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A field study of the performance of a heat pump installed in a low energy house
Applied Thermal Engineering 71 (2014) 596e606
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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
A field study of the performance of a heat pump installed in a low energy house Mariusz Szreder* Department of Mechanical Systems Engineering and Automatization, Warsaw University of Technology Plock Campus, Jachowicza 2/4, 09-402 Plock, Poland
h i g h l i g h t s An efficiency field test of a heat pump was performed. The thermodynamic cycle of the heat pump was controlled by an electronic expansion valve. A high COP of the heat pump was obtained for underfloor space heating.
a r t i c l e i n f o
a b s t r a c t
Article history: Received 27 September 2013 Accepted 22 June 2014 Available online 28 June 2014
The experimental studies of the thermal performance of a ground source heat pump (GSHP) described in this paper were conducted in a single-family house. These studies were aimed at performing a field test for the heating capacity of a heat pump. The studies were conducted for 3 cases of variable demand for heat: when space heating was used on the ground floor (75 m2); when space heating was used on the ground and first floors (140 m2) and with the additional heating of domestic hot water mode on (hot water tank 250 L). During the measurement cycle the evaporation temperature of the R407C refrigerant was maintained at a constant level using an electronic expansion valve. Series of measurements were conducted for the evaporation temperature ranging from 5 C to 2.5 C. The COP increases along with the evaporation temperature of the R407C refrigerant, yet in the case of the heat pump the maximum COP value was obtained for the evaporation temperature about 2.5 C. The obtained results were considerably influenced by the temperature of glycol in the secondary circuit of the low temperature source. The heating performance of the heat pump within the investigated period ranged from 8.4 kW to 9.2 kW. While heating domestic hot water, the system could heat the water in the storage tank to a temperature of 45 C. The control system of the heat pump should enable keeping the condensation temperature as low as possible thus maintain a high COP value. The heating capacity can be adjusted to match the heat demand for space heating by controlling the duration of compressor operation. © 2014 Elsevier Ltd. All rights reserved.
Keywords: Heat pump Underfloor space heating Energy efficiency Field test Low energy house
1. Introduction Heat pumps have been used in Europe for 25 years, yet there has been a sudden increase in their sales recently. This results mainly from the EU Directives whose principal objective is to increase the share of Renewable Energy Sources (RES) in the final energy use up to 15% by 2020 [1]. The forecasts of the RES market development in Poland indicate that low temperature geothermal energy (heat
* Tel.: þ48 24 3675993; fax: þ48 24 262 65 42. E-mail address: [email protected]. http://dx.doi.org/10.1016/j.applthermaleng.2014.06.046 1359-4311/© 2014 Elsevier Ltd. All rights reserved.
pumps) may play a significant role in meeting the requirements of the said Directives. Heat pumps are fitted in popular heating systems in leading European countries (Sweden, Germany and France). These countries have the highest energy saving rates and ecological indexes [2]. These results from the effective promotion of heat pump technologies based on special electricity charge rates dedicated to such units and various forms of subsidies or tax incentives. Theoretical and experimental studies have confirmed that ground source heat pump (GSHP) offer the highest energy performance [3]. Advanced studies of vertical collectors with inclined boreholes were performed by Cui et al. [4]. The numerical and experimental analysis of horizontal collectors is available in the
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work by Esen et al. [5]. Literature also provides results of experimental studies concerning the performance analysis of heat pumps e.g. Nagano et al. [6], where the ground exchanger was used as ground heat source. The use of heat pumps for home heating, owing to its low maintenance costs, is also on the increase in Poland [7]. Because of moderate climate in Poland, GSHP are used with vertical and horizontal collectors. Recently, there has also been an increasing demand for high heating capacity heat pumps for heating multi-family buildings, offices, tourist and leisure facilities that are to undergo thermal upgrades. The market for air source heat pumps used only for producing domestic hot water has also been developing rapidly (Fig. 1). This mainly results from their relatively low price and easy of fitting. Compared to solar systems, they are cheaper and domestic hot water is available throughout the year. The use of an appropriate heat pump technology ensures the most advantageous economic indexes depending on the local climate and terrain conditions [8]. In Europe, heat pumps are equipped with two separate circuits for the heating system and production of domestic hot water. In the U.S. systems with an additional desuperheater are popular enabling a use of domestic hot water DHW and space heating simultaneously [9]. Experimental studies on hybrid ground source heat pumps (HGSHP) have been conducted. These studies aimed at obtaining a higher COP [10]. The demand for heating changes throughout the heating season. The results of experimental studies on heat pumps with variable volume [11] and variable speed compressors [12] are also available. Advanced research studies on various heat pump configurations are supported by experimental research conducted on prototypical research stands. In the literature, there is little information on energy efficiency analyses of heat pumps obtained in field tests [13]. Therefore, this paper describes a field test of a heat pump. The object under study was a single-family house equipped with a heat pump configuration commonly used in Poland.
2. Research station 2.1. Description of the research station A two-story single-family house with a total living area of 156 m2 was investigated. The building was constructed to use low energy technology. A low temperature floor heating system was used on the ground and first floors. In individual rooms, a heating
Fig. 1. Statistical data related to heat pumps installed in 2011 in Poland.
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system with 0.15 m spacing between the pipes was fitted with the exception of the bathrooms where the spacing was 0.1 m. The average length of individual circuits was 60e65 m. Separate manifolds powered by independent circulation pumps were used for heating circuits on the ground and first floors. The applied configuration made it possible to conduct a study of the heat pump for variable demand for the mass heat flow. In the first case only the underfloor space heating was attached only on the ground floor (75 m2 of 0.07 m thick concrete underlayment). In the second case was also attached underfloor space heating on the ground and first floors simultaneously (140 m2 of 0.07 m thick concrete underlayment). Design documentation and specifications give the yearly power requirement at a level of E ¼ 16,400 kWh. Assuming that, on average, the pump works for 1800 h, an approximate heat load of the building at a level of 9.1 kW was obtained. According to data available in relevant literature, the demand for heat energy for new low-energy single-family buildings amounts to 40e50 W/m2 [8]. Assuming the unit heat load of 50 W/m2 and 156 m2, an average heat load of the building amounting to 7.8 kW was obtained. If the heat pump also produces DHW, then, for a 4-person family, an additional 1 kW needs to be added [14]. Finally, for a 156 m2 house, a heat load of 8.8 kW was obtained. In round figures, for further calculations, it should be assumed that the required heating capacity of a heat pump amounts to 9 kW. 2.2. Basic components of a heat pump Heat pumps used for heating low energy houses are controlled by thermostatic expansion valves. Examining the assumed scope of experimental studies requires the use of a heat pump controlled by an electronic expansion valve. Therefore, a mass-market heat pump was not an option. Instead, components produced by reputable manufacturers from the refrigeration industry were used for its construction. Following the guidelines given in Ref. [15], R407C was used as a refrigerant and in the circuit of the condenser and evaporator, plate heat exchangers by WTK (Italy, www.wtk.it) and a spiral hermetic compressor made by SANYO were used. Because of an even load of individual phases of the electrical installation and following the experimental studies using a frequency inverter, a three-phase compressor was applied. A compressor was selected of a heating capacity of 9 kW reaching a COP ¼ 4.39 for the assumed conditions of B0W35 as specified in the manufacturer data given in Table 1. The data shown in Table 1 indicate that a change in the evaporation temperature practically does not result in a change of the motor power consumption. The compressor performance increases along with the evaporation temperature, therefore the COP is also proportionally higher for higher evaporation temperatures. This means that the evaporation temperature of the refrigerant should be maintained at the highest possible level. On the other hand, if the condensation temperature is increased, COP is significantly reduced. The applied R407C refrigerant is a three-component mixture. Each of these components has a different evaporation temperature. In order to ensure 100% evaporation of each of the components, a super-heater of drawn-in gas was used, following guidelines included in the manual [14]. Table 2 includes basic GSHP data, whereas Fig. 2 shows a concept diagram of the applied GSHP system. The plate heat exchangers were selected through Avogadro2 rev. 2.0. software. The standard working parameters for a brine heat pump are B0/W35, which denotes the refrigerant evaporation temperature (0 C) and the condensation temperature (35 C).
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Table 1 An excerpt from the catalog data for the compressor (8.1 C subcooling, 11.1 C superheating). Condensing temperature [ C]
Evaporating temperature [ C]
10
5
0
þ5
þ10
30
Capacity [W] Power consumption Rated current [A] COP Capacity [W] Power consumption Rated current [A] COP Capacity [W] Power consumption Rated current [A] COP Capacity [W] Power consumption Rated current [A] COP
6638 1889 3.7 3.51 6127 2079 4.0 2.95 5645 2310 4.5 2.44 5192 2583 5.0 2.01
8081 1890 3.8 4.28 7473 2080 4.1 3.59 6901 2313 4.5 2.98 6365 2587 5.0 2.46
9846 1885 3.8 5.22 9123 2076 4.2 4.39 8442 2309 4.6 3.66 7805 2584 5.0 3.02
12,005 1873 3.8 6.41 11,146 2064 4.2 5.40 10,338 2298 4.6 4.5 9580 2574 5.0 3.72
e e e e 13,614 2044 4.2 6.66 12,660 2279 4.5 5.56 11,762 2556 5.0 4.60
35
40
45
[W]
[W]
[W]
[W]
Assuming an expected temperature difference in the glycol circuit amounting to DT ¼ 3 C and 5 C for superheating and subcooling of the refrigerant, a P7 series 30-plate heat exchanger was selected for the single-family houses. In the condenser circuit, a P7 series 16-plate exchanger was selected for the temperature difference at the inlet and outlet of the underfloor heating of DT ¼ 8e10 C and for the assumed flow of 1.2 m3/h.
the 2012/13 heating season from November to April. The research station used a system of dual electricity charge rate, therefore, for economical reasons, experimental studies were conducted with a maximum use of the G12 charge rate, (a.k.a. night rate). The heat pump was used in the underfloor heating in constant cycles: Cycle 1 working-time 60 min, break time 60 min, Cycle 2 working-time 120 min, break time 60 min.
2.3. Ground loop selection Due to limited plot development possibilities and unfavorable ground conditions (sand at a depth of 1e2 m) the idea of installing a ground loop as a horizontal collector was given up. Due to high costs of TRT tests, following Mattsson's suggestion [16], tests of heat transfer capabilities of the geothermal probes were also given up. Dedicated Energeo software, by Aspol, Poland (a manufacturer of vertical exchanger probes) was used for selecting the vertical collector. After inputting data into the software for the assumed heat energy coefficient of 38 W/m, the required length of the vertical exchanger amounting to 186 m was obtained. U-tube probes filled with a 20% solution of ethylene glycol (Henock 20E15) were used. Three boreholes were made, 62 m each and individual vertical exchangers were connected to a wall-mounted heating manifold and hydraulically balanced with ball valves. 2.4. Testing procedure After preparing the research station, individual measurement circuits were calibrated. The specifications of the measuring transducers were given in Table 3. The studies were conducted in
Type
Compressor R407C refrigerant, scroll type, Sanyo Condenser Brazed plate heat Exchange, WTK Evaporator Brazed plate heat Exchange, WTK Expansion Electronic proportional, valve E2V Carel Tank
bivalent
using underfloor heating mode on the ground floor (75 m2), using underfloor heating mode simultaneously on the ground floor and on the first floor (140 m2). In order to determine the heating capacity of the condenser, a standard relation was applied
Qcon ¼ Cp;w Mw Tw;out Tw;in
Table 2 Main components of the GSHP. Component
Basic working parameters, vital for determining of the energy balance of the heat pump, were recorded during the studies. The use of two-gear circulation pumps in individual ground loops and underfloor heating circuits made it possible to easily adjust the mass flow of water and volume flow of brine. The use of an electronic expansion valve in the freon circuit provided the possibility of conducting studies on the heat pump for the refrigerant evaporation temperatures ranging from 5 C to 2.5 C. The use of an inverter to power the compressor made it possible to examine the influence of the frequency of the supply voltage on the heat pump basic parameters. The studies were conducted for the underfloor heating and domestic hot water heating modes. For the assumed constant evaporation temperatures of the refrigerant, a number of tests for variable heat demands were conducted. For the space heating, studies were conducted for the following cases:
Properties B0/W35: Q ¼ 9.1 kW, Wcom ¼ 2.07 kW, COP ¼ 4.4 P7-16, area ¼ 0.98 m2, Q ¼ 13.16 kW, flow ¼ 0.3 kg/s P7-30, area ¼ 1.96 m2, Q ¼ 7.53 kW, flow ¼ 0.48 kg/s 0e4.0 MPa, nozzle with a travel of over 14 mm on around 500 steps of the built-in motor 250 L
(1)
where: Mw denotes the mass flow of water. The calculations of the cooling effect of the evaporator were made in an analogical manner. The coefficient of performance of the heat pump at any point in time (t) was determined from a standard relation in the form of
COPHP
Qcon ðtÞ Wcom ðtÞ
(2)
In the space heating circuit, a change in the thermal load (resulting from changed water flowrate or changed heater configuration) forces
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Fig. 2. Schematic diagram of a heat pump experimental system.
a change in the refrigerant condensation temperature along with changes in inlet and outlet temperature of water. Consequently, the average values of COP in an individual measurement cycle was determined from the relationship (3).
Z
t
COP ¼ Z 0t 0
Qcom ðtÞdt (3) Wcom ðtÞdt
where t denotes the duration of the cycle.
3. Results of experimental studies and discussion Experimental studies were conducted at the research station equipped with typically configured heat pumps available for single-family houses. Independently controlled circulation pumps for water and brine were used in the designed installation. The volume flow of glycol in the ground loop obtained during experimental studies was 1.75 m3/h in gear 1 and 2.30 m3/h in gear 2 for simultaneous operation of all 3 loops of the ground source. In the heating circuit, the mass flow of water amounted to 0.09 kg/s in gear 1 and 0.13 kg/s in gear 2 when the circulation pump worked on the ground floor and 0.16 and 0.22 kg/s when the circulation pumps worked simultaneously on the ground and first floors. The COP and heating capacity were determined for each measurement cycle. Upon determining of the measurement errors, the uncertainty of the measurements was estimated by the method of total differential. A calculated measurement error amounted to ±2.45% for the COP and ±3.15% for the heating capacity.
Table 3 Description of the measurement system components. Component
Type
Properties
Pressure sensors Temperature sensors Power meter Flow meter
Ratiometric pressure transmitter, Danfoss PT100
AKS32R-HP 1e34 bar, ±3% AKS32R-LP 1e9 bar, ±3% In-tube installation, 100 < T < 600 C, ±0.5% 0e300 V/0e10 A, ±0.05% 40 < T < 180 C, ±0.2%
Magnetic, MAG, BMETERS
3.1. Space heating Figs. 3e5 show examples of the temperature change patterns of basic elements of the heating system recorded for individual configuration parameters. The data given in Fig. 3 indicate that the temperature of brine recorded at the inlet and outlet of the evaporator decreases successively during the operation of the heat pump. While collecting heat from the ground through the vertical exchangers, the temperature of the ground around these exchangers decreased. During the first hour of the measurement cycle for the assumed constant evaporation temperature of the R407C refrigerant, the efficiency of heat transmission in the evaporator decreased resulting from a decrease in the temperature difference between the glycol and the freon circuits. As a result, a drop of 4% in the condensation temperature of water at the outlet of the condenser was observed. Adjusting the evaporation temperature of freon to 5 C during the next hour of the measurement cycle restored proper heat exchange conditions inside the evaporator (a required difference of temperatures between the glycol and freon circuits). An increase of 5% in the condensation temperature of freon and water at the outlet of the condenser was observed. According to the specifications of the compressor given in Table 1, lowering the evaporation temperature of freon causes a decrease in the heat pump COP. The data included in Fig. 3 indicate that during the first and second hour of the measurement cycle, comparable values of COP were obtained. It was assumed that a higher COP would be obtained in the freon evaporation temperature of 0 C than in the temperature of 5 C. In the 1st hour of the cycle, a lowered heat exchange efficiency in the evaporator prevented the obtainment of a higher COP for the evaporation temperature of 0 C. Fig. 4 presents a chart of temperature changes when the underfloor circuit of the heating system worked on the ground floor. At a constant cooling capacity of the heat pump a change of the mass flow of water from 0.09 to 0.13 kg/s resulted in a water temperature drop at the outlet of the heat exchanger. Increasing the mass flow of water in the heating circuit resulted in a reduction of water temperature differences at the inlet and outlet of the underfloor heating system (DTsh ¼ 18 C / DTsh ¼ 14 C). Obtaining a higher COP of the heat pump is possible through increasing the mass flow of water in the heating circuit.
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Fig. 3. A chart of temperature changes in the underfloor circuit for various evaporation temperatures of the refrigerant.
Fig. 5 shows that when the mass flow of water was set at 0.13 kg/ s in the underfloor circuit on the ground floor a water temperature increase of 14% at the outlet of the condenser was obtained. Simultaneous work of the underfloor circuit on the ground and first floors, caused an increased demand for heat in the heating circuit. This induced a decrease in the water temperature at the outlet of the condenser and the condensation temperature of freon. Consequently, a lower condensation temperature resulted in a higher COP of the heat pump. Maintaining a constant water temperature at the inlet of the underfloor heating system during this measurement cycle resulted from a dual volume increase of the heat receiver (increased thermal inertia of the system). Fig. 6 shows the influence of the refrigerant evaporation temperature on the COP value depending on the variable demand for the heat in underfloor heating. Increasing the mass flow of water
from 0.09 to 0.13 kg/s for space heating working on the ground floor caused the COP to increase by 6%. In the case of a simultaneous operation of ground and first floor heating, increasing the mass flow of water from 0.16 to 0.22 kg/s resulted in a COP increase by 3%. The specifications of the compressor given in Table 1 indicate that an increase in the refrigerant evaporation temperature causes an increase in the COP. The data presented in Fig. 6 indicate that the highest COP was obtained for the evaporation temperature about 2.5 C. Whereas, at the evaporation temperature of 2.5 C, a significant COP decrease was observed. This outcome was significantly influenced by the average temperature of glycol in the ground loop circuit (heat exchange efficiency of the evaporator). The course of the Tgs temperature changes is presented in Fig. 10. Fig. 7 shows the influence of water temperature differences at the inlet and outlet of the space heating system DTsh on the COP
Fig. 4. A chart of temperature changes in the underfloor circuit for a changing mass flow of water.
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Fig. 5. A chart of temperature changes in the underfloor circuit for 75 and 140 m2 of space heating.
value and heating capacity Qcon. The data given indicate that, for DTsh ¼ 18 C (ground floor, 0.09 kg/s), the lowest COP and Qcon values were obtained, whereas the highest values of the COP and Qcon were obtained for DTsh ¼ 8 C (ground floor and first floor, 0.22 kg/s). Therefore, maintaining the lowest possible temperature difference in the heating circuit should be aimed at. At a constant heating capacity of the heat pump under given work conditions, increasing the mass flow of water results in a decrease of the condensation temperature of the refrigerant and water at the outlet of the condenser. A lower condensation temperature causes COP and Qcon values to increase. 3.2. Domestic hot water heating Domestic hot water heating is possible after the switching of the heat pump to DHW mode. In this mode, the circulation pumps of the space heating are off, whereas the DHW pump is on. The
domestic hot water tank with a capacity of 250 L, has a significantly lower thermal inertia in comparison with the underfloor space heating, which results in greater dynamics of the temperature changes in the heating circuit. In Fig. 8 it is clearly visible that, at a constant cooling capacity of the heat pump, a temperature increase in the DHW tank induces an increase in the refrigerant condensation temperature and water temperature at the inlet and outlet of the tank circuit. A water temperature rise at the outlet of the condenser is proportional to a water temperature increase in the DHW tank and amounts to 70%. In one hour of operation of the heat pump, water in the DHW tank is heated from 20 C to 42 C with an average COP ¼ 3.28 for this measurement cycle. Fig. 9 presents the influence of water temperature in the tank on a COP as well as the current supplied to the compressor drive. A water temperature increase in the tank causes a condensation temperature increase of the R407C refrigerant, which consequently
Fig. 6. A chart of the COP depending on the evaporation temperature of the refrigerant.
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Fig. 7. The COP and Qcon depending on water temperature differences in the heating circuit.
induces increased current supplied to the compressor drive along with a decrease in the COP. At the condensation temperature of 50 C, the heat pump energy efficiency decreases to COP ¼ 2.7, whereas the current supplied to 6 A. It is a limit value for the compressor motor, therefore, obtaining a higher water temperature in the tank requires using a heater built in the tank. 3.3. Ground loop During the heating season, an increased demand for heat results in cooling of the ground around the vertical collectors, which manifests itself as a drop in the glycol temperature at the evaporator inlet. According to the trend presented in Fig. 10, the lowest temperature values in the glycol circuit were observed from December to March (the highest demand for thermal energy). Outside temperatures in January and February that were higher
than the statistical average, as well as a relatively low outside temperature in March delayed the temperature increase in the glycol circuit until April. At an average temperature of glycol of 3 C at the evaporator outlet, setting the evaporation temperature at 2.5 C does not guarantee suitable conditions for exchanging heat in the evaporator and results in lowering of the heat pump energy efficiency. The possibility of activation of the circulation pumps independently in the circuits of the vertical collectors, made it possible to carry out tests for cases in which one or two vertical collectors are not working. Based on the conducted studies it may be assessed how the performance of the ground loop changes in case one of the vertical collectors fails. Fig. 11 shows a trend in glycol temperature changes at the inlet and outlet of the evaporator when one, two and three vertical collectors are subsequently activated. For a better presentation of
Fig. 8. A chart of temperature changes in the DHW circuit.
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Fig. 9. The influence of water temperature in the tank on COP as well as current supplied and condensation temperature.
the temperature changes of the ground loop, the results of 3 consecutive measurement cycles were shown in one chart. For all three vertical collectors working simultaneously, an average temperature difference between the inlet and outlet of the evaporator of DTgs ¼ 2.7 C was recorded along with 23% glycol temperature drop at the evaporator inlet within a 30min measurement cycle. When two collectors were activated simultaneously, the recorded values where DTgs ¼ 3.3 C and 31% respectively. Switching off one collector from the circuit resulted in a DTgs increase (lower volume flow of glycol in the evaporator) and an increased glycol temperature drop rate during the measurement cycle. When only one collector was activated, the recorded values were DTgs ¼ 5.8 C and the drop of the glycol temperature 37%. In this case, after 20 min from activation, the glycol temperature at the evaporator outlet decreased to 0 C and the immobilizer of the controller activated (protection against excessive cooling of the ground around the collector). Fig. 12 shows the trend in the temperature changes in the glycol circuit for cycle 2 (working time e 120 min and break time e
60 min). From the analysis of the presented data it may be concluded that the adopted break time of the heat pump is sufficient to level the ground temperature around the vertical collectors. Additionally, the chart presents the course of the floor temperature changes when the heating circuit is activated on the ground floor. 3.4. Summary and results The research studies conducted in the 2012/2013 heating season indicate that the greatest demand for heat was recorded in December (Fig. 13). The outside temperature in January and February, higher than the statistical average, resulted in a significantly lower demand for heat in this period. Low outside temperature in March rendered the demand for heat in that month comparable to the demand observed in January. On average, the circulation pumps used 35 W per hour in the first gear and 50 W per hour in the second gear. The total power consumption of the circulation pumps constituted 3.5% of the compressor power consumption. Admittedly, setting the circulation pump in gear 2 increased the power consumption by 15 W, but
Fig. 10. A chart of temperature changes in the glycol circuit during the heating season.
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Fig. 11. A chart of temperature changes in the glycol circuit when loops 1, 2, and 3 are working.
an increased mass flow of water resulted in a decrease of DTsh as well as a decrease in the condensation temperature. It may be calculated from the recorded data that for the COP ¼ 4.18 the compressor uses 125 W less energy for the production of the same amount of heat than for the COP ¼ 3.95. Therefore, it is economically justified to set the circulation pumps to gear 2 in view of greater savings in the compressor power consumption. The COP increases alongside with an increase in the evaporation temperature of the R407C refrigerant. In the case of GSHP, however, the maximum COP value was obtained for the evaporation temperature about 2.5 C. The results obtained were significantly influenced by the temperature of glycol in the ground loop. This problem was described whilst analyzing the data presented in Fig. 3. The performance of the heat pump during the investigated period changed in the range 8.4 kWe9.2 kW. The highest value was obtained for DTsh ¼ 8 C (gear 2 of the heat pump, 140 m2). The lowest value of heating capacity was obtained for DTsh ¼ 18 C (gear
1 of the heat pump, 75 m2). Therefore, it may be assumed that setting higher values of the mass flow of water in the underfloor circuit induces a reduction of the condensation temperature, which ensures favorable conditions for the obtainment of higher COP and heating capacity values. Additionally, studies with the use of an inverter for the drive of the heat pump compressor were conducted. In the case when a hermetic spiral compressor was used, it was possible to carry out studies in the range of 40e60 Hz. The results presented in Fig. 14 indicate that increasing the frequency of the compressor drive causes an increase in the heating capacity, the COP, however lowers as a consequence. Therefore, the use of an inverter for the drive of the spiral compressor does not result in significant economic benefits. The control system of the heat pump should enable keeping the condensation temperature as low as possible thus maintain a high COP value.
Fig. 12. A chart of temperature changes in the glycol circuit when space heating is working on the ground floor.
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Fig. 13. A summary of power consumption and production of heating energy during particular months of the heating season.
Fig. 14. The influence of the compressor's propulsion frequency on the COP and heating capacity at the evaporation temperature of 2.5 C.
The heating capacity can be adjusted to match the heat demand for space heating by controlling the duration of compressor operation.
4. Conclusions The heating system with the ground source heat pump was successfully installed and investigated in a low energy building. The results obtained were presented and analyzed. The research results presented may lead to the following conclusions: 1. The control system of the heat pump should enable keeping the condensation temperature as low as possible thus maintain a high COP value. 2. The use of an inverter for the drive of the spiral compressor does not bring significant economic benefits.
3. Using the space heating system on the ground floor and on the first floor simultaneously (140 m2 of the floor surface) led to obtaining the highest COP and Qcon values. 4. The use of several vertical collectors in the installation of the ground loop makes it possible to operate the heat pump in case of failure in one of the collectors. 5. The results of operation of the compressor are in line with its specifications. 6. The configuration of the heat pump used in the heating system made it possible to obtain a high COP for space heating, whereas, it failed to meet the expectations as regards the heating of hot domestic water above 45 C. Acknowledgements The author would like to thank Professor Krzysztof Urbaniec from Warsaw University of Technology e Plock Campus for his useful comments on the preparation of this paper.
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Nomenclature
References
Cp E M Q T Wcom
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specific heat, [kJ/kg K] energy mass flow, [kg/s] heating capacity, [kW] temperature, [ C] compressor power input, [kW]
Abbreviations COP coefficient of performance DHW domestic hot water GSHP ground source heat pump HGSHP hybrid ground source heat pump RES renewable energy sources TRT thermal response test Subscripts com compressor con condenser ev evaporator fl floor gs ground source in inlet out outlet sh space heating ta tank w water 1gf pump 1st gear on the ground floor 1gff pump 1st gear on the ground and first floors 2gf pump 2nd gear on the ground floor 2gff pump 2nd gear on the ground and first floors
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A field study of the performance of a heat pump installed in a low energy house Mariusz Szreder* Department of Mechanical Systems Engineering and Automatization, Warsaw University of Technology Plock Campus, Jachowicza 2/4, 09-402 Plock, Poland
h i g h l i g h t s An efficiency field test of a heat pump was performed. The thermodynamic cycle of the heat pump was controlled by an electronic expansion valve. A high COP of the heat pump was obtained for underfloor space heating.
a r t i c l e i n f o
a b s t r a c t
Article history: Received 27 September 2013 Accepted 22 June 2014 Available online 28 June 2014
The experimental studies of the thermal performance of a ground source heat pump (GSHP) described in this paper were conducted in a single-family house. These studies were aimed at performing a field test for the heating capacity of a heat pump. The studies were conducted for 3 cases of variable demand for heat: when space heating was used on the ground floor (75 m2); when space heating was used on the ground and first floors (140 m2) and with the additional heating of domestic hot water mode on (hot water tank 250 L). During the measurement cycle the evaporation temperature of the R407C refrigerant was maintained at a constant level using an electronic expansion valve. Series of measurements were conducted for the evaporation temperature ranging from 5 C to 2.5 C. The COP increases along with the evaporation temperature of the R407C refrigerant, yet in the case of the heat pump the maximum COP value was obtained for the evaporation temperature about 2.5 C. The obtained results were considerably influenced by the temperature of glycol in the secondary circuit of the low temperature source. The heating performance of the heat pump within the investigated period ranged from 8.4 kW to 9.2 kW. While heating domestic hot water, the system could heat the water in the storage tank to a temperature of 45 C. The control system of the heat pump should enable keeping the condensation temperature as low as possible thus maintain a high COP value. The heating capacity can be adjusted to match the heat demand for space heating by controlling the duration of compressor operation. © 2014 Elsevier Ltd. All rights reserved.
Keywords: Heat pump Underfloor space heating Energy efficiency Field test Low energy house
1. Introduction Heat pumps have been used in Europe for 25 years, yet there has been a sudden increase in their sales recently. This results mainly from the EU Directives whose principal objective is to increase the share of Renewable Energy Sources (RES) in the final energy use up to 15% by 2020 [1]. The forecasts of the RES market development in Poland indicate that low temperature geothermal energy (heat
* Tel.: þ48 24 3675993; fax: þ48 24 262 65 42. E-mail address: [email protected]. http://dx.doi.org/10.1016/j.applthermaleng.2014.06.046 1359-4311/© 2014 Elsevier Ltd. All rights reserved.
pumps) may play a significant role in meeting the requirements of the said Directives. Heat pumps are fitted in popular heating systems in leading European countries (Sweden, Germany and France). These countries have the highest energy saving rates and ecological indexes [2]. These results from the effective promotion of heat pump technologies based on special electricity charge rates dedicated to such units and various forms of subsidies or tax incentives. Theoretical and experimental studies have confirmed that ground source heat pump (GSHP) offer the highest energy performance [3]. Advanced studies of vertical collectors with inclined boreholes were performed by Cui et al. [4]. The numerical and experimental analysis of horizontal collectors is available in the
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work by Esen et al. [5]. Literature also provides results of experimental studies concerning the performance analysis of heat pumps e.g. Nagano et al. [6], where the ground exchanger was used as ground heat source. The use of heat pumps for home heating, owing to its low maintenance costs, is also on the increase in Poland [7]. Because of moderate climate in Poland, GSHP are used with vertical and horizontal collectors. Recently, there has also been an increasing demand for high heating capacity heat pumps for heating multi-family buildings, offices, tourist and leisure facilities that are to undergo thermal upgrades. The market for air source heat pumps used only for producing domestic hot water has also been developing rapidly (Fig. 1). This mainly results from their relatively low price and easy of fitting. Compared to solar systems, they are cheaper and domestic hot water is available throughout the year. The use of an appropriate heat pump technology ensures the most advantageous economic indexes depending on the local climate and terrain conditions [8]. In Europe, heat pumps are equipped with two separate circuits for the heating system and production of domestic hot water. In the U.S. systems with an additional desuperheater are popular enabling a use of domestic hot water DHW and space heating simultaneously [9]. Experimental studies on hybrid ground source heat pumps (HGSHP) have been conducted. These studies aimed at obtaining a higher COP [10]. The demand for heating changes throughout the heating season. The results of experimental studies on heat pumps with variable volume [11] and variable speed compressors [12] are also available. Advanced research studies on various heat pump configurations are supported by experimental research conducted on prototypical research stands. In the literature, there is little information on energy efficiency analyses of heat pumps obtained in field tests [13]. Therefore, this paper describes a field test of a heat pump. The object under study was a single-family house equipped with a heat pump configuration commonly used in Poland.
2. Research station 2.1. Description of the research station A two-story single-family house with a total living area of 156 m2 was investigated. The building was constructed to use low energy technology. A low temperature floor heating system was used on the ground and first floors. In individual rooms, a heating
Fig. 1. Statistical data related to heat pumps installed in 2011 in Poland.
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system with 0.15 m spacing between the pipes was fitted with the exception of the bathrooms where the spacing was 0.1 m. The average length of individual circuits was 60e65 m. Separate manifolds powered by independent circulation pumps were used for heating circuits on the ground and first floors. The applied configuration made it possible to conduct a study of the heat pump for variable demand for the mass heat flow. In the first case only the underfloor space heating was attached only on the ground floor (75 m2 of 0.07 m thick concrete underlayment). In the second case was also attached underfloor space heating on the ground and first floors simultaneously (140 m2 of 0.07 m thick concrete underlayment). Design documentation and specifications give the yearly power requirement at a level of E ¼ 16,400 kWh. Assuming that, on average, the pump works for 1800 h, an approximate heat load of the building at a level of 9.1 kW was obtained. According to data available in relevant literature, the demand for heat energy for new low-energy single-family buildings amounts to 40e50 W/m2 [8]. Assuming the unit heat load of 50 W/m2 and 156 m2, an average heat load of the building amounting to 7.8 kW was obtained. If the heat pump also produces DHW, then, for a 4-person family, an additional 1 kW needs to be added [14]. Finally, for a 156 m2 house, a heat load of 8.8 kW was obtained. In round figures, for further calculations, it should be assumed that the required heating capacity of a heat pump amounts to 9 kW. 2.2. Basic components of a heat pump Heat pumps used for heating low energy houses are controlled by thermostatic expansion valves. Examining the assumed scope of experimental studies requires the use of a heat pump controlled by an electronic expansion valve. Therefore, a mass-market heat pump was not an option. Instead, components produced by reputable manufacturers from the refrigeration industry were used for its construction. Following the guidelines given in Ref. [15], R407C was used as a refrigerant and in the circuit of the condenser and evaporator, plate heat exchangers by WTK (Italy, www.wtk.it) and a spiral hermetic compressor made by SANYO were used. Because of an even load of individual phases of the electrical installation and following the experimental studies using a frequency inverter, a three-phase compressor was applied. A compressor was selected of a heating capacity of 9 kW reaching a COP ¼ 4.39 for the assumed conditions of B0W35 as specified in the manufacturer data given in Table 1. The data shown in Table 1 indicate that a change in the evaporation temperature practically does not result in a change of the motor power consumption. The compressor performance increases along with the evaporation temperature, therefore the COP is also proportionally higher for higher evaporation temperatures. This means that the evaporation temperature of the refrigerant should be maintained at the highest possible level. On the other hand, if the condensation temperature is increased, COP is significantly reduced. The applied R407C refrigerant is a three-component mixture. Each of these components has a different evaporation temperature. In order to ensure 100% evaporation of each of the components, a super-heater of drawn-in gas was used, following guidelines included in the manual [14]. Table 2 includes basic GSHP data, whereas Fig. 2 shows a concept diagram of the applied GSHP system. The plate heat exchangers were selected through Avogadro2 rev. 2.0. software. The standard working parameters for a brine heat pump are B0/W35, which denotes the refrigerant evaporation temperature (0 C) and the condensation temperature (35 C).
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Table 1 An excerpt from the catalog data for the compressor (8.1 C subcooling, 11.1 C superheating). Condensing temperature [ C]
Evaporating temperature [ C]
10
5
0
þ5
þ10
30
Capacity [W] Power consumption Rated current [A] COP Capacity [W] Power consumption Rated current [A] COP Capacity [W] Power consumption Rated current [A] COP Capacity [W] Power consumption Rated current [A] COP
6638 1889 3.7 3.51 6127 2079 4.0 2.95 5645 2310 4.5 2.44 5192 2583 5.0 2.01
8081 1890 3.8 4.28 7473 2080 4.1 3.59 6901 2313 4.5 2.98 6365 2587 5.0 2.46
9846 1885 3.8 5.22 9123 2076 4.2 4.39 8442 2309 4.6 3.66 7805 2584 5.0 3.02
12,005 1873 3.8 6.41 11,146 2064 4.2 5.40 10,338 2298 4.6 4.5 9580 2574 5.0 3.72
e e e e 13,614 2044 4.2 6.66 12,660 2279 4.5 5.56 11,762 2556 5.0 4.60
35
40
45
[W]
[W]
[W]
[W]
Assuming an expected temperature difference in the glycol circuit amounting to DT ¼ 3 C and 5 C for superheating and subcooling of the refrigerant, a P7 series 30-plate heat exchanger was selected for the single-family houses. In the condenser circuit, a P7 series 16-plate exchanger was selected for the temperature difference at the inlet and outlet of the underfloor heating of DT ¼ 8e10 C and for the assumed flow of 1.2 m3/h.
the 2012/13 heating season from November to April. The research station used a system of dual electricity charge rate, therefore, for economical reasons, experimental studies were conducted with a maximum use of the G12 charge rate, (a.k.a. night rate). The heat pump was used in the underfloor heating in constant cycles: Cycle 1 working-time 60 min, break time 60 min, Cycle 2 working-time 120 min, break time 60 min.
2.3. Ground loop selection Due to limited plot development possibilities and unfavorable ground conditions (sand at a depth of 1e2 m) the idea of installing a ground loop as a horizontal collector was given up. Due to high costs of TRT tests, following Mattsson's suggestion [16], tests of heat transfer capabilities of the geothermal probes were also given up. Dedicated Energeo software, by Aspol, Poland (a manufacturer of vertical exchanger probes) was used for selecting the vertical collector. After inputting data into the software for the assumed heat energy coefficient of 38 W/m, the required length of the vertical exchanger amounting to 186 m was obtained. U-tube probes filled with a 20% solution of ethylene glycol (Henock 20E15) were used. Three boreholes were made, 62 m each and individual vertical exchangers were connected to a wall-mounted heating manifold and hydraulically balanced with ball valves. 2.4. Testing procedure After preparing the research station, individual measurement circuits were calibrated. The specifications of the measuring transducers were given in Table 3. The studies were conducted in
Type
Compressor R407C refrigerant, scroll type, Sanyo Condenser Brazed plate heat Exchange, WTK Evaporator Brazed plate heat Exchange, WTK Expansion Electronic proportional, valve E2V Carel Tank
bivalent
using underfloor heating mode on the ground floor (75 m2), using underfloor heating mode simultaneously on the ground floor and on the first floor (140 m2). In order to determine the heating capacity of the condenser, a standard relation was applied
Qcon ¼ Cp;w Mw Tw;out Tw;in
Table 2 Main components of the GSHP. Component
Basic working parameters, vital for determining of the energy balance of the heat pump, were recorded during the studies. The use of two-gear circulation pumps in individual ground loops and underfloor heating circuits made it possible to easily adjust the mass flow of water and volume flow of brine. The use of an electronic expansion valve in the freon circuit provided the possibility of conducting studies on the heat pump for the refrigerant evaporation temperatures ranging from 5 C to 2.5 C. The use of an inverter to power the compressor made it possible to examine the influence of the frequency of the supply voltage on the heat pump basic parameters. The studies were conducted for the underfloor heating and domestic hot water heating modes. For the assumed constant evaporation temperatures of the refrigerant, a number of tests for variable heat demands were conducted. For the space heating, studies were conducted for the following cases:
Properties B0/W35: Q ¼ 9.1 kW, Wcom ¼ 2.07 kW, COP ¼ 4.4 P7-16, area ¼ 0.98 m2, Q ¼ 13.16 kW, flow ¼ 0.3 kg/s P7-30, area ¼ 1.96 m2, Q ¼ 7.53 kW, flow ¼ 0.48 kg/s 0e4.0 MPa, nozzle with a travel of over 14 mm on around 500 steps of the built-in motor 250 L
(1)
where: Mw denotes the mass flow of water. The calculations of the cooling effect of the evaporator were made in an analogical manner. The coefficient of performance of the heat pump at any point in time (t) was determined from a standard relation in the form of
COPHP
Qcon ðtÞ Wcom ðtÞ
(2)
In the space heating circuit, a change in the thermal load (resulting from changed water flowrate or changed heater configuration) forces
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Fig. 2. Schematic diagram of a heat pump experimental system.
a change in the refrigerant condensation temperature along with changes in inlet and outlet temperature of water. Consequently, the average values of COP in an individual measurement cycle was determined from the relationship (3).
Z
t
COP ¼ Z 0t 0
Qcom ðtÞdt (3) Wcom ðtÞdt
where t denotes the duration of the cycle.
3. Results of experimental studies and discussion Experimental studies were conducted at the research station equipped with typically configured heat pumps available for single-family houses. Independently controlled circulation pumps for water and brine were used in the designed installation. The volume flow of glycol in the ground loop obtained during experimental studies was 1.75 m3/h in gear 1 and 2.30 m3/h in gear 2 for simultaneous operation of all 3 loops of the ground source. In the heating circuit, the mass flow of water amounted to 0.09 kg/s in gear 1 and 0.13 kg/s in gear 2 when the circulation pump worked on the ground floor and 0.16 and 0.22 kg/s when the circulation pumps worked simultaneously on the ground and first floors. The COP and heating capacity were determined for each measurement cycle. Upon determining of the measurement errors, the uncertainty of the measurements was estimated by the method of total differential. A calculated measurement error amounted to ±2.45% for the COP and ±3.15% for the heating capacity.
Table 3 Description of the measurement system components. Component
Type
Properties
Pressure sensors Temperature sensors Power meter Flow meter
Ratiometric pressure transmitter, Danfoss PT100
AKS32R-HP 1e34 bar, ±3% AKS32R-LP 1e9 bar, ±3% In-tube installation, 100 < T < 600 C, ±0.5% 0e300 V/0e10 A, ±0.05% 40 < T < 180 C, ±0.2%
Magnetic, MAG, BMETERS
3.1. Space heating Figs. 3e5 show examples of the temperature change patterns of basic elements of the heating system recorded for individual configuration parameters. The data given in Fig. 3 indicate that the temperature of brine recorded at the inlet and outlet of the evaporator decreases successively during the operation of the heat pump. While collecting heat from the ground through the vertical exchangers, the temperature of the ground around these exchangers decreased. During the first hour of the measurement cycle for the assumed constant evaporation temperature of the R407C refrigerant, the efficiency of heat transmission in the evaporator decreased resulting from a decrease in the temperature difference between the glycol and the freon circuits. As a result, a drop of 4% in the condensation temperature of water at the outlet of the condenser was observed. Adjusting the evaporation temperature of freon to 5 C during the next hour of the measurement cycle restored proper heat exchange conditions inside the evaporator (a required difference of temperatures between the glycol and freon circuits). An increase of 5% in the condensation temperature of freon and water at the outlet of the condenser was observed. According to the specifications of the compressor given in Table 1, lowering the evaporation temperature of freon causes a decrease in the heat pump COP. The data included in Fig. 3 indicate that during the first and second hour of the measurement cycle, comparable values of COP were obtained. It was assumed that a higher COP would be obtained in the freon evaporation temperature of 0 C than in the temperature of 5 C. In the 1st hour of the cycle, a lowered heat exchange efficiency in the evaporator prevented the obtainment of a higher COP for the evaporation temperature of 0 C. Fig. 4 presents a chart of temperature changes when the underfloor circuit of the heating system worked on the ground floor. At a constant cooling capacity of the heat pump a change of the mass flow of water from 0.09 to 0.13 kg/s resulted in a water temperature drop at the outlet of the heat exchanger. Increasing the mass flow of water in the heating circuit resulted in a reduction of water temperature differences at the inlet and outlet of the underfloor heating system (DTsh ¼ 18 C / DTsh ¼ 14 C). Obtaining a higher COP of the heat pump is possible through increasing the mass flow of water in the heating circuit.
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Fig. 3. A chart of temperature changes in the underfloor circuit for various evaporation temperatures of the refrigerant.
Fig. 5 shows that when the mass flow of water was set at 0.13 kg/ s in the underfloor circuit on the ground floor a water temperature increase of 14% at the outlet of the condenser was obtained. Simultaneous work of the underfloor circuit on the ground and first floors, caused an increased demand for heat in the heating circuit. This induced a decrease in the water temperature at the outlet of the condenser and the condensation temperature of freon. Consequently, a lower condensation temperature resulted in a higher COP of the heat pump. Maintaining a constant water temperature at the inlet of the underfloor heating system during this measurement cycle resulted from a dual volume increase of the heat receiver (increased thermal inertia of the system). Fig. 6 shows the influence of the refrigerant evaporation temperature on the COP value depending on the variable demand for the heat in underfloor heating. Increasing the mass flow of water
from 0.09 to 0.13 kg/s for space heating working on the ground floor caused the COP to increase by 6%. In the case of a simultaneous operation of ground and first floor heating, increasing the mass flow of water from 0.16 to 0.22 kg/s resulted in a COP increase by 3%. The specifications of the compressor given in Table 1 indicate that an increase in the refrigerant evaporation temperature causes an increase in the COP. The data presented in Fig. 6 indicate that the highest COP was obtained for the evaporation temperature about 2.5 C. Whereas, at the evaporation temperature of 2.5 C, a significant COP decrease was observed. This outcome was significantly influenced by the average temperature of glycol in the ground loop circuit (heat exchange efficiency of the evaporator). The course of the Tgs temperature changes is presented in Fig. 10. Fig. 7 shows the influence of water temperature differences at the inlet and outlet of the space heating system DTsh on the COP
Fig. 4. A chart of temperature changes in the underfloor circuit for a changing mass flow of water.
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Fig. 5. A chart of temperature changes in the underfloor circuit for 75 and 140 m2 of space heating.
value and heating capacity Qcon. The data given indicate that, for DTsh ¼ 18 C (ground floor, 0.09 kg/s), the lowest COP and Qcon values were obtained, whereas the highest values of the COP and Qcon were obtained for DTsh ¼ 8 C (ground floor and first floor, 0.22 kg/s). Therefore, maintaining the lowest possible temperature difference in the heating circuit should be aimed at. At a constant heating capacity of the heat pump under given work conditions, increasing the mass flow of water results in a decrease of the condensation temperature of the refrigerant and water at the outlet of the condenser. A lower condensation temperature causes COP and Qcon values to increase. 3.2. Domestic hot water heating Domestic hot water heating is possible after the switching of the heat pump to DHW mode. In this mode, the circulation pumps of the space heating are off, whereas the DHW pump is on. The
domestic hot water tank with a capacity of 250 L, has a significantly lower thermal inertia in comparison with the underfloor space heating, which results in greater dynamics of the temperature changes in the heating circuit. In Fig. 8 it is clearly visible that, at a constant cooling capacity of the heat pump, a temperature increase in the DHW tank induces an increase in the refrigerant condensation temperature and water temperature at the inlet and outlet of the tank circuit. A water temperature rise at the outlet of the condenser is proportional to a water temperature increase in the DHW tank and amounts to 70%. In one hour of operation of the heat pump, water in the DHW tank is heated from 20 C to 42 C with an average COP ¼ 3.28 for this measurement cycle. Fig. 9 presents the influence of water temperature in the tank on a COP as well as the current supplied to the compressor drive. A water temperature increase in the tank causes a condensation temperature increase of the R407C refrigerant, which consequently
Fig. 6. A chart of the COP depending on the evaporation temperature of the refrigerant.
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Fig. 7. The COP and Qcon depending on water temperature differences in the heating circuit.
induces increased current supplied to the compressor drive along with a decrease in the COP. At the condensation temperature of 50 C, the heat pump energy efficiency decreases to COP ¼ 2.7, whereas the current supplied to 6 A. It is a limit value for the compressor motor, therefore, obtaining a higher water temperature in the tank requires using a heater built in the tank. 3.3. Ground loop During the heating season, an increased demand for heat results in cooling of the ground around the vertical collectors, which manifests itself as a drop in the glycol temperature at the evaporator inlet. According to the trend presented in Fig. 10, the lowest temperature values in the glycol circuit were observed from December to March (the highest demand for thermal energy). Outside temperatures in January and February that were higher
than the statistical average, as well as a relatively low outside temperature in March delayed the temperature increase in the glycol circuit until April. At an average temperature of glycol of 3 C at the evaporator outlet, setting the evaporation temperature at 2.5 C does not guarantee suitable conditions for exchanging heat in the evaporator and results in lowering of the heat pump energy efficiency. The possibility of activation of the circulation pumps independently in the circuits of the vertical collectors, made it possible to carry out tests for cases in which one or two vertical collectors are not working. Based on the conducted studies it may be assessed how the performance of the ground loop changes in case one of the vertical collectors fails. Fig. 11 shows a trend in glycol temperature changes at the inlet and outlet of the evaporator when one, two and three vertical collectors are subsequently activated. For a better presentation of
Fig. 8. A chart of temperature changes in the DHW circuit.
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Fig. 9. The influence of water temperature in the tank on COP as well as current supplied and condensation temperature.
the temperature changes of the ground loop, the results of 3 consecutive measurement cycles were shown in one chart. For all three vertical collectors working simultaneously, an average temperature difference between the inlet and outlet of the evaporator of DTgs ¼ 2.7 C was recorded along with 23% glycol temperature drop at the evaporator inlet within a 30min measurement cycle. When two collectors were activated simultaneously, the recorded values where DTgs ¼ 3.3 C and 31% respectively. Switching off one collector from the circuit resulted in a DTgs increase (lower volume flow of glycol in the evaporator) and an increased glycol temperature drop rate during the measurement cycle. When only one collector was activated, the recorded values were DTgs ¼ 5.8 C and the drop of the glycol temperature 37%. In this case, after 20 min from activation, the glycol temperature at the evaporator outlet decreased to 0 C and the immobilizer of the controller activated (protection against excessive cooling of the ground around the collector). Fig. 12 shows the trend in the temperature changes in the glycol circuit for cycle 2 (working time e 120 min and break time e
60 min). From the analysis of the presented data it may be concluded that the adopted break time of the heat pump is sufficient to level the ground temperature around the vertical collectors. Additionally, the chart presents the course of the floor temperature changes when the heating circuit is activated on the ground floor. 3.4. Summary and results The research studies conducted in the 2012/2013 heating season indicate that the greatest demand for heat was recorded in December (Fig. 13). The outside temperature in January and February, higher than the statistical average, resulted in a significantly lower demand for heat in this period. Low outside temperature in March rendered the demand for heat in that month comparable to the demand observed in January. On average, the circulation pumps used 35 W per hour in the first gear and 50 W per hour in the second gear. The total power consumption of the circulation pumps constituted 3.5% of the compressor power consumption. Admittedly, setting the circulation pump in gear 2 increased the power consumption by 15 W, but
Fig. 10. A chart of temperature changes in the glycol circuit during the heating season.
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Fig. 11. A chart of temperature changes in the glycol circuit when loops 1, 2, and 3 are working.
an increased mass flow of water resulted in a decrease of DTsh as well as a decrease in the condensation temperature. It may be calculated from the recorded data that for the COP ¼ 4.18 the compressor uses 125 W less energy for the production of the same amount of heat than for the COP ¼ 3.95. Therefore, it is economically justified to set the circulation pumps to gear 2 in view of greater savings in the compressor power consumption. The COP increases alongside with an increase in the evaporation temperature of the R407C refrigerant. In the case of GSHP, however, the maximum COP value was obtained for the evaporation temperature about 2.5 C. The results obtained were significantly influenced by the temperature of glycol in the ground loop. This problem was described whilst analyzing the data presented in Fig. 3. The performance of the heat pump during the investigated period changed in the range 8.4 kWe9.2 kW. The highest value was obtained for DTsh ¼ 8 C (gear 2 of the heat pump, 140 m2). The lowest value of heating capacity was obtained for DTsh ¼ 18 C (gear
1 of the heat pump, 75 m2). Therefore, it may be assumed that setting higher values of the mass flow of water in the underfloor circuit induces a reduction of the condensation temperature, which ensures favorable conditions for the obtainment of higher COP and heating capacity values. Additionally, studies with the use of an inverter for the drive of the heat pump compressor were conducted. In the case when a hermetic spiral compressor was used, it was possible to carry out studies in the range of 40e60 Hz. The results presented in Fig. 14 indicate that increasing the frequency of the compressor drive causes an increase in the heating capacity, the COP, however lowers as a consequence. Therefore, the use of an inverter for the drive of the spiral compressor does not result in significant economic benefits. The control system of the heat pump should enable keeping the condensation temperature as low as possible thus maintain a high COP value.
Fig. 12. A chart of temperature changes in the glycol circuit when space heating is working on the ground floor.
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Fig. 13. A summary of power consumption and production of heating energy during particular months of the heating season.
Fig. 14. The influence of the compressor's propulsion frequency on the COP and heating capacity at the evaporation temperature of 2.5 C.
The heating capacity can be adjusted to match the heat demand for space heating by controlling the duration of compressor operation.
4. Conclusions The heating system with the ground source heat pump was successfully installed and investigated in a low energy building. The results obtained were presented and analyzed. The research results presented may lead to the following conclusions: 1. The control system of the heat pump should enable keeping the condensation temperature as low as possible thus maintain a high COP value. 2. The use of an inverter for the drive of the spiral compressor does not bring significant economic benefits.
3. Using the space heating system on the ground floor and on the first floor simultaneously (140 m2 of the floor surface) led to obtaining the highest COP and Qcon values. 4. The use of several vertical collectors in the installation of the ground loop makes it possible to operate the heat pump in case of failure in one of the collectors. 5. The results of operation of the compressor are in line with its specifications. 6. The configuration of the heat pump used in the heating system made it possible to obtain a high COP for space heating, whereas, it failed to meet the expectations as regards the heating of hot domestic water above 45 C. Acknowledgements The author would like to thank Professor Krzysztof Urbaniec from Warsaw University of Technology e Plock Campus for his useful comments on the preparation of this paper.
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Nomenclature
References
Cp E M Q T Wcom
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specific heat, [kJ/kg K] energy mass flow, [kg/s] heating capacity, [kW] temperature, [ C] compressor power input, [kW]
Abbreviations COP coefficient of performance DHW domestic hot water GSHP ground source heat pump HGSHP hybrid ground source heat pump RES renewable energy sources TRT thermal response test Subscripts com compressor con condenser ev evaporator fl floor gs ground source in inlet out outlet sh space heating ta tank w water 1gf pump 1st gear on the ground floor 1gff pump 1st gear on the ground and first floors 2gf pump 2nd gear on the ground floor 2gff pump 2nd gear on the ground and first floors