2627–2641.
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¨ nsal, A computational model of a heat pump [17] R. Yumrutas, M. U system with a hemispherical surface tank as the ground heat source, Energy 25 (4) (2000) 371–388. ¨ nsal, Modelling of a space cooling system [18] R. Yumrutas, M. U with underground storage, Applied Thermal Engineering 25 (2005) 227–239. [19] Thermacore Europe Ltd., Northumberland, UK. [20] ASTM, D2216 – Test method for laboratory determination of water (moisture) content of soil and rock, American Society for Testing and Materials, USA, 1998. [21] D.A. de Vries, Thermal properties of soils, in: W.R. van Wijk (Ed.), Physics of Plant Environment, North-Holland, New York, 1963, pp. 210–235. [22] FLUENT, User’s Guide, Fluent Inc., USA, 1995. [23] CIBSE, Guide Volume A, Chartered Institution of Building Services Engineers, UK, 1999. [24] P.D. Dunn, D.A. Reay, Heat Pipes, fourth ed., Pergamon Press, Oxford, 1994. [25] J.M. Ochterbeck, Heat pipes, in: A. Bejan, A.D. Kraus (Eds.), Heat Transfer Handbook, John Wiley & Sons, New Jersey, USA, 2003, pp. 1181–1230. [26] G. Thyrum, E. Cruse, Heat pipe simulation: a simplified technique for modelling heat pipe assisted heat sinks, Advanced Packaging, 23–27 December 2001. Available from: . [27] A. Aziz, Conduction heat transfer, in: A. Bejan, A.D. Kraus (Eds.), Heat Transfer Handbook, John Wiley & Sons, New Jersey, USA, 2003, pp. 161–260.
