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/lp/ou_press/dual-source-heat-pump-a-high-efficiency-and-cost-effective-alternative-8lgyPhIrSY
- Publisher
- Oxford University Press
- Copyright
- © The Author(s) 2018. Published by Oxford University Press.
- ISSN
- 1748-1317
- eISSN
- 1748-1325
- DOI
- 10.1093/ijlct/cty008
- Publisher site
- See Article on Publisher Site
Abstract
This article presents the characteristics and performance of an innovative dual source heat pump (DSHP) for heating, cooling and domestic hot water (DHW) production. The research work was carried out in the frame- work of the H2020 European project: Geot€ch ‘GEOthermal Technology for economic Cooling and Heating’. TheDSHPis abletochoose the most favourable source/sink in such a way that it can work as an air-to-water heat pump using the air as a source/sink, or as a brine-to-water heat pump coupled to the ground. The DSHP is manufactured as an outdoor ‘plug & play’ unit, working with R32 refrigerant and including a vari- able speed compressor, which gives full capabilities for an efficient modulating operation. The DSHP was fully characterized in steady state conditions at the IUIIE laboratory. In order to assess its dynamic perform- ance and to identify key control strategies to optimize its annual operation, a complete integrated model of the DSHP system in TRNSYS including the DSHP and all the other system components was developed. A first energy assessment, carried out for an office building located in the Netherlands, proves that the DSHP system would be able to reach a similar efficiency than a pure ground source heat pump (GSHP) system with half the ground source heat exchanger area needed. Therefore, the DSHP system could become a cost- effective alternative solution for heating, cooling and DHW production in buildings, as the initial investment would be significantly reduced compared to GSHPs, with similar or even higher energy efficiency. Keywords: dual source heat pump; geothermal energy; energy efficiency Received 7 November 2017; revised 23 January 2018; editorial decision 9 February 2018; accepted 12 *Corresponding author: ancamar4@upvnet.upv.es February 2018 ......... ................. ................ ................. ................. ................ ................. ................. . ............... ................. ................. heating and cooling renewable technologies currently available. 1 INTRODUCTION These systems use the ground as a heat source or heat sink, According to the International Energy Agency, buildings account depending on the season, in order to provide buildings with heat- for almost one third of the final global energy consumption, and ing and cooling, respectively. However, they imply the use of they are an important source of CO emissions. In particular, refrigerants in the heat pump refrigeration cycle that might have heating, ventilation and air-conditioning systems (HVAC) an impact in the ozone layer depletion and global warming. account for roughly half of global energy consumption in build- Fortunately, the current trend is to switch to new refrigerants ings. The sector is expanding, so it is bound to increase its energy with no impact in the ozone layer and a low global warming consumption. Therefore, reduction of energy consumption and potential. Nowadays, the GSHPs that are in the market are work- the use of energy from renewable sources in the building sector ing with these type of refrigerants, such as HFCs or HFOs (e.g. constitute important vectors to reduce the greenhouse gas emis- R32). Regarding the direct and indirect emissions, the current sions. When it comes to space heating and cooling using shallow GSHPs are usually factory shield equipment, so the direct emis- geothermal energy as a renewable energy source, ground source sions of refrigerant are negligible and practically the totality heat pump (GSHP) systems become one of the most efficient of the refrigerant is recovered at the end of the heat pump life. International Journal of Low-Carbon Technologies 2018, 13, 161–176 © The Author(s) 2018. Published by Oxford University Press. This is an Open Access article distributed under the terms of the Creative Commons Attribution Non-Commercial License (http://creativecommons.org/licenses/ by-nc/4.0/), which permits non-commercial re-use, distribution, and reproduction in any medium, provided the original work is properly cited. For commercial re- use, please contact journals.permissions@oup.com doi:10.1093/ijlct/cty008 Advance Access Publication 2 March 2018 161 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 J.M. Corberán et al. Furthermore, as the power consumption of these systems is lower corresponding cost of the ground heat exchanger and at the same than conventional ones, the indirect emissions are also reduced. time will offer an interesting potential for energy reduction if an GSHP systems have proved to be more efficient than conven- adequate operation strategy is developed for making use of the tional air-to-water heat pumps, as demonstrated by Urchueguía most convenient heat source at each moment. et al.[1], who concluded that GSHP systems can lead up to a 40% The Geot€ch project ‘Geothermal Technology for economic savings in annual electricity consumption, in comparison to air to Cooling and Heating’ [9]is a 4years’ duration project (2015–18) water conventional heat pumps. Nevertheless, one of the main disad- which intends to demonstrate the next generation of GSHP sys- vantages of GSHPsistheir high investment cost. Therefore, a reduc- tems with a high energy efficiency but also with lower system costs tion in both construction and operation costs is required for these with respect to those already existing in the market. One of the systems to become successful, especially for Southern European aims of the project is to develop system solutions that make the countries where the market of GSHP systems has not taken off yet. best use of hybrid heat pump and control technologies so that effi- A possible approach to save energy in GSHP installations is to cient replicable ‘plug & play’ whole systems can be offered to the combine it with another thermal source in the form of hybrid sys- housing and small building market sectors. To this end, an effi- tems. In the case of heating dominated areas, they are combined cient and comparative low cost ‘plug & play’ system for providing with solar thermal energy as reported in Ref. [2]. A good review of the heating, cooling and DHW needs has been designed and it is this sources combination can be found in Ref. [3]. In the case of going to be installed in three demosites located in Italy, the United heating and cooling systems, a common practice is to combine Kingdom and the Netherlands, respectively. GSHPs with a cooling tower or a dry cooler, to use the ambient as An innovative DSHP has been developed which is capable of an extra heat source/sink. It is also possible to combine GSHPs making optimal use of ground or air environmental heat sources with thermal energy storage, mainly by means of phase change according to operating and climate conditions. The heat pump is materials, as described in Ref. [4]. In fact, both a hybrid system and able to select the most favourable source/sink (ground or air) a thermal storage device can be combined as described in Ref. [5]. depending on their temperatures or other sensed parameters. Hybrid systems, combining ground and air as heat sources, This article first describes the design and characteristics of the have two basic advantages. On one hand, given the cost of an air DSHP designed in the project. Then, a TRNSYS model of the com- heat exchanger (cooling tower or dry cooler) the ground heat plete system, including all the integrated system components (DSHP, exchanger can be significantly reduced, reducing the total cost of ground source heat exchanger, air conditioning and DHW hydraulic the system. On the other hand, if the operation of the system is loops), which has been developed in order to assist both in the opti- optimized, the possibility of employing the most adequate heat mal design and energy optimization of the operation of the system, is source could lead to a considerably higher seasonal performance presented for the demo site located in the Netherlands. Finally, the and consequently to the reduction of the energy consumption. paper presents an initial analysis of the system operation and per- In order to further reduce the size of the installation, its cost, formance based on the model results for a whole year of operation. and simplify the operation, some researchers tried to develop a dual A preliminary energy assessment of the system was presented source heat pump (DSHP), implementing the dual source: water in the 16th International Conference on Sustainable Energy and air, directly into the heat pump design. In Ref. [6], the authors Technologies—SET 2017 with title ‘Modelling and energy analysis developed a DSHP utilizing groundwater and air sources. They of a DSHP system in an office building’ [10]. This article describes found an improvement in the Performance Factor ranging from 2 the dual heat pump design and experimental performance, as well to 7% compared with the system only employing the groundwater as an updated version of the model including the details of the source, and from 4 to 18% when compared with the system only building thermal loads (heating, cooling and DHW), as well as an employing the air source. They pointed out that higher system per- initial energy assessment for the system. formance could be achievable with the use of variable speed for the compressor and the circulation pump. A recent review on hybrid 2 INNOVATIVE DSHP ground heat pumps can be found in Refs. [7]and [8]. The possibility of having the dual source available right at the The innovative concept of the dual source heat pump presented heat pump has several advantages: first the cost will be certainly here is the possibility of having both air and ground source/sink all lower than the one required for the external air source heat integrated in the same refrigerant circuit. This compact solution exchanger alternative (dry cooler or cooling tower), the unit will will allow using either the air or the ground source/sink, whichever be much more compact since it is integrated in the heat pump is more convenient from an efficiency point of view, therefore, lead- and it takes the advantage of a superior heat transfer on the ing to a superior seasonal performance compared with current tech- refrigerant side, and the elimination of the temperature difference nology, and to a significant reduction of the size of the ground heat produced by the intermediate BPHE, and for this same reason the exchanger with the corresponding reduction on the system cost. COP of the unit when operating with the air source, will be super- ior. All this at the marginal cost of adding an extra heat exchanger, and the necessary valves and control to switch from one source to 2.1 Design requirements and conditions another into the heat pump design. Furthermore, as other hybrid The heat pump has to satisfy all the heating, cooling and DHW options, the DSHP will significantly help to reduce the size and demand of a small multifamily house or office. It must be 162 International Journal of Low-Carbon Technologies 2018, 13, 161–176 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 Dual source heat pump, a high efficiency and cost-effective alternative reversible in order to produce heating during winter and cooling during summer. In addition, it has to satisfy the domestic hot water (DHW) needs. The unit has been conceived as a ‘plug & play’ unit for domestic use, therefore, it must be as much compact and simple to operate as possible, as well as extremely automatic and reliable. Nominal heating capacity should be around 8–10 kW. Anyhow, in order to have a unit that can be employed in a large number of different applications, also being able to adapt to a broad range of thermal demands, it was decided to employ a variable speed, inverter driven, scroll compressor. The targeted efficiency for yearly production of heating, cooling and DHW is an annual performance factor >3.5. DHW produc- tion is not a negligible thermal load, especially because of its rela- tively high temperature. In order to increase the efficiency of the unit, it was decided that the DHW will be generated by the HP at 55°C. Legionella treatment of the DHW tank will be done by an electrical resistance in the tank following regulations and recom- Figure 1. Basic structure and components of the dual source heat pump. mendations for this kind of small open loop systems. The decision about which should be the employed refrigerant the use of an air to water heat pump with an extra BPHE for the is very difficult, given the current status of the European F-gas ground loop, as the frame for the evolution to the DSHP. regulation and the volatile evolution of the regulations on refriger- Furthermore, in order to be able to produce DHW apart from ants internationally. On one hand, it seems clear that the use of heating water, a dedicated BPHE for DHW production was incor- HFCs will be severely and progressively restricted. On the other porated in the design. Figure 1 shows the basic structure and hand, the alternatives to the current employed refrigerants are not components. clearatall,and formanyofthe optionsthere arenot commer- cially available components yet, e.g. compressors. The possible choices of refrigerant at the moment of designing 2.2 System layout the unit were: The system layout is designed to satisfy all the operating modes of the heat pump, i.e. heating, cooling and DHW production both – R410A: This is the standard refrigerant for this kind of applica- during winter and during summer. The targeted market is houses tion. It has good characteristics but a high value of GWP, there- or small offices, therefore, heating and cooling are completely fore, it does not seem to be in line with the trend indicated in separated by the intermediate seasons. Only DHW demand can the new EU F-gas regulation (No. 517/2014), thus not being exist simultaneously with either heating or cooling demands. usable in amediumtermhorizon. The DSHP must work in nine different operating modes, – R32: There is still a very limited compressor availability, but it which are summarized in Table 1. They are classified in winter seems that a large part of the AC OEMs are migrating to this and summer mode: when the system operates in summer mode, it old refrigerant because, despite being an HFC, it has consider- will work as a chiller while when it operates in winter mode, it will able lower GWP than the standard refrigerant of the sector work as a heat pump. Apart from the working modes correspond- (R410A). HP cycles with R32 tend to produce an excessively ing to the heat pump, Table 1 also shows three extra working high discharge temperature at the compressor which could be modes for the system (MS, M10 and M11) which correspond to limiting the compressor life. This problem can be solved by free-cooling operating conditions. The heat pump will be switched sequentially injecting some refrigerant liquid at the suction, in offinmodeMSand M10, but itwillbe switchedoninM11 for order to cool down the compressor. the production of DHW using the air as a source. Working mode – R1234ze: This HFO refrigerant is also an alternative with almost MS—Midseason indicated in Table 1 corresponds to those negligible GWP. However, it is an expensive refrigerant, and moments of the year in which the external ambient temperature very new in HP applications, therefore, there is a very little gets very mild values in the range, for instance, of 21 ± 1.5°Cin availability of components. Moreover, there is still some contro- cooling mode, and 19 ± 1.5°C in heating mode. In these condi- versy about its long-term behaviour, so it could be facing also tions, no active heating or cooling is needed and therefore the heat restrictions in future regulations. pumpwillbeswitchedoff. A great number of refrigerant circuit layouts can be considered After the consideration of all the possible alternatives and the con- in order to keep the right circulation of refrigerant through the sultation with several compressor manufacturers, finally R32 was desired components. There are clearly six different topological chosen as the best solution for short and medium term. The use of operation modes, allowing to produce heating, cooling or DHW, R32 pushes the option of an outdoor installed unit. This has led to in winter and summer, from the two available source/sinks air and International Journal of Low-Carbon Technologies 2018, 13, 161–176 163 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 J.M. Corberán et al. Table 1. Heat pump and system operating modes. the smallest variable speed compressor available for R32. The selected compressor is model XHV-25 (R32) by Copeland. Condenser Evaporator Operating mode As explained above, the adopted DSHP concept is based on an Summer Air User M1—Summer Air air to water heat pump. The frame dimensions of the heat pump Ground User M2—Summer Ground prototype were selected in order to have space enough to fit all the –– M10—Free Cooling necessary components. The air coil consisted of a Round Tube –– MS—Midseason and Plate Fin (RTPF) heat exchanger occupying a whole side of DHW User M3—DHW User Air M6—DHW Air theframe.Onthe othersideofthe coil thefansweresituated. The Ground M8—DHW Ground RTPF consisted of two rows of tubes of 8 mm diameter, with five Air M11—Free-Cooling + DHW Air refrigerant circuits. Two ECM fans (A3G450-AC28-51 by EBM) Winter User Air M4—Winter Air were selected with continuous variable speed control. –– MS—Midseason Both BPHEs (USER and GROUND) are employed either as User Ground M5—Winter Ground DHW Air M7—DHW Air evaporator and condenser depending on the operation mode, Ground M9—DHW Ground therefore, they have been designed both as evaporators and both include the corresponding distributor. The plate models finally selected were the F80AS (asymmetric plate) and the F85 by ground. There is an additional mode when it is possible to pro- SWEP. duce DHW in summer at the same time that the unit is producing The BPHE for the DHW production works always as a con- chilled water (this mode is called full recovery). denser, so a condenser model has been selected for this service: The components which must be interconnected in the right plate model B26 (asymmetric) by SWEP. order, depending on the operation mode, are the inverter driven Asymmetric plate models by SWEP have been selected in order compressor, three BPHEs (USER, providing heating or cooling to to minimize the refrigerant charge and reduce the pressure drop the building; GROUND, connecting the unit with the source/sink on the water loops. ground loop; DHW), and finally one air coil (round tube plate fin Electronic expansion valve (EEV) Model E2V14 by CAREL heat exchanger—RTPFHX), which can actuate as a condenser or was employed to keep control of the superheat. As the unit has a evaporator depending on the mode. Additionally, the unit requires large number of different operation modes and the evaporator an adequate liquid receiver to store all liquid refrigerant that will changes from one heat exchanger to another, the control of the be in excess for some of the modes and operating conditions, and superheat is metered at the compressor suction, which is common at least one expansion valve to control superheat and, in some for all operation modes. For the same reason there is no way to occasions, limiting the maximum discharge temperature by inject- implement pressure compensation. Anyhow, the pressure drop ing some liquid refrigerant at the compressor suction. through the different evaporators has been estimated to be low. A significant number of solutions for the refrigerant circuit ACAREL controlsystem(pCO5+) is employed to control the have been analysed and evaluated from the point of view of cost, EEV, the switching of the solenoid valves, the speed of the com- reliability and efficiency, in that same order. The best solution pressor, fans, and the water and brine circulation pumps, and all found is based on the utilization of 10 solenoid valves, which the required safety switches and alarms. Additionally, the control always work on the adequate direction by combining them if system measures the discharge temperature and if this tempera- necessary with a check valve, a unique liquid line, with liquid ture increases over a certain threshold it controls the EEV in a receiver and sight glass, and one direction expansion valve for the way that some pulses of liquid refrigerant are sent to the compres- control of the superheat at the inlet of the compressor. sor suction in order to cool down the compressor. The selection of the adequate diameters for the pipes of the dif- ferent lines of the unit has been based first on two main targets: 2.3 Design and selection of components first, avoiding excessive pressure losses; and second, insuring a This unit is going to incorporate a variable speed compressor in minimum velocity in order to get the return of the oil at low com- ordertobeabletoadapt theperformance to thewiderange of pressor velocities. This is especially critical for the suction line. For operating conditions and to minimize part load losses. Scroll tech- thosepiecesofpipe thatchangetheir role from onemodeto nology has been chosen because its superior efficiency at these another, for instance, from part of the liquid line to part of an operating conditions. expansion line, the most critical criteria has been selected in order Theoptimum compressor sizewould be thesmallest possible to satisfy the design conditions. one that could provide the demand in the worst conditions run- The three required circulation pumps are also included in the ning at maximum speed. In this way, it would be able to cope with frame of the heat pump: ground pump, user pump and DHW the peak demand but still have a good system performance for pump. most of the time in part load and a wide modulation capacity. A modelling study was performed by means of IMST-ART However, the commercial availability of compressors for the cho- [11] in order assist the design of the heat pump, for each of the sen refrigerant (R32) was very small at the moment of designing different operating modes. IMST-ART has been very useful to size the heat pump. Among the available compressors, it was chosen the BPHEs, to design the coil circuitry and all the pipes of the 164 International Journal of Low-Carbon Technologies 2018, 13, 161–176 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 Dual source heat pump, a high efficiency and cost-effective alternative refrigerant circuitry. Besides, it has provided an initial estimation brine temperature to the evaporator, brine flow rate (or brine tem- of the heat pump performance. perature variation across the evaporator), inlet water temperature Figure 2 shows the basic layout and composition of the heat to the usercondenser,userloopmassflowrate(or watertempera- pump in itsframe.Asit can be seen,the aircoilis situatedinthe ture variation across the condenser) and compressor frequency. front, and the fans in the back panel. The three BPHEs, the com- For air source modes, the brine temperature and the brine flow pressor, the liquid receiver and the three circulation pumps are rate, are not employed and instead, the input variables are the out- visible in the picture. It should be pointed out the clean layout of door air temperature and the air flow rate (or the fan frequency). the heat pump and the compactness of the unit. IMST-ART was employed to evaluate the unit performance in the whole range of variation for Winter-ground heating mode, totalling 640 test runs. The performance results, mainly condenser 2.4 Testing campaign and unit performance capacity, evaporator capacity and compressor consumption for all A first set of 29 steady test points were performed, including the 640 points, were employed in order to find a convenient polyno- most important operation modes and conditions, which was mial, containing linear and quadratic terms, and also some crossed employed to analyse the unit performance and prepare the general variables terms. The study was performed employing the fitting by testing campaign. These tests allowed to check the adequate linear regression. The study was carried out manually by inserting refrigerant charge and check the effective oil return and have a new terms and retaining those that have obtained good estimation first estimation of the unit performance. indicators until no further significant improvement was reached Additionally, with these results, the compressor efficiencies in the adjusted R and the maximum relative error. The finally were evaluated and it was found a slight decrease of efficiency selected polynomials were able to predict all the 640 performance from the estimated values obtained from the R410A results sup- points with a maximum error lower than 4%, and a much lower plied by the compressor manufacturer. The compressor efficiency average error. correlations were conveniently readjusted to the experimental Then, taking these performance polynomials as the true data and they were introduced in the IMST-ART software. Then Response Surface (RS) a Design Of Experiments (DOE) was a comparison between predicted and measured results was per- carried out. Several DOE methodologies were tried and tested, formed for all the available test points. A very good agreement in order to find which one was able to give the best compromise was found between the predicted and the measured results, so it between the number of test points and accuracy in the determin- was decided to employ the software to explore the variability of ation of the corresponding RS. The best compromise was found the unit performance when the input variables were changed, with the Central Compact Design methodology and it turned out covering the whole range of possible variation for them. For that selecting only one of the two orthogonal blocks and central instance, for winter-ground mode, the input variables are inlet star it was sufficient to get a very good estimation of the RS all over the entire domain of the five independent variables, resulting in only 30 test points. Once the test matrices were elaborated for each operating mode, the test campaign was followed until all the necessary points were tested for all the seven operating modes. The test matrices were adequately corrected when the variation of some parameter could go beyond the testing capabilities or the test point was in an area of no interest for the targeted application, for instance when the unit will not be in operation because it is much more profitable to employ the free cooling system. All test results were systematically analysed in order to detect possible operation problems or mistakes, and be able to repeat them if necessary. After all the test results became available, the compressor correlation was checked again and readjusted, and the performance for each test point was also evaluated with the soft- ware. This allowed a basis to analyse the results and to check the unit performance and the possible existence of testing mistakes. A comparison between experimental and predicted performance was carried out for each operating mode. This comparison allowed to deeply analyse the results and check all the individual test points. A very good agreement between estimated and measured results was found all across the different test matrices for all the seven operating modes. Figure 3 shows the heating capacity, the compressor power Figure 2. Dual source heat pump plug & play unit. input, and the COP (calculated only taking into account the International Journal of Low-Carbon Technologies 2018, 13, 161–176 165 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 J.M. Corberán et al. compressor consumption) for Winter Ground operating mode. the fact that the air temperature is much higher than the return The results are shown as a function of the compressor fre- brine temperature, so the performance between both working quency, for a hot water supply temperature of 45°C, 0°Cof modes cannot be fairly compared. In any case, the high values of return temperature at the brine loop, a 3 K brine temperature air operation at moderate temperatures 7(6) supports the con- difference through the evaporator and 5 K through the conden- cept of taking the heat from the air when the air temperature is ser. Figure 4 shows the same results for a hot water supply tem- high or mild, reserving the ground for the lowest ambient tem- perature of 35°C. peratures. Again, the maximum COP is obtained around 40 Hz As it can be observed, both the heating capacity and the com- of compressor speed. pressor consumption increase with the compressor speed in a Once the results for each operating mode became available, quite linear way. COP values are considerably high regardless the they were employed to fit the developed performance polynomials, low temperature considered at the inlet of the evaporator, 0°C which will be afterwards employed for the system models and for (brine return temperature from the ground heat exchanger). The the optimization of the system control. maximum COP is obtained around 40 Hz for both applications, with a significant performance decay. Figures 5 and 6 show the same performance, in this case for 2.5 DSHP model Winter Air operating mode with the air at a dry bulb temperature The heat pump will be considered as a black box in the TRNSYS of 7°C (wet bulb temperature of 6°C), keeping the rest of the para- integrated model. For that purpose, the performance of the unit meters (Tco and dTc) with the same values as in Figures 3 and 4. will be calculated by means of polynomial correlations which As it can be observed, the heat pump COP is higher in depend on the working conditions (different source and distribu- Winter Air than in Winter Ground. However, this is only due to tion temperatures and water/brine flow rates) of the heat pump Figure 3. Condenser capacity, compressor power input and compressor COP Figure 5. Condenser capacity, compressor power input and compressor in Winter Ground operating mode (Tco = 45°C, dTc = 5K, Tei = 0°C, COP in Winter Air operating mode (Tco = 45°C, dTc = 5K, T = 7(6), air dTe = 3K). f = 50%). fan Figure 4. Condenser capacity, compressor power input and compressor COP in Figure 6. Condenser capacity, compressor power input and compressor COP in Winter Ground operating mode (Tco = 35°C, dTc = 5K, Tei = 0, dTe = 3K). Winter Air operating mode (Tco = 35°C, dTc = 5K, T = 7(6), f = 50%). air fan 166 International Journal of Low-Carbon Technologies 2018, 13, 161–176 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 Dual source heat pump, a high efficiency and cost-effective alternative for each operating mode. These polynomial correlations were Q = f ⋅(C + C ⋅+ Te C ⋅ Tc + C ⋅⋅ Te Tc evap 0 123 ii i i comp obtained from the experimental results for the condenser capacity, + C ⋅ Te +⋅ C dTe+⋅ C dTc 4 i 56 evaporator capacity and compressor power input, and allow a full +⋅ C Te ⋅ dTe+ C ⋅ Tc ⋅ dTc) (2) characterization of the heat pump performance among all the 78ii operation modes and a large range of operating conditions. An 2 2 example of the employed polynomial correlations is presented in Wf=⋅(C+C⋅Te +C⋅Tc +C⋅dTe comp 01 ii 2 3 comp the following. +⋅ C dTc+ C ⋅ Te ⋅ dTe+ C ⋅ f ⋅ Te 45 i 6 i comp These polynomial correlations were implemented as a new TRNSYS type and integrated in the system model. The main rea- +⋅ C f ⋅ Tc ++C ⋅ f ⋅ Te ⋅ Tc 78 i ii comp comp son for this instead of using the already available heat pump +⋅Cf ⋅dTe+C ⋅ ) (3) TRNSYS types was the need of accurately reproducing the heat comp comp pump performance under different working conditions, which ̇ ̇ depends not only on the source and load inlet temperature but Where, Q is the condenser capacity (W), Q is the evapor- cond evap also on many other variables, as previously stated, for each 1 of ator capacity (W), and W is the compressor power input comp the 11 operation modes. (W). As it can be seen, the polynomials for the heat pump cap- On the other hand, a disadvantage of the already available heat acities are based on nine coefficients, while the compressor pump types in TRNSYS, is that they calculate the performance of power input requires two extra coefficients, i.e. total 11 coeffi- the heat pump based on a fixed number of working points and cients. The compressor power input includes the inverter losses. interpolate the performance when operating under other different When the heat pump is working with the air, for instance conditions. Furthermore, they calculate the heat rejected or Winter Air mode, the same polynomials were found to be absorbed in the source side (ground or air) as the sum (if it is in adequate with the substitution of dTe, which is related to the brine cooling mode) or the subtraction (if it is in heating mode) of the flow rate in the ground mode, by the fan frequency f (%). fan heat pump capacity and the power consumption, not considering Figure 8 shows the comparison between experimental measure- the thermal losses in the heat pump cycle. Contrary to this, the ments and the correlations results for the compressor consump- heat pump model based on the correlations calculates the conden- tion and the condenser capacity in Winter Ground operating ser and evaporator capacity, the power consumption and parasitic mode (M5). It can be observed that the adjustment is very accur- losses separately, based on the experimental results. ate, being the maximum deviation lower than 3.2 and 2.6%, Figure 7 shows a schematic diagram of the heat pump as a respectively. The maximum deviation between experimental and black box model for Winter Ground operating mode. The heat calculated performance values for all operation modes is lower pump model receives the following inputs: the secondary fluid than 5%. inlet temperature to the evaporator, Te (K) (brine in this case); Figures 9 and 10 show the COP maps obtained from the the secondary fluid temperature difference across the evaporator, developed polynomials for the Winter Ground and Winter Air dTe (K); the secondary fluid inlet temperature to the condenser, operating modes. Top graphs correspond to COP evaluated only Tc (K) (user loop in this case); the compressor frequency, f with the compressor consumption whereas bottom graphs cor- comp (Hz) and the secondary fluid temperature difference across the respond to COP evaluated with all electric consumptions, i.e. condenser, dTc (K). compressor, user circulation pump, and either brine circulation Equations (1)–(3) provide the heat pump performance for the pump or fan. Left graphs correspond to a supply water tempera- Winter Ground operating mode (M5). ture of 45°C whereas right hand graphs correspond to 35°C. The return brine temperature is 0°C for the Winter Ground case, whereas the air temperature is 7 (6)°C. The compressor fre- Q = f ⋅(C + C ⋅+ Te C ⋅ Tc + C ⋅⋅ Te Tc cond 0 123 ii i i comp quency is 50 Hz. It should be noticed that the Y axis in Figure 9 + C ⋅ Te +⋅ C dTe+⋅ C dTc 4 i 56 shows an opposite trend to the Y axis in Figure 10, since increas- +⋅ C Te ⋅ dTe+ C ⋅ Tc ⋅ dTc) (1) 78ii ing the brine temperature difference across the evaporator in Figure 9 means decreasing the brine flow rate, whereas Y axis in Figure 10 shows the fan frequency, which is proportionally linked to the air flow rate. As it can be observed in Figure 9, the lower the temperature difference across the evaporator, the higher the compressor COP. In other words, the higher the brine flow rate, the higher the compressor COP. This trend is corrected when it is taken into account the brine circulation losses in the brine loop, since the pumping losses depend on the brine flow rate. The result is that the optimum is situated around 3 K of temperature differ- ence. Interestingly, this is the standard value employed by most Figure 7. Black box model for the heat pump in Winter Ground mode. of the manufacturers for this parameter, moreover, it is the one International Journal of Low-Carbon Technologies 2018, 13, 161–176 167 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 J.M. Corberán et al. Figure 8. Experimental measurements versus polynomial correlations for Winter-ground mode: (a) Compressor consumption. (b) Condenser capacity. Figure 9. COP maps as a function of the water temperature difference across the condenser (dTc) and of the water temperature difference across the evapor- ator (dTe) at different supply water temperatures (Tco = 45°C on the left side, Tco = 35°C on the right side) in Winter Ground operating mode, at 50 Hz of compressor frequency and 0°C brine return temperature. employed in the certification testing standards. In contrast, the so that the user return temperature decreases, allowing a decrease maps show that the higher the water temperature difference in the condensing temperature that makes the COP increase. through the condenser, the higher the COP, although the influ- Similartrendscan be foundin Figure 10: the higher the water ence is lower than at the evaporator. This means that the lower temperaturedifferencethrough thecondenser (dTc), thehigher the user flow rate, the higher is the COP. Low flow rates imply a the COP, although the influence is again lower than the one of the higher thermal resistance at the condenser so it should contrib- fan frequency at the evaporator. Again, the lower the user flow ute to increase the condensing pressure and so to a COP rate, the higher is the COP. In this case, it is much apparent the decrease, which is not observed. This is because the mapping has influence of the fan consumption on the global COP. If the pumps been built with experimental results in which we keep constant and fan consumptions are not taken into account, the higher the the supply water temperature, which is the usual way of control fan frequency, the higher the COP (compressor) because the ther- of these kind of heat pumps. This makes that, in fact, when the mal resistance at the evaporator decreases and therefore the evap- condenser flow rate decreases, the temperature difference increases oration temperature increases. However, if the fan consumption is 168 International Journal of Low-Carbon Technologies 2018, 13, 161–176 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 Dual source heat pump, a high efficiency and cost-effective alternative Figure 10. COP maps as a function of fan frequency and water temperature difference across the condenser (dTc) at different supply water temperatures (Tco = 45°C on the left side, Tco = 35°C on the right side) in Winter Air operating mode, at 50 Hz of compressor frequency and 7(6)°C air temperature. taken into account in the COP (global), then, as the fan consump- is 200 l, but the demand of DHW in the office is 22.5 l/day, corre- tion increases with the fan speed, it reverses the influence and sponding to the demand of three people. This means that the makesthatthe lowerthe fanfrequency, the higher the global COP. DHW tank has a high inertia, and only a small part of the hot After analysing Figures 9 and 10, it can be concluded that water will be extracted during the day, so the heat pump will the influence of the operating parameters on the system per- work in DHW mode only during short periods in the night. On formance is very important. Therefore, the minimization of the the user side, the maximum heating load during the year is energy consumption of these complex systems requires the around 12 kW and the maximum cooling load is around 6 kW. optimization of its operation. The model described in the fol- Regarding the circulation pumps, the user loop circulation pump lowing is a tool to assist the optimization of the sizing and oper- is continuously working during the office schedule (from 08:00 to ation of these type of systems. 18:00), while the ground loop and DHW loop circulation pumps cycle with the compressor operation (they only work when the compressor is working). 3 INTEGRATED SYSTEM MODEL IN 3.1 Innovative coaxial BHE TRNSYS The BHE used in the system model is an innovative coaxial- In order to assist both in the optimal design and energy optimiza- spiral BHE that was developed by Geothex BV (http://geothex.nl; tion of the operation of the system, a model of the ‘plug&play’ 20 February 2018, date last accessed). The main innovations system in TRNSYS including all the integrated system compo- consist of an insulated inner pipe that reduces the heat trans- nents (DSHP, innovative ground source heat exchanger, air con- fer between the inner and outer pipe, together with a ribbed ditioning and DHW hydraulic loops) is presented for the demo outer channel, which makes the fluid follow a spiral path along site located in the Netherlands. The layout of the system model is the outer pipe. According to preliminary investigations, it is shown in Figure 11. possible to obtain a significant increase on the efficiency when The heat pump is able to provide heating and cooling to the compared to conventional BHEs, especially at low Reynolds user, but also DHW. In order to combine the different working numbers [12]. This BHE is under development and optimization modes, it is set that the heat pump will be able to provide DHW inside the framework of the GEOT€CH project. during the night (from 00:00 to 06:00). On the other hand, it will A model of this new coaxial-spiral BHE was developed based only provide heating or cooling during the opening schedule of on the thermal network approach, combined with a vertical dis- the building (from 08:00 to 18:00). The volume of the DHW tank cretization. This model has been adapted to the new configuration International Journal of Low-Carbon Technologies 2018, 13, 161–176 169 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 J.M. Corberán et al. from the B2G dynamic model previously developed for a U-tube loop instead of the ground temperature, as it is easier to measure BHE configuration and presented in Refs. [13]and [14]. The mod- the temperature of the fluid inside the pipe rather than measuring el reproduces the short-term behaviour of the BHE with a high the surrounding ground temperature. accuracy, taking into account only the portion of surrounding ground directly affected by the heat injected/extracted during the 3.3 Temperature compensation period considered. A detailed explanation of the model can be With the aim of increasing the global efficiency of the system, an found in Ref. [15]. optimization strategy based on the supply temperature compen- sation was implemented in the model. This strategy mainly con- sists of adapting the supply temperature to the evolution of the 3.2 Selection of the source/sink outdoor ambient temperature. For example, if the heat pump is It was previously stated that the heat pump is able to work using the providing heating and the outdoor temperature increases, the ground or the air as source/sink. In this context, a control strategy is supply temperature will be decreased, since it will mean a lower needed in order to select the source/sink that, in each moment, will heating demand and therefore there will be no need of supplying be most favourable in order to obtain the highest efficiency. the water so hot. Analogously, if the outdoor temperature As a preliminary control strategy, it was selected a simple strat- decreases, the supply temperature provided by the system will be egy, in which the source with the most favourable temperature increased, since it will mean a higher heating demand. This (highest temperature in heating mode and lowest temperature in allows taking advantage of milder outdoor ambient temperatures winter mode) is selected. In order to prevent the heat pump from to improve the coefficient of performance (COP) of the heat changing from one source to another in a short period (as the air pump whenever it is possible. The methodology used in this temperature changes with a high frequency), a differential control- model is detailed in [16]. According to this methodology, the ler isused, providingsomehysteresistothe control. This means supply temperature is calculated each simulation time step in that, the actual ground temperature is used as the reference tem- such a way that the terminal units are able to meet the user com- perature, and the heat pump changes the source to air when the fort even in the most extreme ambient conditions (maximum air temperature is more favourable (considering the dead band of and minimum ambient temperatures for the location: 29.3 and the differential controller). Figure 12 shows the operation of this −8.1°C, respectively for Amsterdam TRNSYS-Meteonorm wea- control for (a) heating mode and (b) cooling mode. ther file). Therefore, the supply temperature will be set depend- In order to programme this control easily in a control board, it ing on the following parameters: current ambient temperature, is planned to use the fluid return temperature from the ground the maximum and minimum annual ambient temperatures, the desired comfort temperature inside the building (21°C for heat- ing mode and 22°C for cooling mode) and the maximum and minimum limits in the supply temperature. These limits are fixed in order to avoid operating problems in the heat pump (freezing risk, very low pressure ratios in the compressor or very low COPs). In heating mode, the supply temperature limits that have been set are 35 and 45°C; in cooling mode, the supply tempera- ture will have a lower limit of 7°C and a higher limit of 15°C. TT =(14 +ββ ) · − ·T ( ) SB room amb TT − room SB)min β() cooling= ; TT − amb)max room () 5 TT − SB)max room β() heating= Figure 11. Integrated system model in TRNSYS layout. TT − room amb)min Figure 12. Selection of the source depending on the air and ground temperature. (a) Heating mode. (b) Cooling mode. 170 International Journal of Low-Carbon Technologies 2018, 13, 161–176 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 Dual source heat pump, a high efficiency and cost-effective alternative Where, T is the supply temperature to the building, T is calculated using a TRNSYS model ofthisbuildingand then,intro- SB room the comfort temperature desired in each of the air-conditioned duced in the global model of the system as an input. It is assumed spaces, T is the ambient temperature, and T and T that in the building there are three people working in average. amb amb)max amb)min are the maximum and minimum outdoor annual ambient tempera- The model considers the different walls of the building and tures, and T and T are the minimum and maximum windows, as well as the materials, infiltrations and ventilation SB)min SB)max supply temperatures for cooling and heating mode, respectively. required in the building. So, the sensible and latent gains that are needed to meet the comfort inside the different rooms are cal- culated based on the external conditions (outdoor temperature, 4 CASE STUDY AND ANALYSIS OF THE radiation on the different facades and humidity), the internal conditions (indoor temperature, humidity) and the overall heat SYSTEM ENERGY PERFORMANCE transfer coefficient calculated from the building envelope charac- teristics. In order to simplify the load profile and the operation of The model developed in TRNSYS was used to make an assessment the heat pump, two seasons have been considered: one heating of the system energy performance. In order to study the techno- season in which the heat pump provides DHW and heating (ther- economical feasibility of the DSHP system, two scenarios were mal load >0in Figure 13) and one cooling season, in which the compared. In the first scenario, a GSHP (only ground) system was heat pump provides DHW and cooling (load <0in Figure 13). simulated with a borehole field formed by three BHEs, 70 m deep The maximum peak loads are around 11.7 kW in heating and each one. In the second scenario, a DSHP (air/ground) system was around 5.4 kW in cooling. studied. In this case, the length of BHEs is half the one of the Regarding the DHW demand, a profile for DHW in an office GSHP scenario (three BHEs with a depth of 35 m each one). The building extracted from Ref. [17] and an occupancy of three peo- aim of thiscomparisonisto analyse whetherit ispossible to ple is considered. In this profile, the demand (litres per hour) is obtain a similar efficiency of the system but considering half the given for each hour of the day. Figure 14 represents this DHW length of BHEs, which would mean a significantly reduced invest- demand profile during one day. ment cost for the DSHP installation compared to that of a GSHP system. The main design and operation parameters of the system considered in the model are shown in Table 2. 4.2 Seasonal performance factor of the system In order to assess the energy performance of the systems, the sea- sonal performance factor (SPF) of the systems for a whole year of 4.1 Small office building in The Netherlands operation is quantified. The expressions used for the SPFs of the The building that has been used in this analysis is a small office build- systems (equations from (6)to(9)) were defined according to the ing located in the city of Amsterdam, The Netherlands. This building SEPEMO-build ‘SEasonal PErformance factor and MOnitoring has two zones with necessities of heating and cooling: an office room for heat pump systems in the building sector (SEPEMO-Build)’ and a meeting room. The thermal demand loads have been project definition [18]. This project aims at overcoming market barriers to a wider application of heat pumps by developing a uni- Table 2. Parameters of the system. versal methodology for field measurement of heat pump systems Set point heating 35–45°C SPF including a monitoring programme for 46 heat pump installa- Set point cooling 7–15°C tions in six European countries. These expressions are presented Set point DHW 55°C in the following equations. Set point free cooling 10°C Heat pump minimum frequency 30 Hz Heat pump maximum frequency 70 Hz ̇ ̇ ∫ (QQ + )·dt USER DHW Natural heating (Midseason) On: T > 20.5; off: T < 17.5 amb amb Natural cooling (Midseason) Off: T > 22.5; on: T < 19.5 amb amb SPF1 = () 6 Air/ground source, heating mode: Ground: <−2; AIR > 2 (Wd )· t (T – T ) HP air ground Air/ground source, cooling mode: Ground: <−2; AIR > 2 0 (T – T ) ground air Office Schedule (air conditioning 8:00–18:00 on weekdays ̇ ̇ ∫ (QQ + )·dt USER DHW schedule) DHW Schedule 0:00–8:00 and 20:00–24:00 on SPF2 = () 7 weekdays ̇ ̇ ̇ (WW + +W )·dt HP FAN BHE User buffer tank volume 55 l DHW tank volume 200 l DHW profile 22.5 l/day (profile for an office of three people) ̇ ̇ ∫ (QQ + )·dt USER DHW BHEs field Dual source system Three BHEs, 35 m deep BHEs field Ground source system Three BHEs, 70 m deep SPF3 = () 8 Thermal demand load profile Amsterdam ̇ ̇ ̇ ̇ ∫ (WW + +W +W )·dt HP FAN BHE BACKUP Simulation time step 60 s International Journal of Low-Carbon Technologies 2018, 13, 161–176 171 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 J.M. Corberán et al. Figure 13. Thermal demand load profile. Heating loads (>0) and cooling loads (<0). ̇ ̇ BHE circulation pump W , user circulation pump W and BHE USER DHW circulation pump W ), including the back-up system DHW in case that there is any W . BACKUP 4.3 Comparison of the two scenarios One of the main advantages of a DSHP system (air/ground) against a GSHP system (only ground) mainly consists of a reduction in the BHE required length thanks to a lower use of the ground. This means that the investment of the installation will be signifi- cantly reduced, as the BHE field is one of the most expensive parts in a GSHP system. In order to analyse this advantage, two scen- arios were considered using the TRNSYS system model developed. In the first scenario, a GSHP system is analysed. Therefore, the heat pump will work using only the ground as a source/sink. In the second scenario (DSHP system), the heat pump will work usingeitherthe airorthe ground as asource/sink,depending on Figure 14. Domestic Hot Water demand profile for an office building per the temperature of each one as it was detailed in Section 3.2. person. It should be noted that the BHEs field used in the DSHP sys- tem is half the size of the one used in the ground source system (three BHEs with a depth of 70 m in the ground source system SPF4 and three BHEs of 35 m in the dual source system). The efficiency of each system is assessed by calculating the SPFs. As there is no ̇ ̇ (QQ + )·dt USER DHW back-up heater considered in none of the systems, SPF and SPF 0 2 3 = () 9 will get the same values. Figure 15 presents the results for the ̇ ̇ ̇ ̇ ̇ ̇ ∫ (W +WWW + + +W +W )·dt HP FAN BHE BACKUP USER DHW assessment of the systems energy performance over a whole year of operation. where, Q is the useful heat in the user loop (Q ) or DHW As it can be observed, the resulting SPFs are quite similar in USER loop (Q ), and W is the power consumption of each of the both cases, with a yearly SPF of 3.58. So, it can be concluded that DHW 4 ̇ ̇ components existing in the system (heat pump W , fan W , both systems are able to work with a similar efficiency, this means HP FAN 172 International Journal of Low-Carbon Technologies 2018, 13, 161–176 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 Dual source heat pump, a high efficiency and cost-effective alternative Figure 15. Seasonal performance factors for a whole year of operation in the two scenarios: (a) ground source heat pump system (first scenario) and (b)dual source heat pump system (second scenario). that they can provide the same amount of energy to the system happens because during the operation in these modes, the heat with the same power consumption. However, as the DSHP system pumpis switchedoff andthere is no consumptionofthe compres- needs half the length of BHEs compared to that of the GSHP sys- sor, only the consumption of the ground and user loops circula- tem, the investment required will be lower and therefore the pay- tion pumps in the case of the free-cooling and the parasitic losses back time for the DSHP system. of the heat pump in both modes. In this studied case, the reduction in the investment cost could Figure 16 shows the amount of time in which the DSHP works reach values up to 30%. Since, the BHE field length would be in each mode. It is possible to check that most of the year it is reduced to the half (although it is also necessary to pay the cost of working in heating mode (mostly using the ground as a source). the drilling equipment), but the cost of the heat pump would be During the summer, the heat pump works just a few hours in around a 20% higher. The increase in the heat pump cost is due to cooling mode, being the free-cooling mode the most used as the fact that it is a prototype, which includes a non-conventional expected. The rest of the cooling demand is met by natural ventila- compressor and a more complex thermo-hydraulic system tion and the inertia of the building (midseason mode). (including a dedicated BPHE for DHW and two heat exchangers for the use of ground or air as a source/sink). Focusing on the 4.4 Dynamic performance during 1 week (dual DSHP system (Figure 15b) shows that the SPF values are quite higher during the summer, getting values around 5.28, whereas in source system) winter it is a 30% lower, getting values around 3.69. Then, as the This section presents a preliminary analysis of the system energy system is mainly operating in heating mode during most part of performance during 1 week of operation in order to describe the theyear(showninFigure 16), the yearly value of SPF is 3.84 dynamicbehaviour of thedualsource systemand theselectionof approximately, which is more similar to that obtained in winter working mode for 1 typical week in heating mode. Figure 17 shows season. Regarding the rest of the SPFs of the system, it can be con- the selection of working mode during 1 typical week in autumn. It cluded that SPF and SPF get values which are very close to SPF is possibletosee that,asthe airand ground temperature change, 2 3 1 (only 5% lower in the yearly value) which means that the auxiliar- the heat pump will select one source or the other, selecting the ies consumption for the ground loop circulation pump and the most favourable, as it was already explained in Section 3.2 of the fan do not have a great impact in the energy consumption of the present paper. When the air temperature is higher than the ground system. Analogously, when considering the circulation pumps temperature, the heat pump will select the air as source (in heating existing in the DHW loop and the user loop (for heating and cool- mode as well as in DHW mode). Analogously, when the air tem- ing), SPF decreases slightly (around 2%), being the yearly per- perature becomes lower than the ground temperature (considering formance factor of the system around 3.58. the hysteresis in the control), the source will change from air to In general, it can be concluded that the SPFs obtained during ground. On the other hand, if the air temperature is lower than the the summer are quite higher than the ones obtained during the ground temperature and then, the air temperature becomes higher, winter due to the use of free-cooling and natural ventilation. This thesourcewillchangefromgroundtoair. International Journal of Low-Carbon Technologies 2018, 13, 161–176 173 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 J.M. Corberán et al. Figure 16. Dual source system operating modes analysis. Figure 17. Change of working mode during 1 week in autumn. free-cooling (M10—Free cooling) purposes, this would happen 4.5 Working mode analysis (dual source system) during a 7% of the time. The rest of the operating modes during Figure 18 presents the percentage of time during which the system the summer mainly correspond to DHW production (M6—DHW has been working in each operation mode over the year. It should Air) which uses the air as a source, during a 3% of the time. be pointed out that these percentages are extremely influenced by Regarding the winter mode, as it can be observed in Figure 18, the weather and the thermal demand of the demo site existing in the system is working approximately during a 60% of the time in the Netherlands where temperatures are really mild during the heating mode using the ground as a source (M5—Winter Ground). summer (June to September) taking values lower than 21°C dur- The rest of the time is mainly working in heating mode (34%) using ing most of the time. This is the reason why the thermal loads in the air as a source (M4—Winter Air). It can also be observed that summer are very small leading to a high percentage (89%) of the the system only works a 6% of the time for DHW production (3% time working in MS—midseason, which means that the heat of the time using the ground as a source (M9—DHW Ground), pump is off, and the small cooling demand or air renovation needs and another 3% using the air as a source (M7—DHW Air)). This are satisfied just by opening the windows (natural ventilation). In percentage is much lower than that of the heating mode oper- the case that the air is hotter than 22.5°C, then the system would ation, mainly due to the low DHW demand, which is practically work using the ground loop as a sink (M2—Summer Ground), zero during the night. On the other hand, as the system will be during a 1% of the time, which is very low and could lead to an working in heating mode extracting heat from the ground during annual thermal unbalance in the ground. When the temperature most part of the year, special attention should be paid to the sum- of the fluid coming back from the ground loop and entering the mer where the heat injected into the ground is very low. This heat pump is lower than 10°C, the system will use the ground for 174 International Journal of Low-Carbon Technologies 2018, 13, 161–176 Downloaded from https://academic.oup.com/ijlct/article-abstract/13/2/161/4917534 by Ed 'DeepDyve' Gillespie user on 21 June 2018 Dual source heat pump, a high efficiency and cost-effective alternative including circulation pumps are embedded in the same unit box, which only needs the connection of the heating/cooling water loop, the brine loop and the DHW loop. This way the heat pump can be considered as a ‘plug&play’ solution. The DSHP unit has been designed, built and fully tested at the laboratory located at the IUIIE premises, at Universitat Politècnica de València. Furthermore, in the framework of the Geot€ch pro- ject it will be very soon tested in several demosites in the project. The unit has turned to be fully reliable with a smooth simple and full automatic operation. The DSHP works with R32 refrigerant and includes a variable speed compressor which give full capabilit- ies for an efficient modulating operation. The unit has been fully tested at the laboratory with very accurate instrumentation and Figure 18. Working mode time ratio for 1 year of operation of the dual this has allowed the characterization of its performance with sim- source system: (a) Winter season and (b) Summer season. ple polynomials. The article includes a brief summary of its per- formance, including a small study about the influence of some could let the ground thermally recover during the summer, but it important operating parameters on the system performance. may not be enough in order to reach a thermal balance in the In order to assess the energy performance of the heat pump dur- ground, making it necessary to inject heat in the ground during ing 1 year of operation, an integrated system model has been devel- the summer. This highlights the need for developing energy opti- oped in TRNSYS. The assessment consisted of two different steps. mization and control operation strategies to avoid this situation The first step consisted of a comparison between a DSHP sys- and make the system work under its optimal operation point. By tem (air/ground) and a GSHP system (only ground) for heating, using the TRNSYS model developed in this research work, it will cooling and DHW production in an office building located in the be possible to analyse the energy performance of the system and Netherlands. It was concluded that the DSHP system would be its impact in the ground temperature evolution over the years. able to reach a similar efficiency than the GSHP system (yearly Therefore, the model developed can be an assisting tool for the SPF around 3.6) with half the length of BHEs. Therefore, the optimal design and operation of the system and also to assess the DSHP system would be a profitable option against a GSHP sys- system energy performance and its suitability for other European tem, as the initial investment could be significantly reduced (up to countries with higher cooling thermal loads and lower heating a 30%) with a similar energy efficiency. thermal loads than the analysed case. For instance, a previous Then, the second step consisted of an analysis of the DSHP sys- research work was carried out by the authors in Ref. [19], where tem operation and energy performance along 1 year. It was con- the feasibility of the DSHP system was assessed for an installation cluded that, for the type of weather considered (low cooling loads located in Valencia, Spain. It was concluded that this type of and high thermal loads) the system operates mainly in heating DSHP system is even more convenient for Mediterranean cli- mode using the ground as a source during a 60% of the time, while mates (mostly cooling dominated), not only leading to a reduc- it uses the air during a 34% of the time in winter season. In con- tion in the size of the ground source heat exchanger needed, but trast to this, in summer season, the system is switched off during also presenting a higher yearly SPF (SPF equal to 4.62). This is most of thetimedue to themildsummerexistinginthe mainly due to the higher efficiency of the DSHP prototype in Netherlands, and it only works during a low percentage of time cooling mode than in heating mode. Therefore, it can be con- (8%) using the ground as a sink (1% of mechanical cooling and cluded that the greater the cooling thermal demand of the loca- 7% of free-cooling). Regarding the system energy performance, tion, the higher the annual SPF . the SPFs factors according to the SEPEMO-build project definition were obtained for a 1 year operation period. The system presented a yearly performance factor SPF around 3.58, whereas during the summer it took values considerably higher (around 5.09) due to 5 CONCLUSIONS the use of free-cooling and natural ventilation. It was concluded This paperpresentsthe characteristicsand energy performanceof the need for developing key control strategies to optimize the sea- an innovative DSHP developed in the framework of a H2020 sonal energy performance of this type of system. European project with title Geot€ch. The heat pump is able to employ either the air or the brine coming from the ground as heat sources in winter and provide hot water for heating the building. FUNDING The unit is reversible, so it can also provide cooling during sum- mer using the air or brine as a sink. Besides, it provides DHW all The present work has been supported by the European Union along the year, and in summer conditions, it can use the condens- under the Horizon 2020 Framework Programme for European ing waste heat to produce DHW. The heat pump is an outdoor Research and Technological Development (2014–20) inside the unit, very similar to an air–water heat pump unit. 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Journal
International Journal of Low-Carbon Technologies – Oxford University Press
Published: Mar 2, 2018