ASHRAE OR-16-C045-2016 Energy Use of Ground-Source Heat Pumps for Various Load Temperatures.pdf

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1、Nicolas Hache is a graduate student from the Department of Mechanical Engineering, Polytechnique Montreal, Montreal, Canada. Guillaume Soudan is a undergraduate student from the Department of Mechanical Engineering, Polytechnique Mons, Mons, Belgium. Michel Bernier is a professor in the Department o

2、f Mechanical Engineering, Polytechnique Montreal, Montreal, Canada. Energy Use of Ground-Source Heat Pumps for Various Load Temperatures Nicolas Hache Guillaume Soudan Michel Bernier, PhD, PE ASHRAE member ABSTRACT This paper examines the impact of the secondary fluid temperatures at the evaporator

3、and condenser on the energy use of water-to-water ground-source heat pumps (GSHP). In the first part of the paper, the energy consumption reduction associated with small temperature differences between source and load temperatures is evaluated from a thermodynamic point-of-view by examining the coef

4、ficient of performance (COP) of an ideal refrigeration cycle. Then, the performance map of a typical water-to-water heat pump is examined to determine real COPs for a range of source/load temperatures and flow rates. In the second part of the paper, annual simulations are performed on a ground-sourc

5、e heat pump system providing space heating and domestic hot water (DHW) for a well-insulated single-family house. Two different load temperatures and two different source and load flow rates are examined for a total of eight cases. The concept of seasonal performance factors (SPF) is used to account

6、 for all the energy flows into the system including pumping energy. Results show that the highest value of SPF4 (2.44) is obtained when the source and load flow rates are 9.0 gpm (0.56 l/s) and 4.5 gpm (0.28 l/s) respectively, and the return load temperature is 40 C (104 F). There is a difference of

7、 8% between the lowest and highest values of SPF4 for the eight cases studied here indicating that the choice of the source and load flow rates as well as the load temperature is relatively important to limit the energy use of GSHP. INTRODUCTION Classic thermodynamics tells us that the energy requir

8、ed in a compression heat pump decreases with a reduction of the difference between the condensing and evaporating temperatures of the refrigerant. This can be achieved with efficient heat transfer at the evaporator and condenser and by lowering the difference between the source and load temperatures

9、 of the secondary fluids. Recent studies (e.g. Girard et al. 2015, Sarbu and Sebarchievici 2015, and Maivel and Kurnitski 2015) have examined the effects of temperature on heat pump performance. In this study, the impact of the secondary fluid temperatures at the evaporator and condenser on the ener

10、gy use of water-to-water ground-source heat pumps (GSHP) is examined. In addition, various source and load flow rates of the secondary fluids are examined as they modify heat transfer and refrigerant temperatures in the evaporator and condenser. This paper is organized as follows. First, the coeffic

11、ient of performance (COP) for an ideal heat pump is presented. Then, COPs of a commercially available water-to-water heat pump are reviewed and compared to the COP of an ideal heat pump. Finally, seasonal performance factors are calculated based on annual simulations of a GSHP system used for space

12、heating and domestic hot water (DHW) heating of a typical house located in a northern climate. COP The four basic components of a heat pump are shown in Figure 1a. The ideal cycle involves an isentropic compression (1-2), a condensation at constant pressure (2-3), an isenthalpic pressure reduction (

13、3-4) and an evaporation at constant pressure (4-1). There is a finite temperature difference between the secondary fluid temperatures, Tsource and Tload , and the refrigerant evaporating and condensing temperatures, Tevap and Tcond . Hence, the refrigerant temperature in the evaporator is a few degr

14、ees lower than the temperature of the secondary fluid on the source side. Similarly, the refrigerant temperature in the condenser is a few degrees higher than the temperature of the secondary fluid on the load side. Com p r e s s o rE x pa ns i onv a lv eE v a p or a t o rCon d e n s e r1234Q EQ CW

15、a) b) Figure 1 (a) Schematic representation of a compression heat pump and (b) Pressure-enthalpy diagram for two sets of conditions. (1000 kPa = 145 psia ; 300 kJ/kg = 129 Btu/lbm) Figure 1b shows the P-h diagram for two pairs of (Tsource/Tload), i.e. 30/120 F (-1.1/48.9 C) and 60/60 F (15.6/15.6 C)

16、. The refrigerant used is R-410A and temperature differences (Tsource - Tevap) and (Tcond - Tload) of 5 K (9 F) are assumed. The COPideal values calculated for these conditions are given in Table 1. This table also includes data from a commercially available water-to-water heat pump (used later in t

17、his study). The two shaded columns identified by the letter A refer to COPideal. As shown in this table, COPideal decreases as the difference between the load and source temperatures increases. The lowest (3.75) and the highest (25.84) values of COPideal are obtained for source and load temperatures

18、 equal to 30/120 F (-1.1/48.9 C) and 60/60 F (15.6/15.6 C), respectively. This relatively large difference between the lowest and highest values of COPideal can be explained using the P-h diagram shown in Figure 1b. When the difference between the evaporating and condensing temperatures of the refri

19、gerant is small, the heating effect (2 3) is relatively large and the compression work (2 1) is relatively small which leads to high values of COPideal. In practice, heat pumps do not operate ideally mainly because of compressor inefficiencies. Furthermore, isentropic efficiencies of compressors are

20、 not constant but vary as a function of the pressure ratio between the condensing and evaporating temperatures. Thus, the COP of a heat pump will not only vary because of the difference between the condensing and evaporating temperatures but also because the compressor efficiency changes with this t

21、emperature difference. Actual COP values of a commercially available water-to-water heat pump are shown in columns B in Table 1 for four source and four load temperatures as well as two source flow rates and two load flow rates ( and in Figure 1a). As shown in this table, COPs follow the same trend

22、as the values obtained earlier for COPideal , i.e. they decrease with an increase of the temperature difference (Tload -Tsource ). The lowest COP value is 2.20 and the highest is 8.60. High load and source flow rates increase heat transfer coefficients in the evaporator and condenser which reduces t

23、he temperature difference (Tcond - Tload) and (Tsource -Tevap) and increases the value of the COP. For example, for a source temperature of -1.1 C (30 F) and a load temperature of 48.9 C (120 F), the COP increases from a value of 2.20 for source and load flow rates both equal to 0.28 L/s (4.5 gpm) t

24、o a value of 2.50 when both the load and source flow rates are doubled. Two additional columns representing the ratios of the real COP to COPideal have been added to the right of the B columns. This ratio varies from 0.28 to 0.72 with an average of 0.59. Thus, on average, the COP of this heat pump i

25、s 59% of COPideal. Table 1: Ideal COPs and Performance Data of a Water-to-Water Heat Pump In summary, the COP of a heat pump can be increased by lowering the difference between the secondary fluid temperatures (Tload - Tsource) and by increasing the flow rates of the secondary fluids in both the eva

26、porator and condenser. For a GSHP system operating in heating mode, this implies that the ground heat exchanger length should be as long as possible so that the return temperature to the heat pump is as high as possible. In addition, the load temperature should be as low as possible. Thus, it might

27、be advantageous to couple the heat pump to a low temperature radiant floor instead of using radiators or fan-coils that may require higher operating temperatures. Both source and load flow rates should be as high as possible. The resulting increase in the value of the COP can, however, be counterbal

28、anced by increased pumping energy caused by an increase of the pressure drop in the evaporator and condenser and in the rest of the distribution circuits on the source and load sides. The objective of this study is to quantify this impact by determining the energy use of a GSHP system for two load t

29、emperatures and two source and load flow rates to determine the best operating scenario. METHODOLOGY System Description The system under study is shown in Figure 2. It consists of a ground-source heat pump providing space heating and domestic hot water (DHW) for a well-insulated single-family house.

30、 The energy performance of the system is assessed by running annual simulations using readily available models in TRNSYS v17 (Klein et al. 2010). A time step of 0.1h is used. E a u x 2E a u x 1Q a u x 2T DH WHeat P umpE P 1E P 3E P 2E HPS P F 1S P F 2B uf f er t a nkS P F 4S P F 3D HW t a nkQ a u x

31、1T B - t a n kM a i n s w a tertemp er a t u r eT inm l o a d.m s o u rc e.P ump 1P ump 2P ump 3T = 60 CT o u t , DH WT o u t , b u ff e rT o u t , HPQ H PFigure 2 Schematic representation of the system under study. A 10 kW (3 tons) water-to-water heat pump is connected on the load side to a buffer

32、tank for space heating and to a DHW tank. A 140 m (460 ft.) deep borehole is connected on the source side. A total of three circulating pumps are used (P1, P2, and P3 in Figure 2). A three-way valve is used to divert the flow from one tank to the other with priority given to the buffer tank for spac

33、e heating. The set-point temperature for the DHW tank is 45 C (113 F). Thus, when the return temperature from the DHW tank, TDHW, falls below 45 C (113 F) then the heat pump is started along with pumps P1 and P2. Typically, when fully charged, the top temperature in the DHW tank is around 55 C (131

34、F) and a small amount of auxiliary heat is required to reach the desired temperature of 60 C (140 F). Space air temperature in the house is controlled with a two-stage thermostat. When the air temperature drops below 21 C (69.8 F), pump P3 is activated. If the air temperature continues to drop and r

35、eaches 20 C (68 F) then an auxiliary heater is energized to supplement the heat from the buffer tank. If pump P3 operates then the temperature of the buffer tank will decrease. If the temperature in the bottom of the buffer tank, TB-tank, reaches a certain set-point then the heat pump as well as pum

36、ps P1 and P2 are activated. Two set point temperatures for TB-tank are examined here: 40 C (104 F) and 30 C (86 F). The first one may be representative of a system using a fan-coil or radiators while the lower temperature might be associated with a radiant floor heating system. These set point tempe

37、ratures are constant during the year. House and DHW Loads The 220 m2 (2350 ft2) house is modeled using TYPE56 in TRNSYS. It is located in a relatively cold climate (Montreal, Canada). The hourly heating load is given in Figure 3a. Peak demand for space heating is approximately 8.7 kW (29.7 kBTU/hr)

38、and the annual space heating requirement is 20800 kWh ( 71 MBTU). As shown in Figure 2, space heating is achieved using a generic heat exchanger. A constant efficiency of 0.9 is assumed in the modeling of this heat exchanger. The daily hot water consumption is 210 liters (55 gallons). The daily draw

39、 profile used is the one recommended by Hendron (2008) and it is shown in Figure 3b. The mains water temperature is obtained from the weather processor (TYPE15 in TRNSYS). It varies from 3 C (37.4 F) to 15 C (59 F) during the year. The annual energy consumption for DHW is approximately 5000 kWh (17.

40、1 MBTU). Figure 3 (a) Space heating load and (b) Daily domestic hot water demand. Models A borehole model based on the thermal Resistance and Capacitance (RC) approach (Godefroy and Bernier 2014) is used to model the single U-tube borehole. The main characteristics of the borehole are given in Table

41、 2. Table 2: Main Characteristics of the Borehole Parameter Value Unit Value Unit Depth 140 m 460 ft Borehole radius 0.075 m 3 in Outside pipe radius 0.016 m 0.63 in Borehole thermal resistance 0.1 m.K/W 0.173 h.ft.F/Btu Grout conductivity 0.83 W/m.K 0.48 Btu/h.ft.F Ground thermal conductivity 2.2 W

42、/m.K 1.27 Btu/h.ft.F Pipe conductivity 0.42 W/m.K 0.24 Btu/h.ft.F Propylene glycol concentration 25 % 25 % The heat pump is modeled in TRNSYS as a reversible water-to-water single-stage heat pump (TYPE927). The data presented earlier (see Table 1) is used to represent the performance over the range

43、of source and load temperatures and for the two source and load flow rates. Table 3: Characteristics of the Circulators Parameter Value (low, high) flow rates Unit Value (low, high) flow rates Unit P1 Source flow rate 0.28, 0.56 l/s 4.5, 9.0 gpm Pressure drop in borehole 56, 160 kPa 13.4, 53.5 ft Pr

44、essure drop in heat pump 12, 48 kPa 4, 16 ft Pump efficiency 15, 30 % 15, 30 % P2 Load flow rate 0.28, 0.56 l/s 4.5, 9.0 gpm Pressure drop in distribution circuit 8, 32 kPa 2.7, 10.7 ft Pressure drop in heat pump 8, 32 kPa 2.7, 10.7 ft Pump efficiency 10, 20 % 10, 20 % P3 Flow rate 0.14, 0.28 l/s 2.

45、25, 4.5 gpm Pressure drop in distribution circuit 12.5, 50 kPa 4.2 ,16.7 ft Pump efficiency 10, 15 % 10, 15 % 00.010.020.030.040.050.060.0702468101214160 4 8 12 16 20 24DHWflowrate UsgpmDHWflow ratel/hTime hCirculators are modeled using TYPE741 in TRNSYS. This model calculates the circulator energy

46、consumption for given values of the flow rate, pressure drop, and circulator efficiency. The values of pressure drops and circulator efficiencies assumed here are presented in Table 3. Efficiency ranges from 10 to 30% and depends on the size of the circulator. These values were obtained from a study

47、 performed in Europe a dozen years ago (Costic 2003) and may not represent current values in North-America. Therefore, as shown later, these values were doubled to examine the impact of circulator efficiency on the overall results. As noted in Table 3, two flow rates are examined for each pump: 0.28

48、 l/s (4.5 gpm) and 0.56 l/s (9.0 gpm) for P1 and P2; and 0.28 l/s (4.5 gpm) and 0.14 l/s (2.25 gpm) for P3. For P1 and P2, the flow rates correspond to those given in the performance map of the heat pump presented in Table 1. For P3, the two flow rates correspond to cases where TB-tank are 30 C (86

49、F) and 40 C (104 F), respectively. Both tanks are modeled using TYPE534 in TRNSYS. The volume of the DHW and buffer tanks are 350 liters (92 gallons) and 2000 liters (530 gallons), respectively. Both tanks are insulated and the heat loss coefficient is 0.83 W/m.K (0.15 Btu/h.ft.F). RESULTS A total of eight cases are compared in this section. They are identified by three numbers (a/b/c) where